<2s ij > ---------------- + --------------------- = – -------------- + ------------------------ xi t x j x j
(16)
Based on Equation (14), the
(17)
The last three terms that contain the subgrid velocity are therefore the subject of modeling (Ferziger 1977). Breuer (1998) and Spalart (2000) describe some of the many other subgrid models and their performance. The latest developments of LES and its related
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techniques, such as the detached eddy simulation (DES), are described in detail by Spalart (2000).
Direction Numerical Simulation (DNS) Direct numerical simulation (DNS) is used to study turbulent flow. This method is very accurate (sometimes better than experiments), and is used to benchmark performance of other CFD techniques. Because of its stringent requirements on grid, especially in the normal direction within the boundary layer (Grötzbach 1983), DNS is used to study spatially and temporally confined flows with simple geometry (Spalart 2000). Notwithstanding these limits, DNS also has been used to explore more complicated geometry, such as flow over a wavy wall (Cherukat et al. 1998), and flow mechanisms, such as multiphase flow (Ling et al. 1998) and droplet evaporation (Mashayek 1998).
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1.1
MESHING FOR COMPUTATIONAL FLUID DYNAMICS
The first step in conducting a CFD analysis for a fluid region of interest is to divide the region into a large number of smaller regions called cells. The collection of cells that makes up the domain of interest is typically called the mesh or grid, and the process of dividing up the domain is called meshing, gridding, grid generation, or discretization of the computational domain. Meshes can be structured or unstructured, depending on the connectivity of the cells in the mesh to one another. Individual cell shape varies, and each shape has advantages and disadvantages. These shapes range from triangles and quadrilaterals for twodimensional (2D) geometry, to tetrahedrals (four-sided triangularbased shapes) and hexahedrons (typically six-sided boxes) for three-dimensional (3D) geometry. Wedges (a triangle swept into a three-dimensional shape) and rectangular-based pyramids can also be used to transition between the triangular sides of the tetrahedrals and quadrilateral sides of the hexahedrons.
Structured Grids Structured grids have consistent geometrical regularity, wherein families of grid lines (in one direction) do not cross each other. Figure 2 shows examples of structured grids. These grids can be further subclassified as orthogonal and nonorthogonal. Orthogonal structured grids, the simplest scheme, are based on Cartesian/polar-cylindrical coordinate systems. A curved or sloped boundary in the CFD domain is typically approximated by stepwise boundary. Figure 2A shows a meshed 2D domain for flow through a 90° elbow using a Cartesian orthogonal coordinate system. Cells outside the elbow are blocked from CFD analysis or turned into cells that do not participate in the flow field. The stairstep approach to representing the curved surface can lead to numerical errors at curved
Fig. 2 Two-Dimensional CFD Structured Grid Model for Flow Through 90° Elbow
walls. Finer grids are needed to more accurately represent the curved/sloped boundary. The effect of reducing local grid size in one region (grid refinement) may propagate to other sections of the domain and result in an increase in the number of model cells. This, together with the blocked cells outside the flow domain, creates a burden on computing resources. The stepwise approximation of the boundary may also result in errors that negatively affect the CFD solution. Modeling curved/sloped surfaces is possible by using the geometrical flexibility of the nonorthogonal grid, also known as bodyfitted or boundary-fitted grid. An example of a 2D body-fitted nonorthogonal structured grid for a 90° elbow is shown in Figure 2B. Using the body-fitting method, geometric details are accurately represented without using stepwise approximation. An orthogonal grid can be structured (i.e., a single block, as in Figure 2), blockstructured, or overlapping-structured. A block-structured grid consists of a group of meshed regions (blocks) that collectively form the entire region of interest. This is typically referred to as a multiblock domain. The blocks may have a fine grid at the region of interest, to provide more details for flow field analysis, and a coarser grid away from the region of interest. Figure 3 shows block-structured grid for 2D flow through a 90° elbow connected to a rectangular duct. The grid is fine close to the solid surfaces, and is refined at point A, where flow separation is expected. This grid refinement is propagated through blocks 2 and 3. Interblock interfaces could have matching grids, as between blocks 1 and 2, or a nonmatching interface, as between blocks 2 and 3. The nonmatching interface is used to transfer from coarse to finer grid or vice versa. Numerical inaccuracies can occur where blocks are joined together with nonmatching mesh lines. The relative difference in mesh size on either side of the interface is important. Also, there is additional computational overhead associated with managing the nonconformal interface. At the interface of the block-structured grid, the ratio of cell size change (i.e., large to small cells) between two blocks is recommended to be no more than two (Ferziger and Peric 1997), because transporting field variables from a fine to a coarse mesh or vice versa allows inaccuracies to enter the solution. If flow in a domain travels from a group of four cells to a single cell, the flow detail represented by the four cells is lost. In some cases, this rule can be bent, but this is best done by an experienced CFD modeler. Structured grids simplify programming for the CFD code developer and provide regular structure for the matrix of algebraic equations. However, they may not adequately describe complex geometries, and it can be difficult to control grid distribution in the region of interest without propagating through the whole analyzed domain.
Fig. 3 Block-Structured Grid for Two-Dimensional Flow Simulation Through 90° Elbow Connected to Rectangular Duct
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Indoor Environmental Modeling These structured grid types are mainly associated with finitedifference methods. The examples in Figures 2 and 3 are called the physical planes. Finite-difference methods require a uniform rectangular grid called the computational plane. The governing equations must be transformed to give one-to-one correspondences between the physical and computational planes. After analysis of the computational plane, the results are transferred back to the corresponding point on the physical plane. Using data transformation increases the programming efforts and computing costs for CFD. Anderson (1995) has more information on transformation methods.
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Unstructured Grids Unstructured grids (Figure 4) are flexible: they can represent complex geometry boundaries, and can be easily refined in the region of interest without propagating to the rest of the domain. Elements of different shapes can be used in the domain. Either matching or nonmatching nodes can be used between neighboring elements. Figure 4A is an unstructured grid using tetrahedral elements, whereas Figure 4B uses hexahedral elements; note that both have a meshing zone near the pipe wall to resolve the boundary layer. Unlike structured grids, the matrix of algebraic equation does not have a regular diagonal structure and has a slower solver than a structured grid solver (Ferzinger and Peric 1997). Unstructured grids are mostly used for finite-element and finitevolume methods. No transformations are required for finite-volume methods, and analysis can be performed directly on the physical plane of the unstructured grid.
Grid Quality Grid quality measures include the shape of the individual cells, the size of the cell relative to flow field features of interest, and the jump in grid size from one block to the next. Cell quality is important. Values for variables stored at centers of cells must be interpolated to the face of the cell, which allows calculation of fluxes at faces of control volumes. A poor-quality mesh gives less accurate interpolations and can affect the quality of the simulation result by the introducing numerical inaccuracies. Mildly poor grids can increase convergence times; in extreme cases, local poor cell quality can result in overall flow field inaccuracies or cause the simulation to diverge and not reach a solution at all. The examples in Figures 2 to 4 show clean, nonskewed cell shapes: the triangles do not lean over and the quadrilaterals have corners that do not vary significantly from 90°. In many practical meshes, the individual cells become distorted from these ideal shapes (e.g., because four-sided shapes may not fit well into wedgeshaped corners). The amount of distortion is typically referred to as skewness. Different CFD codes allow different levels of skewness, and the solver’s overall sensitivity to skewness may be affected by the method of numerical discretization.
13.5 Grid cell size may be partially determined by the level of geometric detail needed. Cells need to be small enough to resolve the smallest geometric features of the domain. If cell size is too large, any curved elements cannot be represented by a series of straight lines. In addition, a CFD solution’s ability to resolve flow field features is limited by the grid resolution. If a grid has cell sizes of 10 mm, then flow field features smaller than 20 mm cannot be solved. Therefore, sharp gradients of flow field variables necessitate a finer mesh. The additional cells in the finer mesh are required to resolve the rapid change in the flow field variable. Additionally, fine grid resolution at walls is important for methods that use the law of the wall (see the Wall/Surface Boundary Conditions section), because fluxes and shear at the wall require accurate calculations of gradients. Many CFD codes can start with a coarse mesh and add cells as necessary in particular regions, which can save computational time in the set-up and test simulation phases. When generating a mesh, it is important to determine whether the mesh itself will affect the simulation and generate erroneous results. Figure 5 shows three circles with different meshing schemes. Figure 5A shows a grid on the circle that has a pincushion-like look; this mesh distorts the flow field in the “corners,” because the four corner cells have two sides on the perimeter of the circle, whereas the rest of the cells around the circumference have only one. If this mesh is used to simulate flow down a pipe, flow velocity in the corner cells is adversely affected by the additional friction that these cells experience. Figure 5B shows the same geometry, but with a structured mesh created by placing a square in the center of the circle and drawing rays diagonally out from the corners. These rays and the circle’s perimeter define other meshing blocks. If cut halfway through the horizontal or vertical centerline, the mesh can be stretched out into a rectangle. Finally, Figure 5C shows an unstructured mesh over the same geometry. The meshes in Figures 5B and 5C influence results less significantly than that in Figure 5A.
Immersed Boundary Grid Generation The preceding discussion assumes that a 2D or 3D computer model already exists for the geometry, architecture, or region of interest, and that the geometry can be imported into a gridding software package. The grid is then overlaid onto that geometry. In immersed boundary grid generation, a model is not required to exist a priori. This simplifies gridding and geometry generation in one step: the orthogonal structured grid is created within a block and then the geometry (represented by blanked-out sections of cells within the meshed block) is overlaid (see Figure 1A). Groups of meshed blocks with overlaid architecture can be assembled to generate more complicated computational domains. Although this technique can yield high-quality meshes for simple geometrical features (blocked representation), more sophisticated grid-generation techniques may be needed to grid complicated domains.
Grid Independence The level of grid independence from the flow field solution is important to determine in advance. It may be sufficient to demonstrate
Fig. 4 Unstructured Grid for Two-Dimensional Meshing Scheme Flow Simulation Through 90° Elbow Connected to Rectangular Duct
Fig. 5 Circle Meshing
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that grids of similar resolution applied to similar problems give sufficiently low levels of uncertainty. Grid independence can be achieved experimentally by using successive grid refinements in areas with sharp gradients or cell skewness. This allows solutions obtained with coarser and finer grids to be compared. If results of two successive trials are comparable, then both models are grid-independent. In some cases, experienced CFD practitioners may be able to identify a sufficiently grid-independent solution without trials, but new CFD users should not assume a solution is grid independent: different levels of grid can yield results different enough to make conclusions drawn from the flow field unreliable.
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1.2 BOUNDARY CONDITIONS FOR COMPUTATIONAL FLUID DYNAMICS Boundary conditions are integral to CFD modeling’s ability to solve the general Navier-Stokes equations for a particular problem in an indoor environment. Boundary conditions specify physical and/or chemical characteristics at the model’s perimeters. These characteristics could be constant throughout the analysis or timedependent in transient analyses. This section discusses applying boundary conditions for CFD modeling of subsonic fluid flow. Excluding free surface flow, most boundary conditions can be classified as either Dirichlet (variable values specified at the boundary node) or Neumann (variable derivatives are required at the boundary, and the boundary condition must be discretized to provide the required equation). Free surface flow requires moving boundary conditions, such as kinematic and dynamic. Every model has walls, and most have at least one inlet and one outlet boundary. Some cases of natural convection heat transfer (e.g., CFD modeling of an enclosure with heated and cooled sides, modeling of convection from an object such as cylinder in a large fluid medium) may not require an inlet or outlet in the model. Typical boundary condition types for HVAC applications are inlets, outlets, walls or surfaces, symmetry surfaces, and fixed sources or sinks.
Inlet Boundary Conditions Special attention must be paid to inlet boundary conditions, because supply diffusers are usually dominant sources of momentum that create airflow patterns responsible for temperature and concentration distributions. Inlet boundary conditions may be velocity, pressure, or mass flow. When details of flow distribution are unknown, a constantpressure boundary or constant flow rate can be specified. The pressure inlet boundary requires specifying static pressure for incompressible flow. Stagnation pressure and stagnation temperature should be specified at the pressure boundary for compressible flow. In both conditions, velocity components at the boundary are obtained by extrapolation. The mass flow inlet condition requires specifying the mass flow rate and temperature for incompressible flow at the boundary. Velocities and pressure are calculated by extrapolation at inlet boundaries. Experimentally measured values of turbulence quantities at the inlet boundary are also required for accurate CFD simulation for turbulent flow. For the k- turbulence model, turbulent kinetic energy k and turbulent dissipation rate are required. When these values are not available from experimental data, they can be estimated from the following equations: 2 3 k = --- U ref TI 2 3 4
= C
(20)
where Uref is the mean stream velocity, TI is turbulence intensity, l is the turbulence length scale, C is the k- turbulence model constant (C = 0.0845), and L is the characteristic length of the inlet (for a duct, L is the equivalent radius). For indoor environmental modeling, the inlet boundary is especially important because of the potentially complex geometry of supply diffusers designed to produce particular performance characteristics. Detailed diffuser modeling is possible for limited regions near the diffuser, but is not very practical in room-flow simulations: including the small geometric details of the diffuser in the room model results in a mesh with so many cells that current computation resources cannot efficiently find a solution. Therefore, most room airflow simulations should use simplified diffuser modeling that replicates diffuser performance without explicitly modeling the fine geometric details of the diffuser. Other simplifications are possible. The most obvious method is replacing the actual diffuser with a less-complicated diffuser geometry, such as a slot opening, that supplies the same flow momentum and airflow rate to the space as the actual diffuser does (Nielsen 1992; Srebric and Chen 2002). Simplified methods are classified as jet momentum modeling either (1) at air supply devices or (2) in front of air supply devices (Fan 1995). Modeling at the supply device has several variations, including the slot and momentum models; variations of modeling in front of the diffuser include prescribed velocity, box, and diffuser specification (Srebric and Chen 2001, 2002). The most widely used methods are the momentum, box, and prescribed velocity methods. The momentum method decouples the momentum and mass boundary conditions for the diffuser (Chen and Moser 1991). The diffuser is represented with an opening that has the same gross area, mass flux, and momentum flux as a real diffuser. This model allows source terms in the conservation equations to be specified over the real diffuser area. Air supply velocity for the momentum source is calculated from the mass flow rate m· and the diffuser effective area A0: m· U 0 = --------A 0
(21)
The momentum method is very simple, but might not work well for certain types of diffusers (Srebric and Chen 2002). The box method is based on the wall jet flow generated close to the diffuser (Nielsen 1992; Srebric and Chen 2002). Figure 6 shows the location of boundary conditions around the diffuser. Details of flow immediately around the supply opening are ignored, and the supplied jet is described by values along surfaces a and b. There are two advantages of this method compared to the detailed diffuser simulations: (1) the box method does not require as fine a grid as fully numerical prediction of the wall jet development; and (2) twodimensional predictions can be made for three-dimensional supply openings, provided that the jets develop into a two-dimensional wall
(18)
32
k ---------l
l = 0.07L
(19)
Fig. 6 Boundary Condition Locations Around Diffuser Used in Box Method
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Indoor Environmental Modeling
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jet or free jet at a certain distance from the openings. Data for velocity distribution in a wall (or free) jet generated by different commercial diffusers can be obtained from diffuser catalogues or design guide books, Chapter 20, and textbooks [e.g., Awbi (1991), Etheridge and Sandberg (1996), and Rajaratnam (1976)]. The prescribed velocity method has also been used in numerical prediction of room air movement. Figure 7 shows the method’s details. Inlet profiles are given as boundary conditions of a simplified slot diffuser, represented by only a few grid points. All variables except velocities u and w are predicted in a volume close to the diffuser (xa, yb) as well as in the rest of the room. Velocities u and w are prescribed in the volume in front of the diffuser as the fixed analytical values obtained for a wall jet from the diffuser, or they are given as measured values in front of the diffuser (Gosman et al. 1980; Nielsen 1992). For a more detailed description of simplified methods and their applicability to common supply diffusers, see Chen and Srebric (2000). Figure 8 shows how real diffusers can be simplified by the momentum method and box method in CFD simulations (Srebric 2000).
13.7 Outlet Boundary Conditions A mass flow rate or constant pressure can be specified for an outlet boundary condition. The outlet flow rate or pressure boundary is extrapolated to determine the boundary velocity, which needs to be corrected during calculations to satisfy mass conservation in the analyzed domain. Some commercial CFD codes require turbulence values at outlet boundary conditions. These values are used when reversed flow occurs at the outlet pressure boundary.
Wall/Surface Boundary Conditions Wall boundary conditions represent the wet, solid perimeter of the CFD model. All velocity components are set equal to wall velocity for no-slip wall conditions, and to zero for a stationary wall. This is an example of a Dirichlet boundary condition. Wall roughness for both types of flow regimes (laminar and turbulent) should be specified. At the wall, both regimes have laminar flow. For turbulent flow, the wall turbulent boundary layer consists of three sublayers as presented in Figure 9 (Wilcox 1998): a thin, viscous sublayer followed by the log-law layer and the defect layer. Turbulent flow modeling requires very fine mesh inside the boundary layer. This requires extensive computational hardware, which is very costly, but the resource requirements can be reduced by using empirical wall functions in the near wall region instead of directly applying the k- turbulence model with a very fine mesh. Turbulent flow near a wall can be categorized as laminar (viscous sublayer) or turbulent (log-law layer), depending on the dimensionless distance Y + from the wall, defined as Yp w Y + = --------- ----
Fig. 7
Prescribed Velocity Field Near Supply Opening
(22)
where Yp is the distance from the wall to the center of the first cell, w is wall shear stress, is fluid density, and is the fluid kinematic viscosity. The viscous sublayer is very thin (Y + < 5), and is typically smaller than the first cell (2Yp ). For the viscous sublayer, as shown in Figure 9, u+ = Y +
(23)
where u+ is the dimensionless mean velocity (u+ = UP /ut , ut = w , and UP is the is the velocity parallel to the wall at Yp ). In practical terms, most important layer is the turbulent log-law sublayer, which is characterized by the following dimensionless velocity profile:
Fig. 8 Simplified Boundary Conditions for Supply Diffuser Modeling for Square Diffuser
Fig. 9
Typical Velocity Distribution in Near-Wall Region
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2017 ASHRAE Handbook—Fundamentals (SI) 1 u + = --- ln EY +
(24)
where is Von Karman’s constant ( = 0.41), and E is a constant depending on the wall roughness (E = 9.8 for hydraulically smooth walls). The size of the log-law layer is typically 30 < Y + < 500 (Versteeg and Malalasekera 1995). The turbulent kinetic energy k and turbulent dissipation rate use the following functions in the log-law layer: 2
3
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u u k = ------------ and = -----y C
(25)
where C is the k- turbulence model constant. Overall, wall functions are of great practical importance because they allow significant savings of computational time. However, assumptions used to derive the wall functions [i.e., Prandtl mixing hypothesis, Boussinesq eddy viscosity assumption, fully developed flow, and no pressure gradients or other momentum sources (constant shear stresses)] restrict their application to a certain class of flows. For indoor airflow applications, these assumptions are acceptable, and wall functions are widely used. However, predicted heat transfer in the near-wall region tends to be incorrect, depending on the control volume size at the wall (Yuan et al. 1994). Heat transfer calculation can be improved with more accurate temperature profile equations or use of prescribed empirical values for the convective heat transfer coefficient h. Surface temperature and heat transfer are often complicated variables of time and position. However, many CFD simulations use steady-state boundary conditions for a typical or design day. Boundary conditions for surface temperature and heat transfer are illustrated in Figure 10, showing how surface temperature Ts depends on heat transfer to and from the surroundings, on radiation to and from the surfaces in the room, and on the air temperature close to the surface. Boundary conditions for temperature or energy flux can be found from measurements, manual energy calculations, or a building energy performance simulation (BEPS) program. BEPS predicts both energy flow in the building structure and radiation plus detailed dynamic energy flow and consumption of the whole building during a period of time (Figure 11). There are different ways to exchange heat transfer information between BEPS and CFD programs (Zhai et al. 2002); the best method is to transfer surface temperatures from
Fig. 10 Wall Surface Temperature Ts , Influenced by Conduction Tw , Radiation Trad , and Local Air Temperature TP
BEPS to CFD, and convective heat transfer coefficients and air temperature from CFD to BEPS, to achieve a unique solution (Zhai and Chen 2003). Dynamic simulations can be structured in different ways. A BEPS program can be connected to a separate CFD program, which predicts energy flow in selected situations. A CFD program can also be extended to find a combined solution of radiation, conduction, and thermal storage parallel to solving the flow field; this is often called a conjugate heat transfer model. Another possibility is to use additional CFD code in selected rooms as an extension of a large BEPS program. Examples of conjugate heat transfer and combined models are available in Beausoleil-Morrison (2000), Chen (1988), Kato et al. (1995), Moser et al. (1995), Nielsen and Tryggvason (1998), Srebric (2000), and Zhai and Chen (2003). The simplest way to account for heat transfer at CFD boundaries is to prescribe wall temperatures obtained from on-site measurements. Using a turbulence model without wall functions is also recommended when heat and mass flows from surfaces are the important parameters. Predictions of actual flow at surfaces are more accurate than analytical values found from wall functions.
Symmetry Surface Boundary Conditions For a model with symmetrical flow in at least one plane, the symmetry boundary condition represents no flow across the symmetry plane, and all scalar fluxes are set to zero. Select symmetry boundary conditions cautiously. Although the geometry of the model has symmetry, fluid flow might not be
Fig. 11 Combination CFD and BEPS The CFD program predicts flow in room based on heat flux calculated by BEPS program.
Fig. 12
Duct with Symmetry Geometry
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Indoor Environmental Modeling symmetrical at the geometry symmetry plane. In Figure 12, the geometry is symmetrical at plane A-A, but the flow is asymmetrical because of flow instability at the merged flow region. Zhang et al. (2000) studied the symmetry pattern for a an 8.5 by 5 by 3 m room, with air supplied through a slot at the ceiling and along the 5 m long side. Return air exited through a slot at the floor along the same side wall as the inlet air. Velocity on one side of the symmetry plane was up to two times higher than on the other side, at the symmetrical location.
Fixed Sources and Sinks
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Fixed sources are boundary conditions specified as fixed values of the calculated parameter or as mass/momentum/heat/contaminant fluxes. Examples include heat flux from a wall to simulate solar radiation, total heat flow in the occupied zone to simulate heat dissipation from occupants, momentum source from an operating fan, and species generation rate from contaminant sources. These sources could be placed anywhere in the CFD calculation domain, and can vary with time. Fixed-value sources/sinks usually use the wall functions if the source/sink is given as a fixed value, whereas fixed-flux sources/sinks do not use wall functions. These boundary conditions could also be associated with blockages in the flow domain (e.g., furniture, occupants, or other flow obstacles).
Modeling Considerations Pressure boundary conditions are used when there is not enough information to specify the flow distribution, and the boundary pressure is known or assumed. Pressure boundaries are mostly used for buoyancy-driven flow and external flow applications. In inlet/outlet pressure boundary conditions, the stagnation pressure, temperature, concentration, and turbulence quantities k and are required. For external flow away from the wall, the free-stream k and can be set to zero. A pressure boundary velocity cannot be specified, but must be determined by the CFD code from interior conditions relative to the pressure boundary condition (inflow or outflow condition is possible). When using different boundary conditions, be aware that most boundary conditions are just approximations of real physical phenomena. It is the user’s responsibility to evaluate adequacy and influence of different boundary conditions on the accuracy of CFD simulation solutions.
1.3
CFD MODELING APPROACHES
Some steps are common to developing all types of CFD models and can increase the likelihood of getting a reasonable result with appropriate computing time.
Planning Planning a CFD simulation is perhaps the most important step. During this phase, a clear understanding of what is being investigated is important. If the simulation is about thermal comfort in a room, some details about unoccupied space regions may be simplified: there may be little point in determining the thermal comfort in an unoccupied zone. If the purpose is to gain insight into a flow field, then some effort to estimate flow patterns helps during early modeling decisions and later evaluation of results. During planning, a decision of whether to conduct a steady-state or transient simulation must be made. Transient simulations present time-accurate results, such as filling a tank, whereas steady-state simulations represent conditions after the flow field has been flowing long enough to reach equilibrium. It is important to recognize that some flows are inherently unsteady. The effect of choosing a steady-state solver to model unsteady flow should be considered during this process.
13.9 The physics to be examined should also be determined. Turbulence, heat transfer, species transport, and radiation phenomena can be evaluated during CFD modeling. The final stage of planning is to determine how to represent the boundary conditions and flow physics. Diffusers, thermal sources leading to plumes, and contaminant sources are all important flow details that need to be appropriately represented.
Dimensional Accuracy and Faithfulness to Details Highly detailed representation of architecture in a CFD model can significantly add to grid and computational costs. Often, high detail is not required. For instance, including doorknobs on a door will not likely change the simulation results greatly except immediately around the doorknob itself. However, if the room is negatively pressurized, significant flow through a crack beneath the door can cause a jet to propagate into the room, so accurately representing effects of flow through the crack is important. If the simulation is intended to assess thermal comfort in a laboratory with fume hoods, then the fume hood sash details may not be very important. However, if the purpose of the simulation is to evaluate fume hood capture performance, the sash opening detail can be important. For complicated models, it can be helpful to evaluate and include the potential modifications for subsequent simulations before completing the geometry. This allows the simulation to be modified without breaking the grid and geometry, which can be the most time-consuming step of simulation. Preparing the grid for future simulations may reduce overall costs for later simulations.
CFD Simulation Steps The basic mechanical steps in CFD simulation are as follows: 1. 2. 3. 4. 5.
Create the geometry (a 3D model of the simulation environment) Generate the grid Define surfaces and volumes and implement boundary conditions Execute the simulation Evaluate the simulation and conduct quality checks to determine whether CFD simulation is complete; refine the grid, change discretization, and continue to solve 6. Postprocess/analyze the simulation to extract desired information 7. Modify the simulation and redo as required During planning, a large set of physical phenomena (e.g., turbulence, heat transfer, radiation, species transport, combustion) may need to be included within the simulation(s). When executing the simulation itself, starting with a minimal set of physics and then increasing the level of complexity has advantages. For example, for flow in a room with a large convective and radiant heat source from which a contaminant is escaping (e.g., an industrial furnace), (1) solve the fluid velocity field with turbulence, (2) calculate energy to get a temperature distribution, (3) add radiation to redistribute some thermal energy, and (4) track the contaminant by adding species transport equations to the simulation. This stepped approach allows the modeler to build on previous flow field solutions and ensure that each new set of physics is starting from a reasonable estimate of the flow field. It has particular advantages for complicated models or increasing the probability of success for new users.
1.4
VERIFICATION, VALIDATION, AND REPORTING RESULTS
It is important to document and assess the credibility of CFD simulations through verification, validation, and reporting of results. The American Institute of Aeronautics and Astronautics (AIAA 1998) defines verification as “the process of determining that a (physical/ mathematical) model implementation accurately represents the
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developer’s conceptual description of the model and the solution on the model,” and validation as “the process of determining the degree to which a (CFD) model is an accurate representation of the real world from the perspective of the intended uses of the model.” Verification ensures that a CFD code can accurately and correctly solve the equations used in the conceptual model; it does not imply that the computational results represent physical reality. Generally, verification is done during code development. Because very few HVAC engineers who do indoor environment analysis develop CFD codes, this section focuses mostly on code applications, not development. In addition, time and money available for simulation are usually limited, which requires that verification and validation be realistically achievable. Therefore, this section refines the definitions of verification, validation, and reporting of results: • Verification identifies relevant physical phenomena for analysis and provides instructions on how to assess whether a particular CFD code can account for those physical phenomena. • Validation provides instructions on how to demonstrate the coupled ability of a user and a CFD code to accurately conduct representative indoor environmental simulations with available experimental data. • Reporting results provides instructions on how to summarize results so that others can make informed assessments of the value and quality of the CFD work. Therefore, verification should represent physical realities, although the cases can be very simple, containing only one (or a few) flow and heat transfer features of the complete system. The validation cases should be close to reality and include the flow and heat transfer characteristics that need to be analyzed, although approximations may be used in the validation. This section describes a procedure developed by Chen and Srebric (2002) for verification, validation, and reporting of CFD results, but its intent is not to develop standards. The extent of CFD’s capability in modeling has not yet been developed to the point where standards can be written (AIAA 1998).
Verification The basic physical phenomena of the indoor environment are airflow, heat transfer (conduction, convection, and radiation), mass transfer (species concentrations and solid and liquid particles), and chemical reactions (combustion). Therefore, the first step of verification is to identify benchmark cases with one or more flow and heat transfer features of those basic phenomena. In some indoor regions, airflow can be laminar or weakly turbulent. The overall flow features are often considered as turbulent. Most indoor airflows are turbulent because of the high Rayleigh number Ra, and sometimes high Reynolds number Re, defined as Ra = g TL2/k
(26)
Re = UL/
(27)
In most rooms, Ra ranges from 109 to 1012 and Re from 104 to 107 if room height is used as the characteristic length L. Experiments have found that turbulence occurs when Ra > 109 and/or Re > 104. Turbulence modeling approximations, which require more complex numerical schemes so that a converged solution can be achieved, must be made for CFD to solve the flow fields. The ability of a CFD code to simulate airflow in an indoor environment and the fidelity of the computer model to the physical realities may vary, and should be assessed. Predicting indoor physical phenomena may require auxiliary flow and heat transfer models. The following aspects require special attention: • Basic flow and heat transfer (convection, diffusion, conduction, and/or radiation)
• • • •
Turbulence models Auxiliary heat transfer and flow models Numerical methods Assessing CFD predictions
Whether a CFD code can be used to simulate an indoor environment depends on the flow and heat transfer features. For an indoor space with a baseboard heater, a CFD code that can solve natural convection flow may be sufficient. If a radiator replaces the baseboard heater, a radiation model is needed. When heat transfer through the walls must be considered, the code should have a conjugate heat transfer feature. When a duct supplies fresh air, room airflow becomes mixed convection, which requires the capability of mixed convection simulation. For indoor air quality studies, the code should be able to solve species concentrations. The more realistic the model is, the more complex the flow and heat transfer. To verify a CFD code’s capability of simulating the indoor environment of interest, review the code’s manual and any libraries or examples provided by the developer to illustrate successful applications. Discussing the particular application with the code developer can ensure that the physical models required for the application are all available. However, even if a code has been found capable of simulating the physical phenomena in the indoor environment in the past, repeating the verification is helpful because success relies on the joint function of the user and the CFD code. Hands-on verification usually starts with the simplest cases, which contain only one or two flow features and have been thoroughly tested to minimize uncertainties and errors. After successful verification, further simulations can be performed for more realistic cases, which may contain many key features of physical phenomena in an indoor space. The Reynolds or Rayleigh numbers can then be similar to those in reality. Often, verification data are from high-precision experimental measurements. The quantity and quality of the experimental data are usually accompanied by quantified errors. They are generally accurate, have few human errors, cover a large area of interest in the CFD community, and are widely used for testing CFD simulations. These experimental data contain detailed information, such as boundary and initial conditions, and are usually two-dimensional. Different cases represent different flow characteristics. Ideally, verification should be done for all flow features, but in practice, two to three important cases may be sufficient for most indoor environmental analyses. Turbulence Model Identification. With a verification case identified, the next step is to identify a suitable turbulence model. These are divided into two groups: large eddy simulations (LES), and turbulent transport models [Reynolds-averaged Navier-Stokes (RANS) equation modeling]. LES divides turbulent flow into largescale motion (calculated in LES) and small-scale motion (which must be modeled because of its effect on large-scale motion). Using a suitable subgrid scale model for the simulation is the most important factor, because the subgrid’s accuracy and efficiency determines how correct and useful the LES is. RANS models are more common in indoor airflow simulation. In general, eddy viscosity models are accurate for simple airflows, and Reynolds stress models are needed for complex flows. Complex flow exists in a flow domain with complex geometry, such as room and air supply diffuser geometry. Many CFD studies compare different turbulence models, so users may consult the literature for reported studies that are close to the case in question. The k- model is inaccurate for flows with adverse pressure gradient, which seriously limits its general usefulness. If comparisons for a particular case are not available, start with simple, popular models, such as the standard k- model (Launder and Spalding 1974), moving to progressively more complicated models if necessary. Vendors have made selecting different turbulence models as easy as a simple
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Indoor Environmental Modeling mouse click. However, a user should understand the principle of the model, its suitability for the problem to be solved, and the corresponding changes needed in using the model. Identifying Auxiliary Heat Transfer and Flow Models. The indoor environment consists of very complicated physical phenomena, with radiative, conductive, and convective heat transfer almost always occurring simultaneously, and sometimes also including combustion, participating media radiation, and particle transport in multiple phases (air/liquid, air/solid, and air/liquid/solid). It is important to verify whether these physical phenomena can be modeled by a CFD code. Separate verification of auxiliary models and turbulence models reduces the possibility of error. According to AIAA (1998), an error is “a recognizable deficiency in any phase or activity of modeling and simulation that is not due to lack of knowledge.” A complex problem may be verified by separating it into several components that have analytical solutions. For example, a combined conductive, convective, and radiative heat transfer process can be verified by separating it into a conductive and radiative problem, and a convective problem. The two problems can then be verified by the relevant analytical solutions. Another example is liquid particle trajectory in indoor air quality simulations that involve condensation, evaporation, and collision, as well as strong interaction with airflow turbulence. The physical phenomena should be verified separately. Verification does not ensure the correctness of the combined process. Therefore, uncertainty exists in the combined process. Uncertainty is a potential deficiency in any phase or activity of modeling caused by lack of knowledge; no highly accurate solutions are available. This may be addressed during validation. Verification of numerical methods involves investigating the discretization of the continuous space and time (if transient) into finite intervals. The variables are computed at only a finite number of locations (grid points), so the continuous information contained in the solution of differential equations is replaced with discrete values. When a Cartesian mesh system is used for sloped or curved surfaces, the true geometry is not represented in the calculation because it would introduce an error. Thus, for sloped or curved surfaces, similar geometries must be verified rather restricting the simulation to empty rectangular rooms. Different discretization schemes can be verified by comparing the results obtained from two different schemes. For example, to verify an unstructured grid system, Cartesian coordinates can be used as a reference. Case geometry should be simple, such as a rectangular room. Then, the two schemes should generate the same results. If the code has only one grid system, the discretization scheme verification can be combined with model verification. Refining Grid Size and Time Step. Because CFD discretizes partial differential equations into discretized equations, this introduces an error. Verification of grid size and time step is done to reduce error to a level acceptable for the particular application. The time step applies only to transient flow simulation. Therefore, it is not sufficient to perform CFD computations on a single fixed grid. The difference in grid size and time step between two cases should be large enough to identify differences in CFD results. The common method is to repeat the computation by doubling the grid number, and compare the two solutions (Wilcox 1998). It is very important to separate numerical error from turbulence-model error, because the merits of different turbulence models cannot be objectively evaluated unless the discretization error of the numerical algorithm is known. The geometry of an indoor space can be very complicated, and computer speed and capacity are still insufficient for simulating an indoor environment with very fine grid sizes (over a few million grids) and time steps (tens of thousands). Verification estimates the discretization error of the numerical solution. Theoretically, when grid size and time step approach zero, the discretization error of the numerical solution becomes negligible. For LES, when grid size
13.11 and time step become small, flow in the subgrid scale is isotropic and the results become more accurate. When grid size is much smaller than the Kolmogorov length scale, the LES turns into a direct-numerical simulation. Numerical Schemes, Iteration, and Convergence. A numerical scheme is important in CFD code to obtain a fast, accurate, and stable solution. A higher-order differencing scheme should be more accurate than a lower-order scheme for simple cases, such as those suggested for turbulence model verification. However, be aware of the limitations of various differencing schemes. For example, the central differencing scheme (accurate to the second order) is used for small Peclet numbers (Pe < 2), and the upwind scheme [accurate to the first order (but accounts for transportiveness)] is used for a high Peclet number. The Peclet number, the ratio of convection to conduction, is defined as Pe = LUCp /k
(28)
where Cp is specific heat. Solution algorithms in CFD codes can be quite different, ranging from SIMPLE in conventional program with iteration, to the fast Fourier transformation for solving the Poisson pressure equation in LES without iteration. Iteration is normally needed in two situations: (1) globally for boundary value problems (i.e., over the entire domain), or (2) within each time step for transient physical phenomena. Criteria can be set to determine whether a converged solution is reached, such as a specified absolute and relative residual tolerance. The residual is the imbalance of solved variables (e.g., velocities, mass flow, energy, turbulence quantities, species concentrations). For indoor environment modeling, a CFD solution has converged if Sum of absolute residuals in each cell Residual for mass = -----------------------------------------------------------------------------------------Total mass inflow < 0.1% Sum of absolute residuals in each cell Residual for energy = -----------------------------------------------------------------------------------------Total heat gains < 1% Similar convergence criteria can be defined for other solved variables, such as species concentration and turbulence parameters. Note that, for natural convection in a room, net mass flow is zero. Therefore, convergence has most likely been reached if there is little change (no change in the fourth digit) in the major dependent variables (temperature, velocities, and concentrations) within the last 100 iterations. However, a small relaxation factor can always give a false indication of convergence (Anderson et al. 1984). To obtain stable and converged results, iteration often uses relaxation factors for different variables solved, such as underrelaxation factors and false time steps. Underrelaxation factors differ only slightly from false-time-steps. Assessing CFD Predictions. A detailed qualitative and quantitative comparison of CFD results with data from experiments, analytical solutions, and direct numerical simulations is an important final step. All error analyses should be discussed in this section as well. The results indicate whether the CFD code can be used for indoor environment modeling. Although this procedure divides verification into several parts, they often are integrated. The turbulence model and numerical technique must work together to obtain a correct CFD prediction for the flow features selected. However, it is necessary to break them down into individual items in some types of verifications, such as in CFD code development. Indoor environment designers often use commercial software, and it is logical to assume that the codes were verified during development. However, the verifications (if any) may have used different flows that are irrelevant to indoor airflow. In addition, a user may not fully understand the code’s functions. It is imperative for the user to reverify a CFD code’s capabilities for indoor
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13.12 environment simulations. This helps the user become more familiar with the code and eliminates human errors in using the code. Generally, verification cases are not proprietary or restricted for security reasons; the data are usually available from the literature. It is strongly recommended that verification be reported when publishing CFD studies. This is especially helpful in eliminating user errors, because most CFD codes may have been validated by those cases. There are many examples of failed CFD simulations caused by user mistakes. Verification should be done for the following parameters: • All variables solved by the governing equations (e.g., velocity, temperature, species concentrations) • Boundary conditions (e.g., heat flux, mass inflow and outflow rates)
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With these items verified, a CFD code should be able to correctly compute airflow and heat transfer encountered in an indoor environment. The level of accuracy depends on the criteria used in the verification. If the code failed to compute the flow correctly, the problem may be that (1) the code is incapable of solving the indoor airflows, (2) the code has bugs, or (3) there are errors in the user input data that defines the problem to be solved.
Validation Validation demonstrates the ability of both the user and the code to accurately predict representative indoor environmental applications for which some sort of reliable data are available. It estimates how accurately the user can apply the code in simulating a full, realworld indoor environment problem, and gives the user the confidence to use the code for further applications, such as a design tool. A CFD code may solve the physical models selected to describe the real world, but may give inaccurate results because the selected models do not represent physical reality. For example, an indoor environment may simultaneously involve conduction, convection, and radiation, but a user may misinterpret the problem as purely convection. The CFD prediction may be correct for the convection part but fail to describe the complete physics involved. It is obviously a problem on the user’s side, which validation process also tries to eliminate. Note that validation addresses a complete flow and heat transfer system, or several subsystems that, together, represent a complete system. Although the procedure is almost the same, verification addresses only one of the flow aspects in an indoor environment. The basic idea of validation is to identify suitable experimental data, to make sure that all important phenomena in the problem are correctly modeled, and to quantify the error and uncertainty in the CFD simulation. Because the primary role of CFD in indoor environment modeling is to serve as a high-fidelity tool for design and analysis, it is essential to have a systematic, rational, and affordable code validation process. Validation focuses on • Confirming the capabilities of the turbulence model and other auxiliary models in predicting all the important physical phenomena associated with an indoor environment, before applying the CFD model for design and evaluation of a similar indoor environment category • Confirming correctness of the discretization method, grid resolution, and numerical algorithm for flow simulation • Confirming the user’s knowledge of the CFD code and understanding of the basic physics involved Ideally, validation should be performed for a complete indoor environment system that includes all important airflow and heat transfer physics and a full geometric configuration. Experimental data for a complete system can be obtained from on-site measurements or experiments in an environmental chamber. The data usually have a fairly high degree of uncertainty and large errors, and may
2017 ASHRAE Handbook—Fundamentals (SI) contain little information about initial and boundary conditions. Reasonable assumptions are needed to make CFD simulation feasible. Often, experimental data may not be available for a complete indoor environment system. In this case, validations for several subsystems or an incomplete system can be used. A subsystem represents some of the flow features in an indoor environment to be analyzed. The overall effect of several subsystems is equivalent to a complete system. For example, a complete indoor environment system consists of airflow and heat transfer in a room with occupants, furniture, and a forced air unit. If a user can correctly simulate several subsystems such as airflow and heat transfer (1) around a person, (2) in a room with obstacles, and (3) in a room with a forced air unit, the validation is acceptable. In the same example, an incomplete system for this environment can consist of airflow and heat transfer in a room with an occupant and a forced-air unit. Furniture, although it affects the indoor environment, is not as important as the other components, so validation with an incomplete system is acceptable. In either case, the key is that validation should lead to a solid confirmation of the combined capabilities of the user and code. Although validation is for a complete indoor environment system, it is not necessary to start with a very complicated case if alternatives are available. Reliability is better for • A simpler geometry, rather than a complicated one • Convection, rather than combined convection, conduction, and radiation • Single-phase flows, rather than multiphase flows • Chemically inert materials, rather than chemically reactive materials For complex physical phenomena in an indoor space, input data for CFD analysis may involve too much guesswork or imprecision. The available computer power may not be sufficient for high numerical accuracy, and the scientific knowledge base may be inadequate. Complete system validation should be broken down into steps: 1. Setting up building geometry, and then placing inlets and outlets. Isothermal flow indicates the airflow pattern. 2. Adding heat transfer. Species concentration, particle trajectory, etc., should be considered later. This progressive simulation procedure not only builds user confidence in performing the simulation, but also discovers some potential errors in the simulation. If a CFD code has multiple choices, simple, popular models should be considered as the starting point for validation. The starting point can be as basic as • • • • •
Standard k- model No-auxiliary-flow and heat transfer models Structured mesh system Upwind scheme SIMPLE algorithm
The way of measuring real-world accuracy of the representation is to systematically compare CFD simulations to experimental data. The indoor environment systems used in validation are usually complicated, and the corresponding experimental data may contain bias errors and random errors, which should be reported as part of the validation. If the errors are unknown, a report on the equipment used in the measurements is helpful in assessing data quality. Although desirable, it is expensive and time-consuming to obtain goodquality data for a complete system. Therefore, reporting the CFD validation of the complete system cannot be overemphasized. The criteria for accuracy when conducting a validation depend on the application. Very high accuracy, although desirable, is not essential because most design changes are incremental variations from a baseline. As long as the predicted trends are consistent, then
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Indoor Environmental Modeling less-than-perfect accuracy should be acceptable. The validation process should be flexible, allowing varying levels of accuracy, and be tolerant of incremental improvements as time and funding permit. The level of agreement achieved with the test data, taking into the account measurement uncertainties, should be reviewed in light of the CFD application requirements. For example, validation for modeling air temperature in a fire simulation requires much lower minimum accuracy than that for a thermal comfort study for an indoor environment. If validation cases are simple and represent a subsystem of a complex indoor airflow, the validation criteria should be more restrictive than those for the complete system. The criteria can also be selective. For example, if correct prediction of air velocity is more important, the criteria for heat transfer may be relaxed. Although air velocity and temperature are interrelated, one parameter’s effect on the other may be second-order. This allows a fast, less detailed model to be used, such as standard k- model, rather than a detailed but slower model, such as low-Reynolds-number model for heat transfer calculation in boundary layers.
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Reporting CFD Results Reporting involves summarizing CFD simulation results, while providing sufficient information on the value and quality of the CFD work. This is an important quality assurance strategy for CFD analysis of the indoor environment. It is recommended to start with verification and then proceed to validation. In principle, reports for technical audiences should include the information discussed in the verification and validation sections, such as • • • • • •
Experimental design CFD models and auxiliary heat transfer and flow models Boundary conditions Numerical methods Comparison of the CFD results with the data Drawing conclusions
The reporting format, however, can be flexible. If a report were intended for nontechnical readers, including only the last two items would be sufficient. Experimental Design. Thermal and flow conditions of the test environment should be described in enough detail that other people could repeat the simulation. This can be as simple as a reference to the literature or a description of cases in the report. An analysis of uncertainties and errors in the experimental data or a short description of the experimental procedure and equipment should also be included. CFD Models and Auxiliary Heat Transfer and Flow Models. CFD includes hundreds of different models of LES and RANS. Many popular turbulence models have been widely used and reported, making it unnecessary to provide detailed formulation. When reporting CFD results, it is important to specify which turbulence model is used. If the model has not been widely reported, detailed information (including why the model was selected) should be presented. Indoor environment analysis may require auxiliary heat transfer and flow models. For example, a building may use porous material as insulation. Heat transfer through the insulation combines conduction, convection, and radiation (too complex a process for limited computer resources to simulate in detail). Instead, a lumped-parameter model may be used to combine the heat transfer processes and obtain accurate CFD results for the indoor environment. Therefore, it should be described in the CFD analysis report. Boundary Conditions. Accurate specification of boundary conditions is crucial, because they indicate how the user interprets the specific physical phenomena into a computer model or mathematical equations that can be solved by the code. This interpretation requires
13.13 the most skill in CFD modeling. Therefore, detailed description of boundary conditions can help others make informed assessments of the simulation’s quality. Include the following information: • Geometry settings (the size of the computational domain along with sizes and locations of all solid objects represented in the model). If there is an external wall involved that cannot be considered adiabatic, external ambient conditions (e.g., ambient temperature, external radiation temperature, convective heat transfer coefficient) should be reported as well. • Inlet. Airflow from a diffuser greatly affects a room’s airflow pattern. Diffuser geometry, and approximations are often used in a complete system to make indoor airflow solvable. Therefore, the CFD report should give detailed information on the approximations used, as well as the set boundary conditions for the inlet. In some situations, the exact location of an inlet may be difficult to identify (e.g., air infiltration from the outdoors to an indoor space could be through the cracks of windows and doors). Conditions may differ from one window to another. Also, infiltration flow rate can be difficult to estimate because wind magnitude and direction change over time. Furthermore, turbulence parameters for the inlet are generally unknown, and should be estimated. Therefore, how these “inlet” conditions are specified should be clearly stated. • Outlet. An outlet has little effect on room airflow. However, conditions set for the outlet often can significantly influence numerical stability. For example, the outlet may become an inlet during iteration of a calculation. If the default outlet temperature is 0°C, this could lead to a diverged solution. • Walls. Rigid surfaces in an indoor space, such as walls, ceilings, floors, and furniture surfaces, are all considered as walls. Very close to the wall, airflow is laminar, and convective heat transfer often occurs in this region between the flow and surfaces. Many turbulence models cannot accurately handle the laminar sublayer, so ad hoc solutions, such as damping functions, are often used. How a CFD code treats wall boundary conditions greatly affects the accuracy of numerical results. Even if the indoor space is large and the wall effect seems small, accurate prediction of heat transfer from the walls to room air is still important. • Open boundary. When the area of interest is a part of the indoor space, the computational domain does not have to align with a rigid surface; instead, an “open” boundary can be defined. Depending on the inside and outside pressure difference, air may flow in or out across the open boundary. • Source/sink. This boundary condition fixes thermal or dynamic parameters (e.g., heat flux from a wall to simulate solar radiation, total heat flow in the occupied zone to simulate heat dissipation from occupants, momentum source from an operating fan, species generation rate from contaminant sources) in a defined region. Describe the location, size, and parameter being specified. • Coupling between a micro and a macro model. For a large indoor space, CFD analysis may be divided into micro and macro simulations. The micro simulation zooms into a particular area to reveal details of flow and thermal characteristic on a small scale compared with that for the entire space of interest. This allows finer-resolution examination of details of flow in that area. The macro simulation is applied to the entire flow system, and may use the results of the micro simulation so that a coarser grid system can be used. This coupling is usually a complicated procedure that should be detailed in the report. • Other approximations. Approximations are almost always involved when representing the real world in a computer model. For example, when the surface temperature distribution of a heated object is not uniform, the CFD simulation may choose to neglect temperature variation on the surface. When designing a large stadium, it may not be feasible to simulate each individual spectator; instead, the model may combine all the spectators into a human
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13.14
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layer. There are numerous examples in indoor environment modeling that need to be approximated in a CFD simulation. All approximations should be reported. Numerical Methods. It is essential to report the numerical methods used in the analysis, although the report can be brief if the technique is popular and widely available from the literature. The numerical technique includes discretization technique, grid size and quality, time step, numerical schemes, iteration number, and convergence criteria. The report should briefly state why the technique was used, and how suitable it is to the problem under consideration. It is also important to provide the quality indices of the mesh of a bodyfitted coordinate, because mesh quality affects the prediction’s accuracy. Typically, these indices include normal distances from solid surfaces to the centers of the first adjacent cells, maximum scale ratios of each two neighboring cells in each coordinate, and smallest angle of mesh cells. The first index determines the prediction of boundary layer flows, and the other two indicate whether unacceptable numerical errors are introduced into the simulation. Because a coarse grid introduces more numerical viscosity, gridrefinement study is essential to achieving a grid-independent solution, and should be included in the CFD report. Although it may not be realistic to conduct grid refinement for the complete system, it should be conducted for benchmark cases, to estimate the errors introduced in the complete system. If different numerical schemes have been tested, the results should be reported. Knowing the performance of different numerical schemes helps identify whether a numerical scheme or turbulence model causes a discrepancy between the CFD results and experimental data. Iteration number and convergence criteria are interrelated. It is better to use the sum of the absolute residual at each cell for all the variables as convergence criteria. The relaxation method and values should also be reported. Comparison of Results with Data. The most important part of the report is comparing experimental data with analysis results. Qualitative values, such as airflow pattern, should be compared first, followed by first-order parameters, such as air velocity, temperature, and species concentrations. In general, both CFD results and experimental data are more accurate for first-order parameters. Second-order parameters, such as turbulence kinetic energy, Reynolds stresses, and heat fluxes, usually have greater uncertainties and errors than the first-order parameters in both the results and the experimental data, so seeking perfect agreement for these parameters is unnecessary. It is insufficient to describe the comparison between CFD results and experimental data as excellent, good, fair, poor, or unacceptable. For example, a 20% difference can be considered excellent for a complex flow problem, but rather poor for two-dimensional forced convection in an empty room. Therefore, the comparison should be quantitative. The most useful information from comparison is how to interpret discrepancies. If there is little discrepancy, it is important to know why a turbulence model that uses approximations can predict the physical phenomena so well. The comparison should clearly state the uncertainties and errors of the experimental data, if they are known. Conclusions. The most important findings of the CFD analysis should be presented as its conclusions, which should have broad applicability to indoor environment simulation. The report may also recommend measures for further improvements in CFD analyses.
2017 ASHRAE Handbook—Fundamentals (SI) airflow and contaminant transport models. Thermal network models are addressed in Chapter 19.
2.1
MULTIZONE AIRFLOW MODELING
Theory Network airflow models idealize a building as a collection of zones, such as rooms, hallways, and duct junctions, joined by flow paths representing doors, windows, fans, ducts, etc. Thus, the user assembles a building description by connecting zones via the appropriate flow paths. The network model predicts zone-to-zone airflows based on the pressure-flow characteristics of the path models, and pressure differences across the paths. Three types of forces drive flow through the paths: wind, temperature differences (stack effect), and mechanical devices such as fans. As shown in Figure 13, airflow network models resemble electrical networks. Airflow corresponds to electric current, with zone pressure acting like the voltage at an electrical node. Flow paths correspond to resistors and other electrical elements, including active elements like batteries (fans). Unlike CFD models, network models do not prescribe details of airflow in zones. Thus, at any given time, each network zone is characterized by a single pressure. Pressure in the zone varies according to height, for example, using the simple hydrostatic relationship P + gh = constant. Air density is determined by the equation of state = P/(RairT ), based on the zone reference pressure P, temperature T, and the gas constant of the air mixture Rair . Zone temperature is given either directly by the user, or by an independent thermal model. The gas constant is typically assumed to be that of dry air, but can be made a function of other non-trace constituents as well (e.g., water vapor). This lack of detail in the network zone models makes CFD preferable for predicting thermal comfort, or designing displacement ventilation systems, where airflow patterns in a room control the quantities of interest (Emmerich 1997). In network airflow modeling, flow path models provide most of the modeling detail. Typically, the airflow rate Fj,i from zone j to zone i, in kg/s, is some function of the pressure drop Pj – Pi along the flow path: Fj,i = f (Pj – Pi)
Various models represent different types of flow paths, but they are typically nonlinear. For example, the power-law model is commonly implemented as Q = C(P) n where Q = F/ = volumetric airflow rate, m3/s
2. MULTIZONE NETWORK AIRFLOW AND CONTAMINANT TRANSPORT MODELING Multizone or network models are used to address airflow, contaminant transport, heat transfer, or some combination thereof. This section presents the mathematical and numerical background of network
(29)
Fig. 13 Airflow Path Diagram
(30)
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Indoor Environmental Modeling P C n
= = = =
13.15
pressure drop across opening, Pa flow coefficient, (Pa1/n ·m3)/s flow exponent (typically 0.5 to 0.6) density of air in flow path, kg/m3
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Pj,i is assumed to be governed by the Bernoulli equation, which accounts for static pressure on each side of the flow path and pressure differences through the flow path caused by density and height changes. Static pressure at flow path connections depends on the zone pressures, again after accounting for height-dependent pressure changes in the zones. Where a flow path connects to the building facade, the pressure also may depend on pressure imposed by wind (see Chapters 16 and 24). Typically, the calculated pressure drop through a flow path neglects heat transfer and changes in kinetic energy, but this is not an inherent limitation of the model. The power law model is based on engineering equations for orifice flow (see Chapter 16). Models for duct system components (e.g., dampers, bends, transitions) also follow the power law, with flow coefficient C given by tables (see Chapter 21). Other models describe the flow through doors and windows, fans, and so on (Dols and Walton 2002; Fuestel 1998). The network airflow model combines the flow element and zone relations by enforcing mass conservation at each zone. The mass of air mi in zone i is given by the ideal gas law Pi Vi mi = iVi = --------------R air T i
(31)
where mi i Vi Pi Ti Rair
= = = = = =
mass of air in zone i, kg zone density, kg/m3 zone volume, m3 zone pressure, Pa zone temperature, K gas constant for air = 287.055 J/(kg·K)
F j ,i + F i
Solution Techniques In a nodal formulation of the network airflow problem, zone pressures drive the problem. Specifically, the solution algorithm chooses one reference pressure for each zone, and then finds the driving pressure drops across each flow path, after accounting for changes of height in both zones and flow paths. Applying the element pressure/flow relations yields each path’s mass flow. Finally, these flows are summed for each zone to determine whether mass conservation is satisfied. This approach leads to a set of algebraic mass balance equations that must be satisfied simultaneously for any given point in time. Because airflows depend nonlinearly on node pressures, these equations are nonlinear, and therefore must be solved iteratively using a nonlinear equation solver. The simultaneous set of mass balance equations is typically solved using the Newton-Raphson method to “correct” the zone reference pressures until the simultaneous mass balance of all flows is achieved. This method requires a correction vector, which depends on the partial derivatives of relationships between flow and pressure for all flow connections. Therefore, these flow-pressure relationships must be first-order differentiable (Feustel 1998; Walton 1989). The Newton-Raphson method begins with an initial guess of the pressures. A new estimated vector of all zone pressures {P}* is computed from the current estimate of pressures {P} by {P}* = {P} – {C}
For a transient solution, the principle of conservation of mass states that m i V i i --------- = i -------- + V i -------- = t t t
duct junction because of flow in another branch (see Chapter 21), or effects of zone geometry (e.g., short-circuiting of a room when a ventilation supply duct blows air directly into a return air intake). For momentum-based effects, a CFD model of the room is preferable to a network model.
(32)
where the correction vector {C} is computed by the matrix relationship [J]{C} = {B}
(36)
where {B} is a column vector of total flow into each zone, with each element given by
j
m i 1 Pi V i --------- ----- --------------- – m i t – t t t R air Ti t
(35)
Bi = (33)
F j ,i
(37)
j
[J] is the square (i.e., N by N for a network of N zones) Jacobian matrix whose elements are given by
where Fj,i = airflow rate between zones j and i (positive values indicate flows from j to i; negative values indicate flows from i to j), kg/s Fi = nonflow processes that could add or remove significant quantities of air flows from j to i; negative values indicate flows from i to j
However, airflows are typically calculated for steady-state conditions. This is reasonable for most cases where driving forces change slowly compared to the airflow (e.g., because of the building’s large thermal mass, or because rate-limited actuators change damper and fan settings slowly compared to the rate at which the airflow system reestablishes a steady state). Under this quasi-steady assumption, mass conservation in zone i reduces to
Fj , i
=0
(34)
j
This model was based on the assumption that airflows were quiescent and that the zones’ resistance to airflow was negligible relative to the resistance imposed by the airflow paths that connect the zones. Hence, the model enforces conservation of mass in each zone, but does not conserve momentum. This means it cannot model some effects, such as the suction that develops in one branch of a
Ji , j =
F j , i --------- Pj-
(38)
i
In Equations (37) and (38), Fj,i and Fj,i /Pj are evaluated using the current estimate of pressure {P}. Equation (35) represents a set of linear equations which must be solved iteratively until a convergent solution of the set of zone pressures is achieved. In its full form, [J] requires computer memory for N 2 values, and a standard Gauss elimination solution has execution time proportional to N 3. Sparse matrix methods can be used to reduce both the storage and execution time requirements. Two solution methods for the linear equations have been successfully implemented: Skyline (also called the profile method) and preconditioned conjugate gradient (PCG), which may be useful for problems with many zones and junctions (Dols and Walton 2002). The number of iterations needed to find a solution may be reduced by applying descent-based techniques to Newton-Raphson (Dennis and Schnabel 1996). Under a fairly modest set of conditions, line search methods are guaranteed to converge to a unique solution (Lorenzetti 2002).
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13.16
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2.2
2017 ASHRAE Handbook—Fundamentals (SI) CONTAMINANT TRANSPORT MODELING
2.3
MULTIZONE MODELING APPROACHES
Fundamentals
Simulation Planning
Multizone contaminant transport models generally address transport of contaminants by advection via interzone airflows and mechanical system flows while accounting for some or all of the following: contaminant generation by various sources or chemical reaction, removal by filtration, chemical reaction, radiochemical decay, settling, or sorption of contaminants. Unlike CFD models, the details of contaminant distribution within a zone are not modeled: each zone is considered well-mixed and characterized by a single concentration at any given point in time. Therefore, the well-mixed assumption’s applicability to the mixing time and pattern of airflow in a zone should be considered. For example, the well-mixed assumption may be quite appropriate for zones with a mixing time well within the solution time step of interest (e.g., long-term off-gassing of building materials in common ventilation system configurations with relatively steady airflows). However, if a zone is characterized by steep concentration gradients and the time step of interest is relatively short (e.g., a chemical release in a relatively large zone), CFD analysis might be more appropriate. This is especially true if the reason for analysis is to resolve concentration gradients within the zone.
Planning can improve results and reduce the amount of input effort required in multizone simulations. The most important steps are determining what aspects of flow and contaminant transport are being studied, and what driving forces are likely to be most important. One of the first decisions is defining zones in the model. The level of detail needed depends on both the building and scenario being modeled. For a study of contaminant transport from a garage into a house, separate zone models of clothes closets or kitchen cabinets are unnecessary and typically would only be needed if, for example, the source of the contaminant were inside the closet or cabinet. Because zones are typically broken where there are obstructions to air movement and/or differences in air properties, often a doorway between adjacent rooms is an appropriate place to define zones. Therefore, usually, a good starting point is to consider each room as a separate zone, and then model smaller enclosures in more detail or subdivide nonuniform rooms as necessary. On the other hand, sometimes the problem statement allows several rooms to be grouped together as a single zone. This is usually done to save user input time, because multizone models of even very large buildings can be quickly simulated on a desktop PC. HVAC system zoning also provides cues about how to group zones. Considering the primary driving forces (natural, mechanical, etc.) and the relative resistance of the flow elements that connect the zones to these forces can also be helpful. Starting with the assumption that each room is a zone, these changes can be made as the physics of the problem allow. The type of simulation must also be determined. Are only airflow data necessary, or are contaminant concentrations also needed? Are the flow and contaminant transport problems steady-state, cyclical, or transient? Note that they may not be the same. Some models can simulate steady-state flow and transient contaminant transport. During planning, the required model input data and boundary condition information must be specified, with the level of detail depending on the problem. Some items to consider are exterior envelope and interzonal leakage, weather conditions, wind pressure profiles, contaminant characteristics, contaminant source types and strengths, mechanical system flow rates, control algorithms, occupancy, and zone volumes.
Solution Techniques Generally, the goal is to solve a set of mass balance equations for each contaminant in each zone. The mass of contaminant in zone i is m,i = mi C,i
(39)
where mi is the mass of air in zone i and C,i is the concentration mass fraction of (kg of /kg of air). Contaminant is removed from zone i by • Outward airflows from the zone at a rate of j Fi,j C,j , where Fi,j is the rate of air flow from zone i to zone j • Removal at the rate C,i R,i where R,i (kg of air/s) is a removal coefficient • First-order chemical reactions with other contaminants C,i (kg of /kg of air) at rate mi C,i, where (1/s) is the kinetic reaction coefficient in zone i between species and Contaminant is added to the zone by • Inward airflows at rate j (1 – ,j,i)Fj,iC,j where j,i is the filter efficiency in the path from zone j to zone i • Generation at rate G,i (kg of /s) • Reactions of other contaminants
Steps
Conservation of contaminant mass for each species and assuming trace dispersal (i.e., m,i << mi) produces the following basic equation for contaminant dispersal for a given zone in a building:
1. Input zones and building geometry, using just enough detail to capture necessary information. 2. Determine and specify building leakage. This information may be obtained from blower door or tracer gas tests of the actual building, or estimated based on published data (see Chapter 16). Perform the following tests:
dm ,i ------------- = – R ,i C ,i – Fi , j C ,i + Fi, j 1 – , j ,i C , j dt + m i , C ,i + G ,i j
j
(40)
This equation must be developed and solved for all zones to determine each contaminant’s concentration. The various techniques for solving the ensuing set of equations can be categorized by the fundamental control volume used to develop them (i.e., Eulerian or Lagrangian), and by whether the analysis is geared towards solving the steady-state or dynamic system, or determining analytical solutions via eigen-analysis (Axley 1987, 1988; Dols and Walton 2002; Rodriguez and Allard 1992).
The following process is typical of that used by experienced modelers to help catch mistakes and verify that the model is as intended. Always remember to save and test the model often.
• Simulated blower door test: Within the model, set all interior doors in the building to open and put a large pressurization fan in an exterior wall. Pressurize the building and use the fan’s flow rate and consequent pressure difference across the exterior wall to calculate the leakage area per area of exterior wall. Verify specification of the proper amount of leakage on all walls by comparing inputs with data from an actual building or the literature. This is especially important when specifying individual leakage paths. If a result does not make sense, adjust leakage paths to see their effect on overall leakage. • Simulation with typical weather boundary conditions. Verify that the resulting infiltration rate is realistic.
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Indoor Environmental Modeling 3. Check stack effects. Remove the blower door fan from the model, specify a very cold outdoor temperature, and run a simulation. Check the location of the neutral pressure level, which is the collection of points on the building envelope where the pressure difference with the outdoors is zero. The points usually form a plane at the building’s midheight, though it may be a bit higher if roof leakage occurs with no corresponding floor leakage. A very high or low neutral pressure level could indicate large unintended leaks, probably somewhere near the neutral pressure level. (Note that, in complex operating systems or scenarios, the neutral pressure level may not form a plane and could change with time.) 4. Specify wind and wind pressure profiles on exterior leakage paths: Some programs allow wind specification, but this has no effect unless the wind pressure profile for each path is also specified. Perform the following tests: • Run a simulation with no stack effect or mechanical system, and a high wind. Flow should be visible through each exterior path. This allows quick identification of paths that may be missing a wind pressure profile. • Verify that inflows and outflows are as prescribed for the wind pressure profile. (For example, inflow on the windward side, outflow for walls at negative pressure) This helps verify that the pressure profile is correctly input and that the building and wind are oriented properly. • Try other terrain conditions to see if changing this variable significantly affects results. 5. Input air-handling system(s) (if any). Again, use only as much detail as is necessary. For small buildings, it may be reasonable to represent the air-handling system as a simple fan through an exterior wall for ventilation. For internal distribution from zone to zone, HVAC system flows in and out of each zone must be specified. Duct details should be included only if they are an important aspect of the problem. Sometimes supply and return vents can be placed in plenums or other locations where duct leakage is expected, to approximate duct leakage. However, if pressure-driven leakage must be modeled, the ducts should be specified in detail. Keep in mind that leakage in VAV systems, for example, should be separately specified upstream and downstream of the VAV box. Perform the following tests: • Run a simulation with no stack or wind driving forces and check outdoor, return, and exhaust air volumes to verify that the outside air is properly defined. This is a common beginner’s mistake because there are several ways to specify outside air (e.g., setting a percentage of outside air, or scheduling outside air when the default may be 100%), and they may override one another. Also note that the amount of return air specified must equal or exceed the recirculation air needed. Otherwise, outdoor air may be used to make up the difference. • Verify a realistic pressure difference across walls. For example, a 50 Pa pressure difference would not occur in a real house, and probably indicates problems with either the system or leakage input. It is important that the magnitude of the pressure differences makes sense for the situation being modeled. 6. Specify contaminants. Contaminant sources are usually specified on either a mass or volume basis, although numerical counts (e.g., of particles, spores) can also be modeled and can typically be interchanged with mass units. Model refinement is often most appropriate and desirable near the contaminant source, where large concentration gradients are present. Simulation tests for pressure and velocity suggest the expected accuracy when a contaminant is added to the system. Transport of contaminants, particularly aerosols, is also influenced by other mechanisms, for which coefficients are specified in the basic transport equations.
13.17 • Verify that the model predicts conservation of contaminant mass across multiple zones. • Use experimental tracer analysis using dynamically similar, nontoxic materials (if desired and feasible). 7. Run a sensitivity analysis: Deviations in some variables may need to be considered. Depending on the source of the input data and type of simulation, it is often good to know how the system performs over a range of certain variables. Possible items to consider include • Formal sensitivity analysis, if resources permit. • Leakage dependence, tested under a range of values. If conclusions are too leakage-dependent and leakage test data are not available, then a range of possible results should be considered. • System pressure balance. In a building with multiple airhandling systems, their design flow rates may imply perfect balance between systems; however, in real buildings, the balance will never be perfect, which can drive contaminants into shafts and distribute them through the building. Pressurizing or depressurizing a floor slightly compared to others (by specifying slightly imbalanced system flows) can illustrate how big this effect is. • Weather effects, which can be particularly important when studying infiltration or trying to maintain a pressure differential somewhere in the system. Verify that the system can accommodate the expected range of outdoor conditions.
2.4
VERIFICATION AND VALIDATION
Verification and validation of multizone models are similar in many regards to that of CFD models. Because the number of cases a complex multizone model can simulate is unlimited, Herrlin (1992) concluded that absolute validation is impossible. However, validation is still important to identify and eliminate large errors and to establish the model’s range of applicability. Therefore, a model’s performance should be evaluated under a variety of situations, with the recognition that predictions will always have a degree of uncertainty. Herrlin lists three techniques of model validation: • Analytical verification (comparison to simple analytically solved cases) • Intermodel comparison (comparison of one model to another) • Empirical validation (comparison to experimental tests) Herrlin also discussed some specific difficulties in validating multizone airflow models, including input uncertainty (particularly of air leakage distribution) and attempting to simulate processes that cannot be modeled (e.g., using a steady-state airflow model to simulate dynamic airflow). ASTM Standard D5157 provides information on establishing evaluation objectives, choosing data sets for evaluation, statistical tools for assessing model performance, and considerations in applying statistical tools. It stresses that data used for the evaluation process should be independent of the data used to develop the model. Also, sufficiently detailed information should be available for both the measured pollutant concentrations and the required input parameters. Standard D5157 also discusses the fact that model validation consists of multiple evaluations, with each evaluation assessing performance in specific situations.
Analytical Verification Analytical verifications of multizone modeling tools are routinely performed to check the numerical solution. Analytical test cases are simple forms of problems that can be solved analytically to compare the model with an exact solution. For multizone models, these include airflow elements in series and parallel; stack effect; wind pressure effect; fan and duct elements; contaminant generation,
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2017 ASHRAE Handbook—Fundamentals (SI)
dispersal, filtration, and deposition; and simple kinetic reactions. These tests are typically performed by model developers, but may be repeated by the user to develop confidence in the model and to verify the user’s familiarity with the model. Such tests are not routinely published, but some were described by Walton (1989). Unfortunately, most buildings are too complicated for the equations describing airflow and pollutant transport to be solved analytically. Therefore, analytical verification is of limited value in determining the adequacy of a multizone IAQ model for practical applications.
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Intermodel Comparison Intermodel comparison provides a relative check of different models’ assumptions and numerical solutions. As with analytical verification, this is of limited value in evaluating a model’s adequacy for practical applications. Generally, intermodel comparisons are not essential to a user, although good comparisons allow empirical validation conclusions to be generalized beyond the specific model studied. Haghighat and Megri (1996) reported good agreement between CONTAM [the predecessor of CONTAMW (Dols and Walton 2002)], COMIS (Feustel et al. 1989), AIRNET (Walton 1989), CBSAIR (Haghighat and Rao 1991), and BUS (Tuomaala 1993) for airflow predictions for a four-zone model. The model building was two stories tall, with power-law flow elements for leakage. A single set of temperatures and wind-induced pressures was simulated. Model predictions for zone pressures and flow rates were within 5% and 13%, respectively. Orme (2000) also found good agreement overall between CONTAM, COMIS, MZAP (unpublished), and BREEZE (BRE 1994) airflow predictions for a three-story building model. Power-law airflow elements were used to connect the four interior zones with each other and the ambient zone. A single wind speed and ambient temperature condition were applied. Note that both of these intermodel comparisons tested models for only a very limited range of conditions.
Empirical Validation Empirical validation compares model assumptions and numerical solutions to indoor environmental problems of practical interest. However, the standard is only as accurate or realistic as the measurements used to produce it. Not only do all models have uncertainty, but all measurements do as well. Differences between model predictions and measurements could stem from errors in either set of data. As discussed in the section on CFD Modeling Approaches, comparison depends on the numerical model’s capabilities and limiting assumptions, as well as the modeler’s knowledge of both the model being applied and the indoor environment being modeled. It is essential to apply valid statistical tools when interpreting comparisons of measurements and predictions. ASTM Standard D5157 provides three statistical tools for evaluating accuracy of IAQ predictions, and two additional statistical tools for assessing bias. Values for these statistical criteria are provided to indicate whether model performance is adequate. Note that the criteria and specific values in Standard D5157 are not ultimate arbiters of model accuracy, but they provide a useful template for the type of statistical analysis needed. Other valid statistical criteria may be substituted, with values appropriate for the accuracy needed for a specific project. ASTM Standard D5157 suggests the following for assessing agreement between predictions: • The correlation coefficient of predictions versus measurements should be 0.9 or greater. • The line of regression between predictions and measurements should have a slope between 0.75 and 1.25 and an intercept less than 25% of the average measured concentration.
Fig. 14 Floor Plan of Living Area Level of Manufactured House • The normalized mean square error (NMSE) should be less than 0.25. NMSE is calculated as
Cpi – Coi N
NMSE =
2
nC o C p
(41)
i=1
where Cp is the predicted concentration and Co is the observed concentration. For assessing bias, 1. The normalized or fractional bias FB of mean concentrations should be 0.25 or lower, and is calculated as FB = 2 C p – C o C p + C o
(42)
2. The fractional bias FS of variance should be 0.5 or lower. FS is calculated as 2
2
2
2
FS = 2 C p – C o C p + C o
(43)
Emmerich (2001) reviewed the research literature for reports of empirical multizone model validation for residential-scale buildings. Few reviewed reports used either the ASTM Standard D5157 measures or other limited statistical evaluations to evaluate the results. However, for those cases with sufficient published data, Emmerich calculated several statistical measures from Standard D5157. Although these measures specifically address concentrations, they have been used to compare predicted and measured airflow rates also. Table 1 summarizes these published multizone model validation efforts. There are many published validations for residential buildings, but far fewer for large commercial buildings because of the significant effort and cost involved in detailed measuring of a large building. Commercial studies are available by Furbringer et al. (1993), Said and MacDonald (1991), and Upham (1997). Example 1. Ventilation Characterization of a New Manufactured House. Develop a multizone model to investigate various ventilation strategies of a new double-wide manufactured home consisting of three levels: crawlspace, living area, and attic. The crawlspace is divided into two sections by an insulated plastic belly; the region above the belly contains HVAC ductwork, and the volume below vents to the outdoors. The living area is shown in Figure 14. The attic comprises the volume above the vaulted ceiling, with five roof vents and eave vents spanning the perimeter of the house. Figure 15 provides a schematic of the house, showing connections between the levels and the air distribution system. The building has an automated data acquisition system for monitoring air temperatures, building pressures, weather, and HVAC operation. The instrumentation system also has an automated tracer gas system for continuous monitoring of building air change rates. The tracer gas system injects sulfur hexafluoride into the house every 4 to 6 h, allows it to mix to a uniform concentration, and then monitors the
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13.19
Table 1 Summary of Multizone Model Validation Reports Reference
Test Building
Model
Bassett 1990
Five houses
Blomsterberg et al. 1999
Houses and apartment flats
CONTAM Zone air change rates Interzone airflows COMIS Average whole-house air change rates Average room air change rates COMIS Average interzone airflows CONTAM 0.3 to 5.0 m particle concentrations CONTAM Methylene chloride concentration CONTAM Transient CO concentration (zone 1) Transient CO concentration (zone 2) CONTAM Interzone airflows
Borchiellini et al. 1995 Two test houses Emmerich and Nabinger Single-zone test 2000 house Koontz et al. 1992 Test chamber Two-zone research house
Haghighat and Megri 1996
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Lansari et al. 1996
Sextro et al. 1999 Yoshino et al. 1995
Zhao et al. 1998
Multizone laboratory House Two-story house with garage
Three-story test building Three-room test house Test house
Parameter Evaluated
R
m
B
NMSE
FB
0.91 0.27 0.98
1.31 0.10 1.04
–0.23 1.34 –0.03
0.35 2.98 0.01
0.08 0.37 0.01
0.72
0.70
0.32
0.24
0.03
0.84 0.94 to 0.99 0.98
0.60 0.84 to 1.02 1.08
0.18 –0.25 to 0.29 0.07
0.41 0.04 to 0.19 0.20
–0.24 –0.26 to 0.16 0.16
0.94
0.70
0.14
0.15
0.06
0.98
0.85
0.26
0.02
–0.11
0.96
0.90
0.10
0.18
0.002
0.96 0.97
0.84 1.07
0.14 0.10
0.04 0.01
–0.02 –0.03
CONTAM Room airflows CONTAM Tracer gas concentrations in garage Tracer gas concentrations in other rooms CONTAM Tracer gas concentrations
0.92
0.94
0.18
0.12
–0.27
0.97
1.04
0.14
0.10
0.16
COMIS
Air change rates
0.79
0.87
NA
NA
NA
COMIS
Tracer gas concentrations Room air change rates Tracer gas concentrations
0.98 0.72 0.93
1.06 0.92 0.93
NA NA NA
NA NA NA
NA NA NA
Source: Emmerich (2001). Note: R = correlation coefficient; m = slope of regression line; B = ratio of intercept of regression line to average measured value; NMSE = normalized mean square error; FB = fractional bias.
Fig. 15 Schematic of Ventilation System and Envelope Leakage concentration decay in all the major zones of the building. Air change rates are then calculated based on the tracer gas decay rate in the living space. Model Description. The model contains four levels: crawlspace, belly volume, living area (Figure 16), and attic. The duct modeling capabilities (see Figure 17 depicting belly level) were used to model the forced-air system. Leakage values of model airflow paths are listed in Table 2. Leaks in the living space envelope include the exterior wall and interfaces between the ceiling and wall, floor and wall, and the
walls at the corners. In addition, there are two types of windows, the exterior doors, and the living space floor, which contains openings into the belly. There are also interior airflow paths, including leaks in the walls, doorframes, and open doors. Note that for all the tests and simulations performed, all interior doors were open. The attic has leakage in its floor (i.e., the ceiling of the living space), as well as the two types of attic vents to the outdoors. The crawlspace has leaks to the outdoors in the walls, vents in the front and rear of the house, and an access door. The model also includes a leak from the crawlspace into the belly. Finally, the duct leak into the belly, based on the described measurement, is included in the model. Results. Tracer gas decay tests were simulated using the multizone model. Figure 18 shows the results of one of these simulations 30 min after injection of the tracer gas. The darker the shading, the higher the tracer gas concentration. Figure 19A shows the measured and predicted air change rates with the forced-air system off as a function of indoor/outdoor air temperature difference under low wind speed conditions. Values predicted with the model are in good agreement with the measurements, particularly at low values of T, but tend to underpredict by around 20% at higher values. Note that in all reported measurements and predictions, the outdoor air intake on the forced-air system and the window inlet vents are closed. Figure 19B plots the measured and predicted air change rates with the forced-air system on, again for low wind speeds. Under positive temperature differences, the measured air change rates are actually lower than with the system off, which might not be expected with significant duct leakage. Airflow measurements indicate that the system moves about 450 L/s, but about 125 L/s is lost through duct leakage into the belly. Some of this airflow returns to the living space through leaks in the floor, but some flows through the crawlspace to the outdoors, which tends to depressurize the house. A significant air change rate is seen at zero T, but this is not unexpected given the duct leakage. At higher values of T, the stack effect “competes” with duct leakage into the belly, decreasing the air change rate into the house. This effect has actually been proposed as a means of controlling airflow and
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Fig. 16 Multizone Representation of First Floor
Fig. 17 Multizone Representation of Ductwork in Belly and Crawlspace contaminant entry from crawlspaces (Phaff and De Gids 1994). Overall, with the fan on, the agreement between the predicted and measured air change rates is quite good.
2.5 A0 B {B} {C} C Co cp Cp cR C,i C E
= = = = = = = = = = = =
f= FB = Fi = Fj,i =
SYMBOLS
diffuser effective area ratio of intercept of regression line to average measured value column vector of total flow into each zone correction vector flow coefficient, (Pa1/n ·m3)/s; concentration mass fraction observed concentration concentration in air specific heat; predicted concentration mean concentration in return openings concentration mass fraction of , kg /kg of air k- turbulence model constant constant depending on wall roughness (9.8 for hydraulically smooth walls) [see Equation (29)] fractional bias of mean concentrations nonflow processes that could add or remove significant quantities of air mass flow rate from zone j to zone i, kg/s
FS g G,i G h [J] k l L m M mi m· n NMSE P P {P} {P*} Pe Pi, Pj Q R Rair
= = = = = = = = = = = = = = = = = = = = = = = =
fractional bias of variance gravitational acceleration generation rate of contaminant , kg /s filter function height; convective heat transfer coefficient square Jacobian matrix thermal conductivity; turbulent kinetic energy turbulence length scale characteristic length (e.g., room height, diffuser height) mass of air, kg slope of regression line mass of air in zone i, kg mass flow rate flow exponent (typically 0.5 to 0.6) normalized mean square error pressure pressure drop across opening, Pa estimated pressures estimated vector of all zone pressures Peclet number (ratio of convection to conduction) zone pressure, Pa volumetric airflow rate, F/, m3/s removal coefficient; correlation coefficient gas constant of air, 287.055 J/(kg·K)
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13.21 Table 2 Leakage Values of Model Airflow Components Exterior Airflow Paths
14 mm2/m2 81 mm2/m 124 mm2/m 500 mm2 194 mm2 80.8 mm2/m 1870 mm2 365 mm2/m2
Interior airflow paths
Interior walls Bedroom doorframe Open interior doors Bathroom doorframe Interior doorframe Closet doorframe
200 mm2/m2 4100 mm2 2 by 0.9 m 3300 mm2 2500 mm2 460 mm2
Attic
Attic floor Roof vents Eave vents
200 mm2/m2 0.135 m2 10 600 mm2/m
Crawlspace and belly
Exterior walls of crawlspace Rear crawlspace vents Front crawlspace vents Crawlspace access door Crawlspace to belly Duct leak into belly
2500 mm2/m2 32 300 mm2 46 500 mm2 20 600 mm2 25 800 mm2 3200 mm2
Fig. 18 Test Simulation of Concentration of Tracer Gas Decay in Manufactured House 30 min After Injection
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ELA at 4 Pa
Living space envelope Exterior wall Ceiling wall interface Floor wall interface Window #1 Window #2 Corner interface Exterior doors Living space floor to belly volume
Ti U ui ui u i uj UP Uref U0 Vi xi xj Y+ Yp
= = = = = = = = = = = = = =
zone temperature, K velocity instantaneous velocity ensemble average of for steady flow fluctuation velocity velocity in j direction, m/s velocity parallel to wall at Yp mean stream velocity air supply velocity for momentum source zone volume, m3 distance in i direction, m distance in j direction, m dimensionless distance from wall distance from wall to center of first cell
Greek
Fig. 19
Ra Re sij S t T T TI Ts
= = = = = = = = =
Measured and Predicted Air Change Rates for Wind Speeds less than 2 m/s Rayleigh number Reynolds number strain rate tensor source or sink time, s temperature, °C temperature change turbulence intensity, % surface temperature
= contaminants = generalized diffusion coefficient or transport property of fluid flow = dissipation rate of turbulent kinetic energy = Kolmogorov length scale; filter efficiency = von Karmann’s constant (0.41) , = kinetic reaction coefficient between and = dynamic viscosity t = eddy viscosity = kinematic viscosity = density, kg/m3 i = zone density, kg/m3 = standard deviation ij = viscous tensor stress w = wall shear stress = transport property (1 for mass continuity, momentum, temperature, or species concentration)
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore.
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13.22 AIAA. 1998. Guide for the verification and validation of computational fluid dynamics simulations. AIAA Standard G-077-1998. American Institute of Aeronautics and Astronautics, Reston, VA. Anderson, J.D., Jr. 1995. Computational fluid dynamics: The basics with applications. McGraw-Hill, New York. Anderson, D.A., J.C. Tannehill, and R.H. Pletcher. 1984. Computational fluid dynamics and heat transfer. Hemisphere, Washington, D.C. ASTM. 1997. Standard guide for statistical evaluation of indoor air quality models. Standard D5157-97(2003)e1. American Society for Testing and Materials, West Conshohocken, PA. Awbi, H.B. 1991. Ventilation of buildings. Chapman & Hall, London. Axley, J.A. 1987. Indoor air quality modeling, phase II report. NBSIR 873661. National Institute of Standards and Technology, Gaithersburg, MD. Axley, J.A. 1988. Progress toward a general analytical method for predicting indoor air pollution in buildings—Indoor air quality modeling, phase III report. NBSIR 88-3814. National Institute of Standards and Technology, Gaithersburg, MD. Baker, A.J., P.T. Williams, and R.M. Kelso. 1994. Numerical calculation of room air motion—Part 1. ASHRAE Transactions 100(1):514-530. Bassett, M. 1990. Infiltration and leakage paths in single family houses—A multizone infiltration case study. AIVC Technical Note 27. Air Infiltration and Ventilation Centre, Brussels, Belgium. Beausoleil-Morrison, I. 2000. The adaptive coupling of heat and air flow modelling with dynamic whole-building. Ph.D. dissertation, Department of Mechanical Engineering, University of Strathclyde, Glasgow. Bernard, P.S., and J.M. Wallace. 2002. Turbulent flow: Analysis, measurement, and prediction. John Wiley & Sons, Hoboken, NJ. Blomsterberg, A., T. Carlsson, C. Svensson, and J. Kronvall. 1999. Air flows in dwellings—Simulations and measurements. Energy and Buildings 30:87-95. Borchiellini, R., M. Cali, and M. Torchio. 1995. Experimental evaluation of COMIS results for ventilation of a detached house. ASHRAE Transactions 101(1). BRE. 1994. BREEZE 6.0 user manual. Building Research Establishment, Garston, U.K. Breuer, M. 1998. Large eddy simulation of the subcritical flow past a circular cylinder: Numerical and modeling aspects. International Journal of Numerical Methods in Fluids 28:1281-1302. Chen, Q. 1988. Indoor airflow, air quality and energy consumption of buildings. Ph.D. dissertation, Delft University of Technology, The Netherlands. Chen, Q. 1996. Prediction of room air motion by Reynolds-stress models. Building and Environment 31(3):233-244. Chen, Q., and Z. Jiang. 1992. Significant questions in predicting room air motion. ASHRAE Transactions 98(1):929-939. Chen, Q., and A. Moser. 1991. Simulation of a multiple-nozzle diffuser. Proceedings of the 12th AIVC Conference 2:1-14. Chen, Q., A. Moser, and A. Huber. 1990. Prediction of buoyant, turbulent flow by a low Reynolds-number k- model. ASHRAE Transactions 96(1):564-573. Chen, Q., and J. Srebric. 2000. Simplified diffuser boundary conditions for numerical room airflow models. ASHRAE Research Project (RP) 1009, Final Report. Chen, Q., and J. Srebric. 2002. A procedure for verification, validation, and reporting of indoor environment CFD analyses. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 8(2):201-216. Chen, Q., and W. Xu. 1998. A zero-equation turbulence model for indoor airflow simulation. Energy and Buildings 28(2):137-144. Cherukat, P., Y. Na, and T. J. Hanratty. 1998. Direct numerical simulation of a fully developed turbulent flow over a wavy wall. Theoretical and Computational Fluid Dynamics 11:109-134. Chester, S., F. Charlette, and C. Meneveau. 2001. Dynamic model for LES without test filtering: Quantifying the accuracy of Taylor series approximations. Theoretical and Computational Fluid Dynamics 15:165-181. Corrsin, S. 1961. Turbulent flow. American Scientist 49:300-325. Dennis, J.E., and R.B. Schnabel. 1996. Numerical methods for unconstrained optimization and nonlinear equations. Society for Industrial and Applied Mathematics, Philadelphia. Dols, W.S., and G. Walton. 2002. CONTAMW 2.0 user manual. NISTIR 6921. National Institute of Standards and Technology, Gaithersburg, MD.
2017 ASHRAE Handbook—Fundamentals (SI) Emmerich, S.J. 1997. Use of computational fluid dynamics to analyze indoor air quality issues. NISTIR 5997. National Institute of Standards and Technology, Gaithersburg, MD. Emmerich, S.J. 2001. Validation of multizone IAQ modeling of residentialscale buildings: A review. ASHRAE Transactions 107. Emmerich, S.J., and K.B. McGrattan. 1998. Application of a large eddy simulation model to study room airflow ASHRAE Transactions 104(1):1-9. Emmerich, S.J., and S.J. Nabinger. 2000. Measurement and simulation of the IAQ impact of particle air cleaner in a single-zone building. NISTIR 6461. National Institute of Standards and Technology, Gaithersburg, MD. Etheridge, D., and M. Sandberg. 1996. Building ventilation, theory and measurements. John Wiley & Sons, Chichester, NY. Fan, Y. 1995. CFD modeling of the air and contaminant distribution in rooms. Energy and Buildings 23:33-39. Fanger, P.O. 1972. Thermal comfort: Analysis and application in environmental engineering. McGraw-Hill. Ferziger, J.H. 1977. Large eddy numerical simulations of turbulent flows. AIAA Journal 15(9):1261-1267. Ferziger, J., and M. Peric. 1997. Computational methods for fluid dynamics. Springer, New York. Feustel, H.E. 1998. COMIS—An international air-flow and contaminant transport model. LBNL 42182. Lawrence Berkeley National Laboratory. Feustel, H.E., F. Allard, V.B. Dorer, M. Grosso, M. Herrlin, L. Mingsheng, J.C. Phaff, Y. Utsumi, and H. Yoshino. 1989. The COMIS infiltration model. Proceedings of the 10th AIVC Conference, Air Infiltration and Ventilation Centre. Furbringer J.-M., V. Dorer, F. Huck, and A. Weber. 1993. Air flow simulation of the LESO building including a comparison with measurements and a sensitivity analysis. Proceedings of Indoor Air 1993, p. 5. Gosman, A.D., P.V. Nielsen, A. Restivo, and J.H. Whitelaw. 1980. The flow properties of rooms with small ventilation openings. Transactions of the ASME. Grötzbach, G. 1983. Spatial resolution requirements for direction numerical simulation of the Rayleigh-Bénard convection. Journal of Computational Physics 49:241-264. Haghighat, F., and A.C. Megri. 1996. A comprehensive validation of two airflow models—COMIS and CONTAM. Indoor Air 6:278-288. Haghighat, F., and J. Rao. 1991. Computer-aided building ventilation system design—A system theoretic approach. Energy and Buildings 1:147-155. Herrlin, M.K. 1992. Air-flow studies in multizone buildings—Models and applications. Royal Institute of Technology, Stockholm. Hinze, J.O. 1975. Turbulence, 2nd ed. McGraw-Hill, New York. Horstman, R.H. 1988. Predicting velocity and contamination distribution in ventilated volumes using Navier-Stokes equations. Proceedings of ASHRAE IAQ ’88 Conference, pp. 209-230. Jones, P.J., and G.E. Whittle. 1992. Computational fluid dynamics for building air flow prediction—Current status and capabilities. Building and Environment 27(3):321-338. Kato, S., S. Murakami, and Y. Kondo. 1994. Numerical simulation of twodimensional room airflow with and without buoyancy by means of ASM. ASHRAE Transactions 100(1):238-255. Kato, S., S. Murakami, S. Shoya, F. Hanyu, and J. Zeng. 1995. CFD analysis of flow and temperature fields in atrium with ceiling height of 130 m. ASHRAE Transactions 101(2):1144-1157. Koontz, M.D., H.E. Rector, and N.L. Nagda. 1992. Consumer products exposure guidelines: Evaluation of indoor air quality models. GEOMET Report IE-1980. GEOMET Technologies, Germantown, MD. Lansari, A., J.J. Streicher, A.H. Huber, G.H. Crescenti, R.B. Zweidinger, J.W. Duncan, C.P. Weisel, and R.M. Burton. 1996. Dispersion of automotive alternative fuel vapors within a residence and its attached garage. Indoor Air 6:118-126. Launder, B.E. 1989. Second-moment closure: Present . . . and future? International Journal of Heat and Fluid Flow 10:282-300. Launder, B.E., and D.B. Spalding. 1974. The numerical computation of turbulent flows. Computer Methods in Applied Mechanics and Engineering 3:269-289. Launder, B.E., G.J. Reece, and W. Rodi. 1975. Progress in the development of a Reynolds-stress turbulence closure. Journal of Fluid Mechanics 68(3):537-566. Launder, B.E., and B.I. Sharma. 1974. Application of the energy dissipation model of turbulence to the calculation of flow near a spinning disc. Letters in Heat and Mass Transfer 1(2):131-138.
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Indoor Environmental Modeling Lin, C.H., M.F. Ahlers, A.K. Davenport, L.M. Sedgwick, R.H. Horstman, and J.C. Yu. 2001. A numerical model for airborne disease transmission in a 767-300 passenger cabin. Final report to the National Institute for Occupational Safety and Health, contract no. 200-2000-08001. Boeing Commercial Airplanes Group. Lin, C.H., T. Han, and C.A. Koromilas. 1992. Effect of HVAC design parameters on passenger thermal comfort. SAE Paper No. 920264. Society of Automotive Engineers, Warrendale, PA. Ling, W., J.N. Chung, T.R. Troutt, and C.T. Crowe. 1998. Direct numerical simulation of a three-dimensional temporal mixing layer with particle dispersion. Journal of Fluid Mechanics 358:61-85. Lorenzetti, D.M. 2002. Computational aspects of nodal multizone airflow systems. Building and Environment 37:1083-1090. Mashayek, F. 1998. Direct numerical simulation of evaporating droplet dispersion in forced low Mach number turbulence. International Journal of Heat and Mass Transfer 41(17):2601-2617. Mathieu, J., and J. Scott. 2000. An introduction to turbulent flow. Cambridge University. Monin, A.S., and A.M. Yaglom. 1971. Statistical fluid mechanics, vol. 1. MIT Press, Cambridge. Moser, A., F. Off, A. Schälin, and X. Yuan. 1995. Numerical modeling of heat transfer by radiation and convection in an atrium with thermal inertia. ASHRAE Transactions 101(2):1136-1143. Murakami, S. Kato, and R. Ooka. 1994. Comparison of numerical predictions of horizontal nonisothermal jet in a room with three turbulence models—k-, EVM, ASM, and DSM. ASHRAE Transactions 100(2): 697-706. Nielsen, P.V. 1975. Prediction of air flow and comfort in air conditioned spaces. ASHRAE Transactions 81(2):247-259. Nielsen, P.V. 1992. The description of supply openings in numerical models for room air distribution. ASHRAE Transactions 98(1):963-971. Nielsen, P.V. 1995. Air flow in an exposition pavilion studied by scale-model experiments and computational fluid dynamics. ASHRAE Transactions 101(2):1118-1126. Nielsen. P.V. 1998. The selection of turbulence models for prediction of room airflow. ASHRAE Transactions 104(1B):1119-1127. Nielsen, P.V., and T. Tryggvason. 1998. Computational fluid dynamics and building energy performance simulation. Proceedings of ROOMVENT ’98: Sixth International Conference on Air Distribution in Rooms, Stockholm, 1:101-107. Orme, M. 2000. Applicable input data for a proposed ventilation modeling data guide. ASHRAE Transactions 106(2). Phaff, H.J.C., and W.F. deGids. 1994. The air lock floor. Proceedings of 5th Air Infiltration and Ventilation Centre Conference. Air Infiltration and Ventilation Centre, Brussels, Belgium. Pope, S.B. 2000. Turbulent flows. Cambridge University. Rajaratnam, N. 1976. Turbulent jets. Elsevier, Amsterdam. Reynolds, O. 1895. On the dynamical theory of incompressible viscous fluids and the determination of the criterion. Philosophical Transactions of the Royal Society, London A(186):123. Rodriguez, E.A., and F. Allard. 1992. Coupling COMIS airflow model with other transfer phenomena. Energy and Buildings 18:147-157. Russell, M.B., and P.N. Surendran. 2000. Use of computational fluid dynamics to aid studies of room air distribution: A review of some recent work. Proceedings of CIBSE A: Building Services Engineering Research and Technology 21(4):241-247. Said, M.N., and R.A. MacDonald. 1991. An evaluation of a network smoke control model. ASHRAE Transactions 97(1):275-282. Schälin, A., and P.V. Nielsen. 2004. Impact of turbulence anisotropy near walls in room air flow. Indoor Air 14(3):159-168. dx.doi.org/10.1111 /j.1600-0668.2004.00201e.x. Sextro, R.G., J.M. Daisey, H.E. Feustel, D.J. Dickerhoff, and C. Jump. 1999. Comparison of modeled and measured tracer gas concentrations in a multizone building. Proceedings of Indoor Air ’99, vol. 1. Smagorinsky, J. 1963. General circulation experiments with primitive equations. Monthly Weather Review 91:99-165. Spalart, P.R. 2000. Strategies for turbulence modeling and simulations. International Journal of Heat and Fluid Flow 21:252-263. Srebric, J. 2000. Simplified methodology for indoor environment design. Ph.D. dissertation, Department of Architecture, Massachusetts Institute of Technology, Cambridge.
13.23 Srebric, J., and Q. Chen. 2001. A method of test to obtain diffuser data for CFD modeling of room airflow. ASHRAE Transactions 107(2):108-116. Srebric, J., and Q. Chen. 2002. Simplified numerical models for complex air supply diffusers. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 8(3): 277-294. Tennekes, H., and J. L. Lumley. 1972. A first course in turbulence. MIT Press, Cambridge. Tuomaala, P. 1993. New building air flow simulation model: Theoretical bases. Building Services Engineering Research and Technology 14: 151-157. Upham, R.D. 1997. A validation study of multizone air flow and contaminant migration simulation program CONTAM as applied to tall buildings. M.S. thesis. The Pennsylvania State University, University Park. Versteeg, H., and W. Malalasekera. 1995. An introduction to computational fluid dynamics: The finite volume method. Prentice Hall, Old Tappan, NJ. Walton, G.N. 1989. AIRNET—A computer program for building network airflow modeling. NISTIR 89-4072. National Institute of Standards and Technology, Gaithersburg, MD. Wilcox, D.C. 1998. Turbulence modeling for CFD, 2nd ed. DCW Industries, La Cañada, CA. Williams, P.T., A.J. Baker, and R.M. Kelso. 1994a. Numerical calculation of room air motion—Part 2. ASHRAE Transactions 100(1):531-548. Williams, P.T., A.J. Baker, and R.M. Kelso. 1994b. Numerical calculation of room air motion—Part 3. ASHRAE Transactions 100(1):549-564. Yoshino, H., Z. Yun, H. Kobayashi, and Y. Utsumi. 1995. Simulation and measurement of air infiltration and pollutant transport using a passive solar test house. ASHRAE Transactions 101(1). Yuan, X., A. Moser, and P. Suter. 1994. Wall functions for numerical simulations of turbulent natural convection. Proceedings of the 10th International Heat Transfer Conference, Brighton, U.K., pp. 191-196. Zhai, Z., and Q. Chen. 2003. Solution characters of iterative coupling between energy simulation and CFD programs. Energy and Buildings 35(5):493-505. Zhai, Z., Q. Chen, P. Haves, and J.H. Klems. 2002. On approaches to couple energy simulation and computational fluid dynamics programs. Building and Environment 37:857-864. Zhang, G., S. Morsing, B. Bjerg, K. Svidt, and J.S. Strom. 2000. Test room for validation of airflow patterns estimated by computational fluid dynamics. Journal of Agricultural Engineering Research 76:141-148. Zhao, Y., H. Yoshino, and H. Okuyama. 1998. Evaluation of the COMIS model by comparing simulation and measurement of airflow and pollutant concentration. Indoor Air 8:123-130.
BIBLIOGRAPHY Feustel, H.E., and B.V. Smith. 1997. COMIS 3.0 user’s guide. Lawrence Berkeley National Laboratory. Jiang, Y., D. Alexander, H. Jenkins, R. Arthur, and Q. Chen. 2003. Natural ventilation in buildings: measurement in a wind tunnel and numerical simulation with large eddy simulation. Journal of Wind Engineering and Industrial Aerodynamics 91(3):331-353. Jiang, Y., and Q. Chen. 2001. Study of natural ventilation in buildings by large eddy simulation. Journal of Wind Engineering and Industrial Aerodynamics. 89(13):1155-1178. Jiang, Y., and Q. Chen. 2002. Effect of fluctuating wind direction on cross natural ventilation in building from large eddy simulation. Building and Environment 37(4):379-386. Jiang, Y., and Q. Chen. 2003. Buoyancy-driven single-sided natural ventilation in buildings with large openings. International Journal of Heat and Mass Transfer 46(6):973-988. Jiang, Y., M. Su, and Q. Chen. 2003. Using large eddy simulation to study airflows in and around buildings. ASHRAE Transactions 109(2). Liddament, M., and C. Allen. 1983. The validation and comparison of mathematical models of air infiltration. AIC Technical Note 11. Air Infiltration Centre, Brussels, Belgium. Spalart, P.R., and S.R. Allmaras. 1994. A one-equation turbulence model for aerodynamic flows. La Recherche Aérospatiale 1:5. Su, M., Q. Chen, and C.-M. Chiang. 2001. Comparison of different subgridscale models of large eddy simulation for indoor airflow modeling. Journal of Fluids Engineering 123:628-639.
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Climatic design data tables for 8118 locations in the United States, Canada, and worldwide
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CHAPTER 14
CLIMATIC DESIGN INFORMATION CLIMATIC DESIGN CONDITIONS ....................................... 14.1 CALCULATING CLEAR-SKY SOLAR RADIATION............... 14.8 TRANSPOSITION TO RECEIVING SURFACES OF VARIOUS ORIENTATIONS............................................... 14.10
Generating Design-Day Data ................................................ Estimation of Degree-Days .................................................... Representativeness of Data and Sources of Uncertainty ....... Other Sources of Climatic Information..................................
HIS chapter and the data on the accompanying CD-ROM provide the climatic design information for 8118 locations in the United States, Canada, and around the world. This is an increase of 1675 stations from the 2013 ASHRAE Handbook—Fundamentals. As in the previous edition, the large number of stations made printing the whole tables impractical. Consequently, the complete table of design conditions for only Atlanta, GA, appears in this printed chapter to illustrate the table format. However, a subset of the table elements most often used is presented in the Appendix at the end of this chapter for selected stations representing major urban centers in the United States, Canada, and around the world. The complete data tables for all 8118 stations are contained on the CD-ROM that accompanies this book. This climatic design information is commonly used for design, sizing, distribution, installation, and marketing of heating, ventilating, air-conditioning, and dehumidification equipment, as well as for other energy-related processes in residential, agricultural, commercial, and industrial applications. These summaries include values of dry-bulb, wet-bulb, and dew-point temperature, and wind speed with direction at various frequencies of occurrence. Also included are monthly degree-days to various bases, parameters to calculate clearsky irradiance, and monthly averages of daily all-sky solar radiation. Sources of other climate information of potential interest to ASHRAE members are described later in this chapter. Design information in this chapter was developed largely through research project RP-1699 (Roth 2017). The information includes design values of dry-bulb with mean coincident wet-bulb temperature, design wet-bulb with mean coincident dry-bulb temperature, and design dew-point with mean coincident dry-bulb temperature and corresponding humidity ratio. These data allow the designer to consider various operational peak conditions. Design values of wind speed facilitate the design of smoke management systems in buildings (Lamming and Salmon 1996, 1998). Warm-season temperature and humidity conditions are based on annual percentiles of 0.4, 1.0, and 2.0. Cold-season conditions are based on annual percentiles of 99.6 and 99.0. The use of annual percentiles to define design conditions ensures that they represent the same probability of occurrence in any climate, regardless of the seasonal distribution of extreme temperature and humidity. Monthly precipitation data are also included. They are used mostly to determine climate zones for ASHRAE Standard 169, but may also be helpful in developing green technologies such as vegetative roofs. The clear-sky solar radiation model introduced in the 2009 edition and slightly modified in the 2013 edition is unchanged in its general formulation. However, the site-specific coefficients have been recalculated, based on the latest atmospheric information available. Additionally, all-sky solar radiation values have been added; these are useful in assessing solar technologies (solar heating, photovoltaics), which are typically necessary in the quest for designing net-zero buildings. Design conditions are provided for locations for which long-term hourly observations were available (1990-2014 for most stations in
the United States and Canada). Compared to the 2013 chapter, the number of U.S. stations increased from 1406 to 1952 (39% increase); Canadian stations increased from 562 to 765 (36% increase); and stations in the rest of the world increased from 4475 to 5401 (21% increase; see Figure 1 for map).
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T
The preparation of this chapter is assigned to TC 4.2, Climatic Information.
14.1 Copyright © 2017, ASHRAE
1.
14.12 14.12 14.13 14.16
CLIMATIC DESIGN CONDITIONS
Table 1 shows climatic design conditions for Atlanta, GA, to illustrate the format of the data available on the CD-ROM. A limited subset of these data for 1445 of the 8118 locations for 21 annual data elements is provided for convenience in the Appendix. The top part of the table contains station information as follows: • Name of the observing station, state (USA) or province (Canada), country. • World Meteorological Organization (WMO) station identifier. • Weather Bureau Army Navy (WBAN) number (99999 denotes missing). • Latitude of station, °N/S. • Longitude of station, °E/W. • Elevation of station, m. • Standard pressure at elevation, in kPa (see Chapter 1 for equations used to calculate standard pressure). • Time zone, h ± UTC. • Time zone code (e.g., NAE = Eastern Time, USA and Canada). The CD-ROM contains a list of all time zone codes used in the tables. • Period analyzed (e.g., 90-14 = data from 1990 to 2014 were used).
Annual Design Conditions Annual climatic design conditions are contained in the first three sections following the top part of the table. They contain information as follows: Annual Heating and Humidification Design Conditions. • Coldest month (i.e., month with lowest average dry-bulb temperature; 1 = January, 12 = December). • Dry-bulb temperature corresponding to 99.6 and 99.0% annual cumulative frequency of occurrence (cold conditions), °C. • Dew-point temperature corresponding to 99.6 and 99.0% annual cumulative frequency of occurrence, °C; corresponding humidity ratio, calculated at standard atmospheric pressure at elevation of station, grams of moisture per kg of dry air; mean coincident drybulb temperature, °C. • Wind speed corresponding to 0.4 and 1.0% cumulative frequency of occurrence for coldest month, m/s; mean coincident dry-bulb temperature, °C. • Mean wind speed coincident with 99.6% dry-bulb temperature, m/s; corresponding most frequent wind direction, degrees from north (east = 90°). Annual Cooling, Dehumidification, and Enthalpy Design Conditions. • Hottest month (i.e., month with highest average dry-bulb temperature; 1 = January, 12 = December).
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14.2
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 1 Locations of Weather Stations • Daily temperature range for hottest month, °C [defined as mean of the difference between daily maximum and daily minimum drybulb temperatures for hottest month]. • Dry-bulb temperature corresponding to 0.4, 1.0, and 2.0% annual cumulative frequency of occurrence (warm conditions), °C; mean coincident wet-bulb temperature, °C. • Wet-bulb temperature corresponding to 0.4, 1.0, and 2.0% annual cumulative frequency of occurrence, °C; mean coincident drybulb temperature, °C. • Mean wind speed coincident with 0.4% dry-bulb temperature, m/s; corresponding most frequent wind direction, degrees true from north (east = 90°). • Dew-point temperature corresponding to 0.4, 1.0, and 2.0% annual cumulative frequency of occurrence, °C; corresponding humidity ratio, calculated at the standard atmospheric pressure at elevation of station, grams of moisture per kg of dry air; mean coincident dry-bulb temperature, °C. • Enthalpy corresponding to 0.4, 1.0, and 2.0% annual cumulative frequency of occurrence, kJ/kg; mean coincident dry-bulb temperature, °C. • Extreme maximum wet-bulb temperature, °C. Extreme Annual Design Conditions. • Wind speed corresponding to 1.0, 2.5, and 5.0% annual cumulative frequency of occurrence, m/s. • Mean and standard deviation of extreme annual minimum and maximum dry-bulb temperature, °C. • 5-, 10-, 20-, and 50-year return period values for minimum and maximum extreme dry-bulb temperature, °C. • Mean and standard deviation of extreme annual minimum and maximum wet-bulb temperature, °C. • 5-, 10-, 20-, and 50-year return period values for minimum and maximum extreme wet-bulb temperature, °C.
Monthly Design Conditions Monthly design conditions are divided into subsections as follows:
Temperatures, Degree-Days, and Degree-Hours. • Average temperature, °C. This parameter is a prime indicator of climate and is also useful to calculate heating and cooling degreedays to any base. • Standard deviation of average daily temperature, °C. This parameter is useful to calculate heating and cooling degree-days to any base. Its use is explained in the section on Estimation of DegreeDays. • Heating and cooling degree-days (bases 10 and 18.3°C). These parameters are useful in energy estimating methods. They are also used to classify locations into climate zones in ASHRAE Standard 169. • Cooling degree-hours (bases 23.3 and 26.7°C). These are used in various standards, such as Standard 90.2-2004. Wind. • Monthly average wind speed, m/s. This parameter is useful to estimate the wind potential at a site; however, the local topography may significantly alter this value, so close attention is needed. Precipitation. • Average precipitation, mm. This parameter is used to calculate climate zones for Standard 169, and is of interest in some green building technologies (e.g., vegetative roofs). • Standard deviation of precipitation, mm. This parameter indicates the variability of precipitation at the site. • Minimum and maximum precipitation, mm. These parameters give extremes of precipitation and are useful for green building technologies and stormwater management. Monthly Design Dry-Bulb, Wet-Bulb, and Mean Coincident Temperatures. These values are derived from the same analysis that results in the annual design conditions. The monthly summaries are useful when seasonal variations in solar geometry and intensity, building or facility occupancy, or building use patterns require consideration. In particular, these values can be used when determining
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Climatic Design Information air-conditioning loads during periods of maximum solar radiation. The values listed in the tables include
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• Dry-bulb temperature corresponding to 0.4, 2.0, 5.0, and 10.0% cumulative frequency of occurrence for indicated month, °C; mean coincident wet-bulb temperature, °C. • Wet-bulb temperature corresponding to 0.4, 2.0, 5.0, and 10.0% cumulative frequency of occurrence for indicated month, °C; mean coincident dry-bulb temperature, °C. For a 30-day month, the 0.4, 2.0, 5.0 and 10.0% values of occurrence represent the value that occurs or is exceeded for a total of 3, 14, 36, or 72 h, respectively, per month on average over the period of record. Monthly percentile values of dry- or wet-bulb temperature may be higher or lower than the annual design conditions corresponding to the same nominal percentile, depending on the month and the seasonal distribution of the parameter at that location. Generally, for the hottest or most humid months of the year, the monthly percentile value exceeds the design condition for the same element corresponding to the same nominal percentile. For example, Table 1 shows that the annual 0.4% design dry-bulb temperature at Atlanta, GA, is 34.4°C; the 0.4% monthly dry-bulb temperature exceeds 34.4°C for June, July, and August, with values of 34.7, 36.4, and 36.3°C, respectively. Fifth and tenth percentiles are also provided to give a greater range in the frequency of occurrence, in particular providing less extreme options to select for design calculations. A general, very approximate rule of thumb is that the n% annual cooling design condition is roughly equivalent to the 5n% monthly cooling condition for the hottest month; that is, the 0.4% annual design dry-bulb temperature is roughly equivalent to the 2% monthly design dry-bulb temperature for the hottest month; the 1% annual value is roughly equivalent to the 5% monthly value for the hottest month, and the 2% annual value is roughly equivalent to the 10% monthly value for the hottest month. Mean Daily Temperature Range. These values are useful in calculating daily dry- and wet-bulb temperature profiles, as explained in the section on Generating Design-Day Data. Three kinds of profile are defined: • Mean daily temperature range for month indicated, °C (defined as mean of difference between daily maximum and minimum drybulb temperatures). • Mean daily dry- and wet-bulb temperature ranges coincident with the 5% monthly design dry-bulb temperature. This is the difference between daily maximum and minimum dry- or wet-bulb temperatures, respectively, averaged over all days where the maximum daily dry-bulb temperature exceeds the 5% monthly design dry-bulb temperature. • Mean daily dry- and wet-bulb temperature ranges coincident with the 5% monthly design wet-bulb temperature. This is the difference between daily maximum and minimum dry- or wet-bulb temperatures, respectively, averaged over all days where the maximum daily wet-bulb temperature exceeds the 5% monthly design wet-bulb temperature. Clear-Sky Solar Irradiance. Clear-sky irradiance parameters are useful in calculating solar-related air conditioning loads for any time of any day of the year. Parameters are provided for the 21st day of each month. The 21st of the month is usually a convenient day for solar calculations because June 21 and December 21 represent the solstices (longest and shortest days) and March 21 and September 21 are close to the equinox (days and nights have the same length). Parameters listed in the tables are • Clear-sky optical depths for beam and diffuse irradiances, which are used to calculate beam and diffuse irradiance as explained in the section on Calculating Clear-Sky Solar Radiation.
14.3 • Clear-sky beam normal and diffuse horizontal irradiances at solar noon. These two values can be calculated from the clear-sky optical depths but are listed here for convenience. All-Sky Solar Radiation. All-sky solar radiation parameters are useful for evaluating the potential of solar technologies (e.g., solar heating, photovoltaics), which are valuable in the design of net-zero energy buildings. Parameters listed in the tables are • Monthly average daily global radiation on a horizontal surface. This is a traditional way to characterize the solar resource at a site. • Standard deviation of monthly average daily radiation on a horizontal surface. This parameter gives an idea of the year-to-year variability of the solar resource at the site.
Data Sources The following two primary sources of observational data sets were used in calculating design values: • For most Canadian stations, meteorological data were obtained directly from Environment Canada (climate.weather.gc.ca) for the years 1982-2014. • Data were obtained from the U.S. Climate Reference Network (CRN) (www.ncdc.noaa.gov/crn) (Diamond et al. 2013). • Most stations, including some in Canada with inadequate data, were sourced through the Integrated Surface Database (ISD) from NOAA (www.ncdc.noaa.gov) (Smith et al. 2011) for the years 1982-2015. In most cases, the period of record used in the calculations spanned 25 years (1990 to 2014). This choice of period is a compromise between trying to derive design conditions from the longest possible period of record, and using the most recent data to capture climatic or land-use trends from the past two decades. The actual number of years used in the calculations for a given station depends on the amount of missing data, and, as discussed in the next section, may be as little as 8 years. The first and last years of the period of record used to calculate design conditions are listed in the top section of the tables of climatic design conditions, as shown in Table 1 for Atlanta. For a limited number of stations, years as far back as 1982 or as recent as 2015 were used instead of 1990 to 2014 because that time frame lacked the necessary data. Precipitation data were derived from a number of sources, including station data from the Global Historical Climatology Network, version 2 (GHCN 2015) and the United Nations Food and Agriculture Organization (FAO 2011), as well as gridded data from the Global Precipitation Climatology Centre, version 7 (GPCC 2015), and the Global Precipitation Climate Project (GPCP). Clear-sky solar irradiance parameters listed in the tables constitute a simple parameterization of the more sophisticated REST2 broadband clear-sky radiation model (Gueymard 2008; Gueymard and Thevenard 2009; Thevenard 2009). The REST2 model requires detailed knowledge of various atmospheric constituents, such as aerosols, water vapor, or ozone. To extend applicability of the model to the whole world, multiple data sets, mainly derived from space observations and reanalysis models, were used to obtain these inputs. These sources of data have changed or have been updated since the 2013 edition, which explains the coincident changes in site-specific coefficients. Water vapor, ozone, and ground albedo data are now derived from the National Aeronatucis and Space Administration (NASA) Modern-Era Retrospective Analysis for Research and Applications, version 2 (MERRA-2) reanalysis dataset (Molod et al. 2015), corrected for elevation in the case of water vapor (Gueymard and Thevenard 2009). The period of data is now uniform and longer, from 2000 to 2014. An exception is nitrogen dioxide, for which a database from ozone monitoring instrument
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14.4
2017 ASHRAE Handbook—Fundamentals (SI)
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Table 1 Design Conditions for Atlanta, GA, USA (see Table 1A for Nomenclature)
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Climatic Design Information Table 1A Nomenclature for Tables of Climatic Design Conditions CDDn CDHn DB DBAvg DBSD DP Ebn,noon
Licensed for single user. © 2017 ASHRAE, Inc.
Edh,noon Elev Enth HDDn HR Lat Long MCDB MCDBR MCWB MCWBR MCWS MDBR PCWD Period PrecAvg PrecMax PrecMin PrecStd RadAvg RadStd StdP taub taud Time Zone WB WBAN WMO# WS WSAvg
Cooling degree-days base n°C, °C-day Cooling degree-hours base n°C, °C-hour Dry-bulb temperature, °C Average daily dry-bulb temperature, °C Standard deviation of average daily dry-bulb temperature, °C Dew-point temperature, °C Clear-sky beam normal irradiances at solar noon, W/m2 Clear-sky diffuse horizontal irradiance at solar noon, W/m2 Elevation, m Enthalpy, kJ/kg base 0°C and 101.325 kPa pressure Heating degree-days base, n°C, °C-day Humidity ratio, gmoisture/kgdry air Latitude, °N Longitude, °E Mean coincident dry-bulb temperature, °C Mean coincident dry-bulb temp. range, °C Mean coincident wet-bulb temperature, °C Mean coincident wet-bulb temp. range, °C Mean coincident wind speed, m/s Mean dry-bulb temp. range, °C Prevailing coincident wind direction, ° (0 = North; 90 = East) Years used to calculate the design conditions Average precipitation, mm Maximum precipitation, mm Minimum precipitation, mm Standard deviation of precipitation, mm Monthly mean daily all-sky radiation, kWh/(m2 ·day) Standard deviation of monthly mean daily radiation, kWh/m2 ·day Standard pressure at station elevation, kPa Clear-sky optical depth for beam irradiance Clear-sky optical depth for diffuse irradiance Hours ahead or behind UTC, and time zone code Wet-bulb temperature, °C Weather Bureau Army Navy number Station identifier from the World Meteorological Organization Wind speed, m/s Monthly average wind speed, m/s
Note: Numbers (1) to (45) and letters (a) to (p) are row and column references to quickly point to an element in the table. For example, the 5% design wetbulb temperature for July can be found in row (31), column (k).
(OMI) satellite observations is used over the period 2005 to 2014. Pressure is estimated from station’s elevation. Aerosol turbidity data (in the form of separate evaluations of aerosol optical depth and Ångström exponent) received special attention, because they are the primary inputs that affect the accuracy of direct and diffuse irradiance predictions under clear skies. Spaceborne retrievals of aerosol optical depth at various wavelengths from NASA’s Multi-angle Imaging SpectroRadiometer (MISR; www-misr.jpl.nasa.gov) and two Moderate Resolution Imaging Spectroradiometer (MODIS; modis-atmos.gsfc.nasa.gov) instruments were used between 2000 and 2014 and compared to reference data from a large number of ground-based sites, mostly from the Aerosol Robotic Network (AERONET; aeronet.gsfc.nasa.gov), after appropriate scale-height corrections to remove artifacts from the effect of elevation (Gueymard and Thevenard 2009). Regional corrections of the satellite data were devised to remove as much bias
14.5 as possible, compared to the reference ground-based data. To fill missing data or correct biased satellite observations, modeled aerosol datasets were used, including 10 years (2003 to 2012) of simulated monthly-average aerosol optical depth from the Monitoring Atmospheric Composition and Climate (MACC) reanalysis model (Eskes et al. 2015; Inness et al. 2013) and 13 years (2002 to 2014) of MERRA-2 reanalysis data (Molod et al. 2015). Results from the REST2 model (Guyemard 2008) were then fitted to the simple twoparameter model described in this chapter. The fits enable a concise formulation requiring tabulation, on a monthly basis, of only two parameters per station, referred to here as the clear-sky beam and diffuse optical depths. Details about the fitting procedure can be found in Thevenard and Gueymard (2013). Global horizontal irradiance at the surface, and its standard deviation, were calculated from the Clouds and the Earth’s Radiant Energy System (CERES) Energy Balanced and Filled (EBAF) dataset (ceres.larc.nasa.gov/products.php?product=EBAF-Surface). From the available 1°×1° dataset, a bilinear interpolation, without altitude adjustment, was made given the station latitude and longitude for the period 2000 to 2014.
Calculation of Design Conditions Values of ambient dry-bulb, dew-point, and wet-bulb temperature and wind speed corresponding to the various annual percentiles represent the value that is exceeded on average by the indicated percentage of the total number of hours in a year (8760). The 0.4, 1.0, 2.0, and 5.0% values are exceeded on average 35, 88, 175, and 438 h per year, respectively, for the period of record. The design values occur more frequently than the corresponding nominal percentile in some years and less frequently in others. The 99.0 and 99.6% (cold-season) values are defined in the same way but are usually viewed as the values for which the corresponding weather element is less than the design condition for 88 and 35 h, respectively. Simple design conditions were obtained by binning hourly data into frequency tables, then deriving from the binned data the design condition having the probability of being exceeded a certain percentage of the time. Mean coincident values were obtained by double-binning the hourly data into joint frequency matrices, then calculating the mean coincident value corresponding to the simple design condition. Coincident temperature ranges were also obtained by doublebinning daily temperature ranges (daily maximum minus minimum) versus maximum daily temperature. The mean coincident daily range was then calculated by averaging all bins above the simple design condition of interest. The weather data sets used for the calculations often contain missing values (either isolated records, or because some stations report data only every third hour). Gaps up to 6 h were filled by linear interpolation to provide as complete a time series as possible. Dry-bulb temperature, dew-point temperature, station pressure, and humidity ratio were interpolated. However, wind speed and direction were not interpolated because of their more stochastic and unpredictable nature. Some stations in the ISD data set also provide data that were not recorded at the beginning of the hour. When data at the exact hour were missing, they were replaced by data up to 0.5 h before or after, when available. Finally, psychrometric quantities such as wet-bulb temperature or enthalpy are not contained in the weather data sets. They were calculated from dry-bulb temperature, dew-point temperature, and station pressure using the psychrometric equations in Chapter 1. Measures were taken to ensure that the number and distribution of missing data, both by month and by hour of the day, did not introduce significant biases into the analysis. Annual cumulative frequency distributions were constructed from the relative frequency distributions compiled for each month. Each individual month’s data were
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14.6 included if they met the following screening criteria for completeness and unbiased distribution of missing data after data filling:
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• The number of hourly dry-bulb temperature values for the month, after filling by interpolation, had to be at least 85% of the total hours for the month. • The difference between the number of day and nighttime dry-bulb temperature observations had to be less than 60. Although the nominal period of record selected for this analysis was 25 years (1990 to 2014 for most stations), some variation and gaps in observed data meant that some months’ data were unusable because of incompleteness. Some months were also eliminated during additional quality control checks. A station’s dry-bulb temperature design conditions were calculated only if there were data from at least 8 months that met the quality control and screening criteria from the period of record for each month of the year. For example, there had to be 8 months each of January, February, March, etc. for which data met the completeness screening criteria. These criteria were ascertained from results of RP-1171 (Hubbard et al. 2004) and were the same as used in calculating the design conditions in the 2001 to 2013 editions of the ASHRAE Handbook—Fundamentals. Dew-point temperature, wet-bulb temperature, and enthalpy design conditions were calculated for a given month only if the number of dew-point, wet-bulb, or enthalpy values was greater than 85% of the minimum number of dry-bulb temperature values defined previously; wind speed and direction conditions were calculated for a given month only if the number of values was greater than 28.3% (i.e., one-third of 85%) the minimum number of dry-bulb temperature values. For example, a month of January was included in calculations if the number of dry-bulb temperature values exceeded 85% of 744 h, or 633 h. The month was included in calculation of dewpoint temperature design conditions only if dew-point temperature was present for at least 85% of 633 h, or 538 h. The month was included in calculation of wind speed design conditions only if wind speed was present for at least 28.3% of 633 h, or 179 h. Annual dry-bulb temperature extremes were calculated only for years that were 85% complete. At least 8 annual extremes were required to calculate the mean and standard deviation of extreme annual dry-bulb temperatures. Daily minimum and maximum temperatures were calculated only for complete days; so were daily temperature ranges and mean coincident temperature ranges. Details about quality checks and other steps taken during data processing to ensure results as free from error as possible are detailed in Roth (2017).
Differences from Previously Published Design Conditions • Climatic design conditions in this chapter are generally similar to those in previous editions, because similar if not identical analysis procedures were used. There are some differences, however, owing to a more recent period of record (generally 1990-2014 versus 1982-2006). For example, when compared to the 2009 edition, 99.6% heating dry-bulb temperatures have increased by 0.05 K on average, and 0.4% cooling dry-bulb temperatures have increased by 0.08 K on average. Similar trends are observed for other design temperatures. The root mean square differences are 0.63 K for the 99.6% heating dry-bulb values and 0.36 K for 0.4% cooling dry-bulb. The increases noted here are generally consistent with the discussion in the section on Effects of Climate Change. • Further details concerning differences between design conditions in the 2013, 2009, and 2005 editions are described in Thevenard (2009) and Thevenard and Gueymard (2013). Differences between the 2005 and the 2001 editions are described in Thevenard
2017 ASHRAE Handbook—Fundamentals (SI) et al. (2005). Differences between the 1993 and previous editions are described in Colliver et al. (2000).
Applicability and Characteristics of Design Conditions Climatic design values in this chapter represent different psychrometric conditions. Design data based on dry-bulb temperature represent peak occurrences of the sensible component of ambient outdoor conditions. Design values based on wet-bulb temperature are related to the enthalpy of the outdoor air. Conditions based on dew point relate to the peaks of the humidity ratio. The designer, engineer, or other user must decide which set(s) of conditions and probability of occurrence apply to the design situation under consideration. Additional sources of information on frequency and duration of extremes of temperature and humidity are provided in the section on Other Sources of Climatic Information. Further information is available from Harriman et al. (1999). This section discusses the intended use of design conditions in the order they appear in Table 1. Annual Heating and Humidification Design Conditions. The month with the lowest mean dry-bulb temperature is used, for example, to determine the time of year where the maximum heating load occurs. The 99.6 and 99.0% design conditions are often used in sizing heating equipment. The humidification dew-point and mean coincident dry-bulb temperatures and humidity ratio provide information for cold-season humidification applications. Wind design data provide information for estimating peak loads accounting for infiltration: extreme wind speeds for the coldest month, with the mean coincident dry-bulb temperature; and mean wind speed and direction coincident to the 99.6% design dry-bulb temperature. Annual Cooling, Dehumidification, and Enthalpy Design Conditions. The month with the highest mean dry-bulb temperature is used, for example, to determine the time of year where the maximum sensible cooling load occurs, not taking into account solar loads. The mean daily dry-bulb temperature range for the hottest month is the mean difference between the daily maximum and minimum temperatures during the hottest month and is calculated from the extremes of the hourly temperature observations. The true maximum and minimum temperatures for any day generally occur between hourly readings. Thus, the mean maximum and minimum temperatures calculated in this way are about 0.5 K less extreme than the mean daily extreme temperatures observed with maximum and minimum thermometers. This results in the true daily temperature range generally about 1 K greater than that calculated from hourly data. The mean daily dry-bulb temperature range is used in cooling load calculations. The 0.4, 1.0, and 2.0% dry-bulb temperatures and mean coincident wet-bulb temperatures often represent conditions on hot, mostly sunny days. These are often used in sizing cooling equipment such as chillers or air-conditioning units. Design conditions based on wet-bulb temperature represent extremes of the total sensible plus latent heat of outdoor air. This information is useful for design of cooling towers, evaporative coolers, and outdoor-air ventilation systems. The mean wind speed and direction coincident with the 0.4% design dry-bulb temperature is used for estimating peak loads accounting for infiltration. Design conditions based on dew-point temperatures are directly related to extremes of humidity ratio, which represent peak moisture loads from the weather. Extreme dew-point conditions may occur on days with moderate dry-bulb temperatures, resulting in high relative humidity. These values are especially useful for humidity control applications, such as desiccant cooling and dehumidification, coolingbased dehumidification, and outdoor-air ventilation systems. The
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Climatic Design Information
14.7
values are also used as a check point when analyzing the behavior of cooling systems at part-load conditions, particularly when such systems are used for humidity control as a secondary function. Humidity ratio values are calculated from the corresponding dew-point temperature and the standard pressure at the location’s elevation. Annual enthalpy design conditions give the annual enthalpy for the cooling season; this is used for calculating cooling loads caused by infiltration and/or ventilation into buildings. Enthalpy represents the total heat content of air (the sum of its sensible and latent energies). Cooling loads can be calculated knowing the conditions of both the outdoor ambient and the building’s interior air. The extreme maximum wet-bulb temperature provides the highest wet-bulb temperature observed over the entire period of record and is the most extreme condition observed during the data record for evaporative processes such as cooling towers. For most locations, the extreme maximum wet-bulb value is significantly higher than the 0.4% wet-bulb (discussed previously) and should be used only for design of critical applications where an occasional short-duration capacity shortfall is not acceptable. Extreme Annual Design Conditions. Extreme annual design wind speeds are used in designing smoke management systems. The mean and standard deviation of the extreme annual maximum and minimum dry-bulb temperatures are used to calculate the probability of occurrence of very extreme conditions. These can be required for design of equipment to ensure continuous operation and serviceability regardless of whether the heating or cooling loads are being met. These values were calculated from extremes of hourly temperature observations. The true maximum and minimum temperatures for any day generally occur between hourly readings. Thus, the mean maximum and minimum temperatures calculated in this way are about 0.5 K less extreme than the mean daily extreme temperatures observed with maximum and minimum thermometers. The 5-, 10-, 20- and 50-year return periods for maximum and minimum extreme dry-bulb temperature are also listed in the table. Return period (or recurrence interval) is defined as the reciprocal of the annual probability of occurrence. For instance, the 50-year return period maximum dry-bulb temperature has a probability of occurring or being exceeded of 2.0% (i.e., 1/50) each year. This statistic does not indicate how often the condition will occur in terms of the number of hours each year (as in the design conditions based on percentiles) but describes the probability of the condition occurring at all in any year. The following method can be used to estimate the return period (recurrence interval) of extreme temperatures: Tn = M + IFs
(1)
where Tn = n-year return period value of extreme dry-bulb temperature to be estimated, years M = mean of annual extreme maximum or minimum dry-bulb temperatures, °C s = standard deviation of annual extreme maximum or minimum dry-bulb temperatures, K I = 1 if maximum dry-bulb temperatures are being considered = –1 if minimum dry-bulb temperatures are being considered 6 n F = – -------- 0.5772 + ln ln ------------ n – 1
For example, the 50-year return period extreme maximum dry-bulb temperature estimated for Atlanta, GA, is 41.2°C (according to Table 1, M = 35.9°C, s = 2.0, and n = 50; I = 1). Similarly, the 50year return period extreme minimum dry-bulb temperature for Atlanta, GA, is –16.1°C [M = –9.5°C, s = 2.6, and n = 50; I = –1]. The n-year return periods can be obtained for most stations using ASHRAE’s Weather Data Viewer 6.0 (ASHRAE 2017), which is discussed in the section on Other Sources of Climatic Information.
New in 2017 are the parameters required to calculate the 5-, 10-, 20- and 50-year return periods for maximum and minimum extreme wet-bulb temperature. The maximum conditions in particular may be useful in determining very extreme wet-bulb temperatures during which evaporative systems may have to operate. Calculation of the n-year return period is based on assumptions that annual maxima and minima are distributed according to the Gumbel (Type 1 Extreme Value) distribution and are fitted with the method of moments (Lowery and Nash 1970). The uncertainty or standard error using this method increases with standard deviation, value of return period, and decreasing length of the period of record. It can be significant. For instance, the standard error in the 50-year return period maximum dry-bulb temperature estimated at a location with a 12-year period of record can be 3 K or more. Thus, the uncertainties of return period values estimated in this way are greater for stations with fewer years of data than for stations with the complete period of record from 1990 to 2014. Temperatures, Degree-Days, and Degree-Hours. Monthly average temperatures and standard deviation of daily average temperatures are calculated using the averages of the minimum and maximum temperatures for each complete day within the period analyzed. They are used to estimate heating and cooling degree-days to any base, as explained in the section on Estimation of DegreeDays. Heating and cooling degree-days (base 10 or 18.3°C) are calculated as the sum of the differences between daily average temperatures and the base temperature. For example the number of heating degree-days (HDD) in the month is calculated as HDD =
Tbase – Ti N
+
(2)
i=1
where N is the number of days in the month, Tbase is the reference temperature to which the degree-days are calculated, and T i is the mean daily temperature calculated by adding the maximum and minimum temperatures for the day, then dividing by 2. The + superscript indicates that only positive values of the bracketed quantity are taken into account in the sum. Similarly, monthly cooling degree-days (CDD) are calculated as CDD =
Ti – Tbase N
+
(3)
i=1
Degree-days are used in energy estimating methods, and to classify stations into climate zones for ASHRAE Standard 169. Monthly Design Dry-Bulb and Mean Coincident Wet-Bulb Temperatures. These values provide design conditions for processes driven by dry-bulb air temperature. In particular, air-conditioning cooling loads are generally based on dry-bulb design conditions (plus clear-sky solar irradiance). Monthly Design Wet-Bulb and Mean Coincident Dry-Bulb Temperatures. Wet-bulb design conditions are of use in analysis of evaporative coolers, cooling towers, and other equipment involving evaporative transfer. Note also that air wet-bulb temperature and enthalpy are closely related, so applications with large ventilation flow rates may have maximum cooling requirements under high wet-bulb conditions. Mean Daily Temperature Range. Mean daily range values are computed using all days of the month, as opposed to coincident values that derive from design days. Mean daily range values have been published in previous Handbook editions and are included for completeness. Coincident daily range values should be used for generating design-day profiles. Clear-Sky Solar Irradiance. Clear-sky solar irradiance data are used in load calculation methods. Beam normal irradiance refers to solar radiation emanating directly from the solar disk and measured
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14.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 2
Month
Licensed for single user. © 2017 ASHRAE, Inc.
Day of year Eo, W/m2 Equation of time (ET), min Declination , degrees
Approximate Astronomical Data for 21st Day of Each Month
Jan
Feb
Mar
Apr
May
Jun
Jul
Aug
Sep
Oct
Nov
Dec
21 1410 –10.6 –20.1
52 1397 –14.0 –11.2
80 1378 –7.9 –0.4
111 1354 1.2 11.6
141 1334 3.7 20.1
172 1323 –1.3 23.4
202 1324 –6.4 20.4
233 1336 –3.6 11.8
264 1357 6.9 –0.2
294 1380 15.5 –11.8
325 1400 13.8 –20.4
355 1411 2.2 –23.4
perpendicularly to the rays of the sun. Diffuse horizontal irradiance refers to solar radiation emanating from the sky dome, sun excluded, and measured on a horizontal surface. Because the beam and diffuse irradiances vary during the course of the day, current load calculation methods require their estimation at various times, which can be done with the method described in the section on Calculating Clear-Sky Solar Radiation. The method uses the clear-sky optical depths b and d , listed in Table 1 as taub and taud, respectively, as inputs. Clear-sky beam normal and diffuse horizontal irradiances at solar noon are also listed in Table 1 for convenience. All-Sky Solar Radiation. All-sky solar radiation data are used in the design of solar energy systems (either thermal or photovoltaic). Monthly average daily radiation on the horizontal refers to average amount of solar radiation received on a horizontal surface during the course of a day, for the month under consideration. The standard deviation of monthly average daily radiation on the horizontal is the standard deviation of the previous monthly quantity, calculated over the period of record used for the Handbook, and is an indicator of the year-to-year variability of solar radiation.
2.
Table 3
Local Standard TZ Meridian Longitude (Hours ± UTC) (°E)
Time Zone Name Newfoundland standard time Atlantic standard time Eastern standard time Central standard time Mountain standard time Pacific standard time Alaska standard time Hawaii-Aleutian standard time
–3.5 –4 –5 –6 –7 –8 –9 –10
–52.5 –60 –75 –90 –105 –120 –135 –150
Equation of Time and Solar Time The earth’s orbital velocity also varies throughout the year, so apparent solar time (AST), as determined by a solar time sundial, varies somewhat from the mean time kept by a clock running at a uniform rate. This variation is called the equation of time (ET) and is approximated by the following formula (Iqbal 1983): ET = 2.2918[0.0075 + 0.1868 cos() – 3.2077 sin() – 1.4615 cos(2) – 4.089 sin(2)]
CALCULATING CLEAR-SKY SOLAR RADIATION
Knowledge of clear-sky solar radiation at various times of year and day is required by several calculation methods for heat gains in HVAC loads and solar energy applications. The tables of climatic design conditions include the parameters required to calculate clearsky beam and diffuse solar irradiances using the equations in the following section. The section on Transposition to Receiving Surfaces of Various Orientations explains how to use these values to calculate clear-sky solar radiation incident on arbitrary surfaces. Note that in all equations in this section, angles are expressed in degrees. This includes the arguments appearing in trigonometric functions.
n–1 = 360° -----------365
where n is the day of year (1 for January 1, 32 for February 1, etc.) and the argument inside the cosine is in degrees. Table 2 tabulates values of Eo for the 21st day of each month.
(7)
where AST LST ET LSM
= = = =
apparent solar time, decimal hours local standard time, decimal hours equation of time in minutes, from Table 2 or Equation (5) longitude of local standard time meridian, °E of Greenwich (negative in western hemisphere) LON = longitude of site, °E of Greenwich
Most standard meridians are found every 15° from 0° at Greenwich, U.K., with a few exceptions, such as the province of Newfoundland in Canada. Standard meridian longitude is related to time zone as follows: LSM = 15TZ
(4)
(6)
Table 2 tabulates the values of ET for the 21st day of each month. The conversion between local standard time and solar time involves two steps: the equation of time is added to the local standard time, and then a longitude correction is added. This longitude correction is four minutes of time per degree difference between the local (site) longitude and the longitude of the local standard meridian (LSM) for that time zone; hence, AST is related to the local standard time (LST) as follows: AST = LST + ET/60 + (LON – LSM)/15
The solar constant Esc is defined as the intensity of solar radiation on a surface normal to the sun’s rays, just beyond the earth’s atmosphere, at the average earth-sun distance. One frequently used value is that proposed by the World Meteorological Organization in 1981, Esc = 1367 W/m2 (Iqbal 1983). Because the earth’s orbit is slightly elliptical, the extraterrestrial radiant flux Eo varies throughout the year, reaching a maximum of 1412 W/m2 near the beginning of January, when the earth is closest to the sun (aphelion) and a minimum of 1322 W/m2 near the beginning of July, when the earth is farthest from the sun (perihelion). Extraterrestrial solar irradiance incident on a surface normal to the sun’s ray can be approximated with the following equation:
(5)
with ET expressed in minutes and
Solar Constant and Extraterrestrial Solar Radiation
n – 3 Eo = Esc 1 + 0.033 cos 360 ---------------- 365
Time Zones in United States and Canada
(8)
where TZ is the time zone, expressed in hours ahead or behind coordinated universal time (UTC). TZ is listed for each station on the CD-ROM accompanying this book. Table 3 lists time zones and standard time meridians for the United States and Canada. If daylight saving time (DST) is to be used, rather than local standard time, an additional correction has to be performed. In most
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Climatic Design Information
14.9
locales, local standard time can be obtained from daylight savings time by subtracting one hour: LST = DST – 1
(9)
where DST is in decimal hours.
Declination Because the earth’s equatorial plane is tilted at an angle of 23.45° to the orbital plane, the solar declination (the angle between the earth/sun line and the equatorial plane) varies throughout the year, as shown in Figure 2. This variation causes the changing seasons with their unequal periods of daylight and darkness. Declination can be obtained from astronomical or nautical almanacs; however, for most engineering applications, the following equation provides sufficient accuracy: n + 284 = 23.45 sin 360 ------------------ 365
(10)
where is in degrees and the argument inside the sine is also in degrees. Table 2 provides for the 21st day of each month.
Sun Position The sun’s position in the sky is conveniently expressed in terms of the solar altitude above the horizontal and the solar azimuth measured from the south (Figure 3). The solar altitude angle is defined as the angle between the horizontal plane and a line emanating from the sun. Its value ranges from 0° when the sun is on the horizon, to 90° if the sun is directly overhead. Negative values correspond to night times. The solar azimuth angle is defined as angular displacement from south of the projection, on the horizontal plane, of the earth/sun line. By convention, it is counted positive for afternoon hours and negative for morning hours. Solar altitude and azimuth angles, in turn, depend on the local latitude L (N, negative in the southern hemisphere); the solar declination , which is a function of the date [see Table 2 or Equation (10)]; and the hour angle H, defined as the angular displacement of the sun east or west of the local meridian caused by the rotation of the earth, and expressed in degrees as
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H = 15(AST – 12)
(11)
where AST is the apparent solar time [Equation (7)]. H is zero at solar noon, positive in the afternoon, and negative in the morning. Equation (12) relates the solar altitude angle to L, , and H: sin = cos L cos cos H + sin L sin
(12)
Note that at solar noon, H = 0 and the sun reaches its maximum altitude in the sky: max = 90° – |L – |
(13)
The azimuth angle is uniquely determined by its sine and cosine, given in Equations (14) and (15):
Fig. 2 Motion of Earth around Sun
sin = sin H cos /cos
(14)
cos = (cos H cos sin L – sin cos L)/cos
(15)
Fig. 3 Solar Angles for Vertical and Horizontal Surfaces
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14.10
2017 ASHRAE Handbook—Fundamentals (SI)
Example 1. Calculate the position of the sun in Atlanta, GA, for July 21 at noon solar time. Solution: From Table 1, Atlanta is at latitude L = 33.64°N. From Table 2 or Equation (10), declination = 20.44°. Solar altitude is given by Equation (13): = 90 – |33.64 – 20.44| = 76.80° At solar noon, the sun is due south, so the azimuth angle is simply 0°. Example 2. Perform the same calculation as in Example 1, but for 3:00 PM eastern daylight saving time. Solution: Compared to Example 1, a few extra steps are required to calculate AST. From Table 1, for Atlanta, LON = 84.43°W = –84.43°E and TZ = –5.00. Also, from Table 1 or Equation (5), ET = –6.4 min. Then, from Equation (8): LSM = 15(–5.00) = –75° Because 3 PM daylight saving time is 2 14, Equation (7) leads to
PM
standard time, or hour
AST = 14 – 6.4/60 + [(–84.43) – (–75)]/15 = 13.27 h Then, from Equation (11):
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H = 15(13.27 – 12) = 18.97° Solar altitude is given by Equation (12), using the same latitude and declination as in Example 1: sin = cos(33.64°) cos(20.44°) cos(18.97°) + sin(33.64°) sin(20.44°) = 0.931
Therefore, = 56.69°.
The relative air mass m is the ratio of the mass of atmosphere in the actual earth/sun path to the mass that would exist if the sun were directly overhead. Air mass is solely a function of solar altitude and is obtained from (Kasten and Young 1989) m = 1/[sin + 0.50572(6.07995 +
)–1.6364]
(16)
where is expressed in degrees.
Equations (17) to (20) describe a simple parameterization of a sophisticated broadband radiation model and provide accurate predictions of Eb and Ed , even at sites where the atmosphere is very hazy or humid most of the time. Example 3. Calculate clear-sky beam and diffuse solar irradiance in Atlanta, GA, for July 21 at noon solar time. Note that Table 1 already lists clear-sky beam and diffuse solar irradiance for solar noon. Calculations are shown here to illustrate the application of the method.
ad = 0.507 + 0.205 × 0.515 – 0.080 × 2.066 – 0.190 × 0.515 × 2.066 = 0.245 and from Equations (17) and (18), Eb = 1324 exp(–0.515 × 1.0270.714) = 784 W/m2 Ed = 1324 exp(–2.066 × 1.0270.245) = 166 W/m2 These are the values listed for Ebn,noon and Edh,noon in Table 1.
Clear-Sky Solar Radiation Solar radiation on a clear day is defined by its beam (direct) and diffuse components. The direct component represents the part of solar radiation emanating directly from the solar disc, whereas the diffuse component accounts for radiation emanating from the rest of the sky. These two components are calculated as
Example 4. Perform the same calculation as in Example 3, but for 3 eastern daylight saving time.
m = 1/[sin(68.62°) + 0.50572(6.07995 + 68.62)–1.6364] = 1.073
(17)
Eb = 1324 exp(–0.515 × 1.0730.714) = 770 W/m2
Ed = Eo exp[–d mad]
(18)
Ed = 1324 exp(–2.066 × 1.0730.245) = 162 W/m2
Eb = beam normal irradiance (measured perpendicularly to rays of
the sun) diffuse horizontal irradiance (measured on horizontal surface) extraterrestrial normal irradiance [Equation (4) or Table 2] air mass [Equation (16)] beam and diffuse optical depths (b and d are more correctly termed pseudo-optical depths, because optical depth refers to an air mass coefficient without exponentiation; “optical depth” is used here for convenience.) beam and diffuse air mass exponents
PM
Solution: This is the same calculation as in the solution of Example 3, but using the solar altitude = 68.62° calculated in Example 2 (ab and ad are unchanged from Example 3):
mab]
where
ab and ad =
(20)
ab = 1.454 – 0.406 × 0.515 – 0.268 × 2.066 + 0.021 × 0.515 × 2.066 = 0.714
Air Mass
m =
ad = 0.507 + 0.205 b – 0.080 d – 0.190 bd
From Table 1, the beam and diffuse optical depths for Atlanta in July are b = 0.515 and d = 2.066. From Table 2 or Equation (4), normal extraterrestrial irradiance on July 21 is Eo = 1324 W/m2. Then, from Equations (19) and (20)
cos = [cos(18.97°) cos(20.44°) sin (33.64°) – sin(20.44°) cos(33.64°)]/cos(68.62°) = 0.549
b and d =
(19)
m = 1/[sin(76.80°) + 0.50572(6.07995 + 76.80)–1.6364] = 1.027
sin = sin(18.97°) cos(20.44°)/cos(68.62°) = 0.836
Ed = Eo =
ab = 1.454 – 0.406 b – 0.268 d + 0.021 bd
Solution: From Example 1, at solar noon on July 21 in Atlanta solar altitude is = 76.80°. From Equation (16):
Therefore, = 68.62°. Solar azimuth is obtained through Equations (14) and (15):
Eb = Eo exp[–b
Values of b and d are location-specific, and vary during the year. They embody the dependence of clear-sky solar radiation on local conditions, such as elevation, precipitable water, aerosols, ozone, and surface reflectance. In previous editions, their average values were determined through ASHRAE research projects RP1453 (Thevenard 2009) and RP-1613 (Thevenard and Gueymard 2013). For this edition, the results are from RP-1699 (Roth 2017), and are tabulated for the 21st day of each month for all the locations in the tables of climatic design conditions. Values for other days of the year should be found by interpolation. Air mass exponents ab and ad are correlated to b and d through the following empirical relationships:
3. TRANSPOSITION TO RECEIVING SURFACES OF VARIOUS ORIENTATIONS Calculations developed in the previous section are chiefly concerned with estimating clear-sky solar irradiance either normal to the rays of the sun (direct beam) or on a horizontal surface (diffuse). However, in many circumstances, calculation of clear-sky solar irradiance is required on surfaces of arbitrary orientations. Receiving
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Climatic Design Information Table 4 Orientation
14.11 which leads to = 8.74°.
Surface Orientations and Azimuths, Measured from South N
NE
Surface azimuth 180°
E
SE
–135° –90° –45°
S
SW
W
NW
0
45°
90°
135°
surfaces can be vertical (e.g., walls and windows) or tilted (e.g., skylights or active solar devices). This section describes transposition models that enable calculating solar irradiance on any surface, knowing beam normal and diffuse horizontal irradiance.
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Solar Angles Related to Receiving Surfaces The orientation of a receiving surface is best characterized by its tilt angle and its azimuth, shown in Figure 3. The tilt angle (also called slope) is the angle between the surface and the horizontal plane. Its value lies between 0 and 180°. Most often, slopes are between 0° (horizontal) and 90° (vertical). Values above 90° correspond to surfaces facing the ground. The surface azimuth is defined as the displacement from south of the projection, on the horizontal plane, of the normal to the surface. Surfaces that face west have a positive surface azimuth; those that face east have a negative surface azimuth. Surface azimuths for common orientations are summarized in Table 4. Note that, in this chapter, surface azimuth is defined as relative to south in both the northern and southern hemispheres. Other presentations and software use relative-to-north or relative-to-equator; care is required. The surface-solar azimuth angle is defined as the angular difference between the solar azimuth and the surface azimuth : =–
(21)
Values of greater than 90° or less than –90° indicate that the surface is in the shade. Finally, the angle between the line normal to the irradiated surface and the earth-sun line is called the angle of incidence . It is important in fenestration, load calculations, and solar technology because it affects the intensity of the direct component of solar radiation striking the surface and the surface’s ability to absorb, transmit, or reflect the sun’s rays. Its value is given by cos = cos cos sin + sin cos
(22)
Note that for vertical surfaces ( = 90°) Equation (22) simplifies to cos = cos cos
(23)
whereas for horizontal surfaces ( = 0°) it simplifies to = 90 –
(24)
Example 5. For Atlanta, GA, on July 21 at 3 PM eastern daylight saving time, find the angle of incidence at a vertical widow facing 60° west of south. Solution: The azimuth of the receiving surface is = +60°. According to Example 2, solar azimuth angle is = 56.69°. Then, Equation (21) gives the surface-solar azimuth angle as = 56.69° – 60° = –3.31° Still from Example 2, solar altitude angle is = 68.62°. Equation (23) leads to cos = cos(68.62°) cos(–3.31°) = 0.364 Therefore, = 68.66°. Example 6. For the same conditions as in Example 5, find the angle of incidence at a skylight tilted at 30° and facing 60° west of south. Solution: The azimuth of the receiving surface is still = +60°, but its slope is = 30°. Other angles are unchanged from Example 5. Equation (22) now applies: cos = cos(68.62°) cos(–3.31°) sin(30°) + sin(68.62°) cos(30°) = 0.988
Calculation of Clear-Sky Solar Irradiance Incident On Receiving Surface Total clear-sky irradiance Et reaching the receiving surface is the sum of three components: the beam component Et,b originating from the solar disc; the diffuse component Et,d, originating from the sky dome; and the ground-reflected component Et,r originating from the ground in front of the receiving surface. Thus, Et = Et,b + Et,d + Et,r
(25)
Only a simple method for computing all the factors on the right side of Equation (25) is presented here. More elaborate methods, particularly with regard to the calculating the diffuse component, can be found in Gueymard (1987) and Perez et al. (1990). Beam Component. The beam component is obtained from a straightforward geometric relationship: Et,b = Eb cos
(26)
where is the angle of incidence. This relationship is valid only when cos > 0; otherwise, Et,b = 0. Diffuse Component. The diffuse component is more difficult to estimate because of the anisotropic nature of diffuse radiation: some parts of the sky, such as the circumsolar disc or the horizon, tend to be brighter than the rest of the sky, which makes the development of a simplified model challenging. For vertical surfaces, Stephenson (1965) and Threlkeld (1963) showed that the ratio Y of clear-sky diffuse irradiance on a vertical surface to clear-sky diffuse irradiance on the horizontal is a simple function of the angle of incidence : Et,d = EdY
(27)
Y = max(0.45, 0.55 + 0.437 cos + 0.313 cos2 )
(28)
with For a nonvertical surface with slope , the following simplified relationships are sufficient for most applications described in this volume: Et,d = Ed (Y sin + cos ) Et,d = Ed Y sin
if 90° if 90°
(29) (30)
where Y is calculated for a vertical surface having the same azimuth as the receiving surface considered. Note that Equations (27) to (30) are appropriate for clear-sky conditions, but should not be used for cloudy skies. Ground-Reflected Component. Ground-reflected irradiance for surfaces of all orientations is given by 1 + cos Et,r = (Eb sin + Ed)g --------------------2
(31)
where g is ground reflectance, often taken to be 0.2 for a typical mixture of ground surfaces. Table 5 provides estimates of g for other surfaces, including in the presence of snow. Example 7. Find the direct, diffuse and ground-reflected components of clear-sky solar irradiance on the window in Example 5. Solution: Clear-sky beam normal irradiance Eb and diffuse horizontal irradiance Ed were calculated in Example 4 as Eb = 770 W/m2 and Ed = 162 W/m2. Example 2 provided the solar altitude as = 68.62° and Example 5 provided the angle of incidence as = 68.66°. The surface slope is = 90°, and ground reflectance is assumed to be 0.2. Substituting these values into Equations (26), (27), (28), and (31) leads to Et,b = 770 cos(68.66°) = 280 W/m2
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14.12
2017 ASHRAE Handbook—Fundamentals (SI)
Y = max[0.45, 0.55 + 0.437 cos(68.66°) + 0.313 cos2(68.66°)] = 0.750 Et,d = 162 × 0.750 = 121 W/m2 1 + cos 68.62 Et,r = [770 sin (68.62°) + 162]0.2 --------------------------------------- = 120 W/m2 2 Example 8. Find the direct, diffuse and ground-reflected components of clear-sky solar irradiance on the skylight in Example 6. Solution: This example uses the same values as Example 7, except that the surface slope is = 30° and the angle of incidence, calculated in Example 6, is = 8.74°. The clear-sky irradiance components are then calculated from Equations (26), (29) and (31); the ratio Y is calculated for a vertical surface having the same azimuth as the receiving surface, so the value calculated in Example 7 is unchanged. Et,b = 770 cos(8.74°) = 761 W/m2 Et,d = 162[0.750 sin(30°) + cos(30°)] = 201 W/m2
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1 – cos 68.62 Et,r = [770 sin (68.62°) + 162]0.2 --------------------------------------- = 120 W/m2 2
3.1
GENERATING DESIGN-DAY DATA
This section provides procedures for generating 24 h temperature data sequences suitable as input to many HVAC analysis methods, including the radiant time series (RTS) cooling load calculation procedure described in Chapter 18. Temperatures. Table 6 gives a normalized daily temperature profile in fractions of daily temperature range. Recent research projects RP-1363 (Hedrick 2009) and RP-1453 (Thevenard 2009) have shown that this profile is representative of both dry-bulb and wetbulb temperature variation on typical design days. To calculate hourly temperatures, subtract the Table 6 fraction of the dry- or wet-bulb daily range from the dry- or wet-bulb design temperature (limiting by saturation in the case of the wet-bulb). This procedure is applicable to annual or monthly data and is shown in Example 9. Table 7 specifies the input values to be used for generating several design-day types. Because daily temperature variation is driven by heat from the sun, the profile in Table 6 is, strictly speaking, specified in terms of solar time. Typical HVAC calculations (e.g., hourly cooling loads)
are performed in local time, reflecting building operation schedules. The difference between local and solar time can easily be 1 or 2 h, depending on site longitude and whether daylight saving time is in effect. This difference can be included by accessing the temperature profile using apparent solar time (AST) calculated with Equation (7), as shown in Example 9. Additional Moist-Air Properties. Once hourly dry-bulb and wet-bulb temperatures are known, additional moist air properties (e.g., dew-point temperature, humidity ratio, enthalpy) can be derived using the psychrometric chart, equations in Chapter 1, or psychrometric software. Example 9. Deriving Hourly Design-Day Temperatures. Calculate hourly temperatures for Atlanta, GA, for a July dry-bulb design day using the 5% design conditions. Solution: From Table 1, the July 5% dry-bulb design conditions for Atlanta are DB = 33.1°C and MCWB = 23.5°C. Daily range values are MCDBR = 11.2°C and MCWBR = 3.4°C. Daylight saving time is in effect for Atlanta in July. Apparent solar time (AST) for hour 1 local daylight saving time (LDT) is –0.73. The nearest hour to the AST is 23, yielding a Table 6 profile value of 0.75. Then tdb,1 = 33.1 – 0.75 11.2 = 24.7°C. Similarly, twb,1 = 23.5 – 0.75 3.4 = 21.0°C. With psychrometric formulas, derive tdp,1 = 19.3°C. Table 8 shows results of this procedure for all 24 h.
3.2
ESTIMATION OF DEGREE-DAYS
Monthly Degree-Days The tables of climatic design conditions in this chapter list heating and cooling degree-days (bases 10 and 18.3°C). Although 10 and 18.3°C represent the most commonly used bases for the calculation of degree-days, calculation to other bases may be necessary. With that goal in mind, the tables also provide two parameters (monthly average temperature T, and standard deviation of daily average temperature sd) that enable estimation of degree-days to any base with reasonable accuracy. The calculation method was established by Schoenau and Kehrig (1990). Heating degree days HDDb to base Tb are expressed as HDDb = Nsd [ZbF(Zb) + f (Zb)]
where N is the number of days in the month and Zb is the difference between monthly average temperature T and base temperature Tb , normalized by the standard deviation of the daily average temperature sd :
Table 5 Ground Reflectance of Foreground Surfaces Foreground Surface Water (near normal incidences) Coniferous forest (winter) Asphalt, new weathered Bituminous and gravel roof Dry bare ground Weathered concrete Green grass Dry grassland Desert sand Light building surfaces Snow-covered surfaces: Typical city center Typical urban site Typical rural site Isolated rural site Source: Adapted from Thevenard and Haddad (2006).
Tb – T Zb = ---------------sd
Reflectance 0.07 0.07 0.05 0.10 0.13 0.2 0.2 to 0.3 0.26 0.2 to 0.3 0.4 0.6 0.2 0.4 0.5 0.7
(32)
(33)
Function f is the normal (Gaussian) probability density function with mean 0 and standard deviation 1, and function F is the equivalent cumulative normal probability function: –Z 2 1 f (Z) = ------------ exp --------- 2 2
(34)
Table 6 Fraction of Daily Temperature Range Time, h
Fraction
Time, h
Fraction
Time, h
Fraction
1 2 3 4 5 6 7 8
0.88 0.92 0.95 0.98 1.00 0.98 0.91 0.74
9 10 11 12 13 14 15 16
0.55 0.38 0.23 0.13 0.05 0.00 0.00 0.06
17 18 19 20 21 22 23 24
0.14 0.24 0.39 0.50 0.59 0.68 0.75 0.82
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Climatic Design Information
14.13 Table 7 Input Sources for Design-Day Generation
Design Day Type
Design Conditions
Daily Ranges
Limits
Dry-bulb Annual Monthly
0.4, 1, or 2% annual cooling DB/MCWB 0.4, 2, 5, or 10% DB/MCWB for month
Hottest month 5% DB MCDBR/MCWBR 5% DB MCDBR/MCWBR for month
Hourly wet-bulb temp. = min(dry-bulb temp., wet-bulb temp.)
Wet-bulb Annual Monthly
0.4, 1, or 2% annual cooling WB/MCDB 0.4, 2, 5, or 10% WB/MCDB for month
Hottest month 5% WB MCDBR/MCWBR 5% WB MCDBR/MCWBR for month
Hourly dry-bulb temp. = max(dry-bulb temp., wet-bulb temp.)
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Table 8 Derived Hourly Temperatures for Atlanta, GA for July for 5% Design Conditions, °C Hour (LDT)
tdb
twb
tdp
Hour (LDT)
tdb
twb
tdp
1 2 3 4 5 6 7 8 9 10 11 12
24.7 23.9 23.2 22.8 22.5 22.1 21.9 22.1 22.9 24.8 26.9 28.8
21.0 20.7 20.5 20.4 20.3 20.2 20.1 20.2 20.4 21.0 21.6 22.2
19.3 19.3 19.3 19.3 19.3 19.3 19.3 19.3 19.3 19.3 19.3 19.4
13 14 15 16 17 18 19 20 21 22 23 24
30.5 31.6 32.5 33.1 33.1 32.4 31.5 30.4 28.7 27.5 26.5 25.5
22.7 23.1 23.3 23.5 23.5 23.3 23.0 22.7 22.2 21.8 21.5 21.2
19.4 19.5 19.5 19.5 19.5 19.5 19.5 19.4 19.4 19.3 19.3 19.3
HDD15 = 31 × 3.93[–0.636 × 0.262 + 0.326] = 19.4°C-day. For cooling degree-days, Zb = 0.636. Note that f(–Zb) = f (Zb) and F(–Zb) = 1 – F(Zb), hence f (Zb) = 0.326
f z dz – Z
(35)
Both f and F are readily available as built-in functions in many scientific calculators or spreadsheet programs, so their manual calculation is rarely warranted. Cooling degree days CDDb to base Tb are calculated by the same equation: CDDb = Nsd [ZbF(Zb) + f (Zb)]
(36)
except that Zb is now expressed as T – Tb Zb = ---------------sd
(37)
Alternative Equations. The following formulas from ISO Standard 15927-6 give results very similar to Equations (32) and (36) but are somewhat simpler: N Tb – T HDDb = ------------------------------------------------------------------1 – exp – 2 T b – T s d
(38)
N T – Tb CDDb = ------------------------------------------------------------------1 – exp – 2 T – T b s d
(39)
When T = T b , the right-hand side of these equations become Ns d 2 .
Annual Degree-Days Annual degree-days are simply the sum of monthly degree days over the twelve months of the year. Example 10. Calculate heating and cooling degree-days (base 15°C) for Atlanta for the month of October.
and
F(Zb) = 0.738
and CDD15 = 31 × 3.93(0.636 × 0.737 + 0.326) = 96.9°C-day.) For most stations, the monthly degree-days calculated with this method are within 5°C-day of the observed values.
3.3
LDT = Local daylight saving time.
F(Z) =
Solution: For October in Atlanta, Table 1 provides T = 17.5°Cand sd = 3.93°C. For heating degree-days, Equation (33) provides Zb = (15 – 17.5)/3.93 = –0.636. From a scientific calculator or a spreadsheet program f (Zb) = 0.326, and F(Zb) = 0.262. Equation (32) then gives
REPRESENTATIVENESS OF DATA AND SOURCES OF UNCERTAINTY
Representativeness of Data The climatic design information in this chapter was obtained by direct analysis of observations from the indicated locations. Design values reflect an estimate of the cumulative frequency of occurrence of the weather conditions at the recording station, either for single or jointly occurring elements, for several years into the future. Several sources of uncertainty affect the accuracy of using the design conditions to represent other locations or periods. The most important of these factors is spatial representativeness. Most of the observed data for which design conditions were calculated were collected from airport observing sites, the majority of which are flat, grassy, open areas, away from buildings and trees or other local influences. Temperatures recorded in these areas may be significantly different from built-up areas where the design conditions are being applied. For example, the maximum urban heat island intensity may be 10 K or more (Oke 1987), although intraurban variability is typically quite large. Urban microclimate is affected by the three-dimensional density of building construction, usually represented by the ratio of building height to street width (H/ W); by type and extent of plant cover; and by anthropogenic heat emissions from buildings and vehicles. Significant variations can also occur with changes in local elevation, even if elevations differ by a few hundred metres, or in the vicinity of large bodies of water. It should be emphasized that such variations are not constant in time: intraurban differences in temperature and humidity fluctuate not only in predictable diurnal patterns, but also in response to changes in synoptic conditions and wind direction. Urban heat islands, for example, are typically prominent on clear nights with little or no wind, and are weaker or nonexistent in windy conditions and during daytime. Therefore, judgment must always be used in assessing the representativeness of the design conditions. Consult an applied climatologist regarding estimating design conditions for locations not listed in this chapter. For online references to applied climatologists in the United States, see www.ncdc.noaa.gov/customer -support/partnerships/regional-climate-centers. Also, GIS-compatible files (KML format) are provided as a special feature in ASHRAE
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14.14 Handbook Online. This allows use of the data in a GIS environment such as Google Earth or ArcGIS, which provides capabilities to overlay various layers of information such as elevation, land use, and bodies of water. This type of information can greatly assist in determining the most representative location to use for an application. Depending on a site’s specific geographic location and setting (e.g., proximity to large body of water or hills), the data in this chapter for the nearest weather station may not be representative of the actual climate experienced at the project site. In these instances, it may be beneficial to obtain climate data using procedures developed by ASHRAE research project RP-1561 (Qiu et al. 2016). The methodologies provide a protocol for using state-of-the-art mesoscale modeling techniques to derive meteorological conditions specific to the study area. The research project included the methodology based on the Weather Research and Forecasting (WRF) model designed to develop site-specific climate data where standard weather stations are unavailable or not representative of site conditions. The methodology was evaluated by using observations in various geographic regions, including coastal, mountain valley, mountain plateau, and major cities. The underlying data also depend on the method of observation. During the 1990s, most data gathering in the United States and Canada was converted to automated systems designated either an automated surface observation system (ASOS) or an automated weather observing system (AWOS). This change improved completeness and consistency of available data. However, changes have resulted from the inherent differences in type of instrumentation, instrumentation location, and processing procedures between the prior manual systems and ASOS. These effects were investigated in ASHRAE research project RP-1226 (Belcher and DeGaetano 2004). Comparison of one-year ASOS and manual records revealed some biases in dry-bulb temperature, dew-point temperature, and wind speed. These biases are judged to be negligible for HVAC engineering purposes; the tabulated design conditions in this chapter were derived from mixed automated and manual data as available. Changes in the location of the observing instruments often have a larger effect than changes in instrumentation. On the other hand, ASOS measurements of sky coverage and ceiling height differ markedly from manual observations and are incompatible with solar radiation models used in energy simulation software. An updated solar model, compatible with ASOS data, was developed as part of RP-1226. The ASOS-based model was found less accurate than models based on manually observed data when compared to measured solar radiation. Weather conditions vary from year to year and, to some extent, from decade to decade because of the inherent variability of climate. Similarly, values representing design conditions vary depending on the period of record used in the analysis. Thus, because of short-term climatic variability, there is always some uncertainty in using design conditions from one period to represent another period. Typically, values of design dry-bulb temperature vary less than 1 K from decade to decade, but larger variations can occur. Differing periods used in the analysis can lead to differences in design conditions between nearby locations at similar elevations. Design conditions may show trends in areas of increasing urbanization or other regions experiencing extensive changes to land use. Longer-term climatic change brought by human or natural causes may also introduce trends into design conditions. This is discussed further in the section on Effects of Climate Change. Wind speed and direction are very sensitive to local exposure features such as terrain and surface cover. The original wind data used to calculate the wind speed and direction design conditions in Table 1 are often representative of a flat, open exposure, such as at airports. Wind engineering methods, as described in Chapter 24, can be used to
2017 ASHRAE Handbook—Fundamentals (SI) Table 9
Locations Representing Various Climate Types
Cold Snow Forest Dry
Warm Rainy
Portland, ME Grand Island, NE Minot, ND Indianapolis, IN
Huntsville, AL Key West, FL Wilmington, NC West Palm Portland, OR Beach, FL Quillayute, WA
Amarillo, TX Bakersfield, CA Sacramento, CA Phoenix, AZ
Tropical Rainy
account for exposure differences between airport and building sites. This is a complex procedure, best undertaken by an experienced applied climatologist or wind engineer with knowledge of the exposure of the observing and building sites and surrounding regions.
Uncertainty from Variation in Length of Record ASHRAE research project RP-1171 (Hubbard et al. 2004) investigated the uncertainty associated with the climatic design conditions in the 2001 ASHRAE Handbook—Fundamentals. The main objectives were to determine how many years are needed to calculate reliable design values and to look at the frequency and duration of episodes exceeding the design values. Design temperatures in the 1997 and 2001 editions were calculated for locations for which there were at least 8 years of sufficient data; the criterion for using 8 years was based on unpublished work by TC 4.2. RP-1171 analyzed data records from 14 U.S. locations (Table 9) representing four different climate types. The dry-bulb temperatures corresponding to the five annual percentile design temperatures (99.6, 99, 0.4, 1, and 2%) from the 33-year period 1961-1993 (period used for the 2001 edition’s U.S. stations) were calculated for each location. The temperatures corresponding to the same percentiles for each contiguous subperiod ranging from 1 to 33 years in length was calculated, and the standard deviation of the differences between the resulting design temperature from each subperiod and the entire 33-year period was calculated. For instance, for a 10-year period, the dry-bulb values corresponding to each of the 23 subperiods 1961-1970, 1962-1971, … 1984-1993 were calculated and the standard deviation of differences with the dry-bulb value for the same percentile from the 33-year period calculated. The standard deviation values represent a measure of uncertainty of the design temperatures relative to the design temperature for the entire period of record. The results for the five annual percentiles are summarized in Figures 4A to 4E, each of which shows how the uncertainty (the average standard deviation for each of the locations in each climate type) varies with length of period. To the degree that the differences used to calculate the standard deviations are distributed normally, the short-period design temperatures can be expected to lie within one standard deviation of the long-term design temperature 68% of the time. For example, from Figure 4A, the uncertainty for the cold snow forest for a 1-year period is 3.6 K. This can be interpreted that the probability is 68% that the difference in a 99.6% dry-bulb in any given year will be within 6.3 K of the long-term 99.6% dry-bulb. Similarly, there is a 68% probability that the 99.6% dry-bulb from any 10-year period will be within 1 K of the long-term value for a location of the cold snow forest climate type. The uncertainty for the cold season is higher than for the warm season. For example, the uncertainty for the 99.6% dry-bulb for a 10-year period ranges from 0.6 to 1.0 K for the five climate types, whereas the uncertainty for the 0.4% dry-bulb for a 10-year period ranges from 0.4 to 0.6 K. A variety of other general characteristics of uncertainty are evident from an inspection of Figure 4. For example, the highest uncertainty of any climate type for a 10-year period is 1.1 K for the cold snow forest 99% dry-bulb case. The smallest uncertainty is 0.2 K for the tropical rainy 1% and 2% dry-bulb cases. Based on these results, it was concluded that using a minimum of 8 years of data would provide reliable (within ±1 K) climatic design calculations for most stations.
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Climatic Design Information
Fig. 4 Uncertainty versus Period Length for Various Dry-Bulb Temperatures, by Climate Type
Effects of Climate Change The evidence is unequivocal that the climate system is warming globally (IPCC 2007). The most frequently observed effects relate to increases in average, and to some degree, extreme temperatures. This is partly shown by the results of an analysis of design conditions conducted as part of calculating the values for the 2009 edition of this chapter (Thevenard 2009). For 1274 observing sites worldwide with suitably complete data from 1977 to 2006, selected
design conditions were compared between the period 1977-1986 and 1997-2006. The results, averaged over all locations, are as follows: • • • • •
The 99.6% annual dry-bulb temperature increased 1.52 K The 0.4% annual dry-bulb increased 0.79 K Annual dew point increased by 0.55 K Heating degree-days (base 18.3°C) decreased by 237°C-days Cooling degree-days (base 10°C) increased by 136°C-days
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14.16 Although these results are consistent with general warming of the world climate system, there are other effects that undoubtedly contribute, such as increased urbanization around many of the observing sites (airports, typically). There was no attempt in the analysis to determine the reasons for the changes. A more recent study by Thevenard and Shephard (2014), using stations used in the 2013 edition of this chapter, looked at trends for yearly average dry-bulb temperature and other quantities over the 1986 to 2010 period using statistical methods. The study showed that statistically significant increases in average dry-bulb temperature can be detected in only 19% of stations on an individual basis. However, trends become more apparent when stations are evaluated in groups. Stations were grouped in 5°×5° cells covering the globe. Of these cells, 44% showed an increase in average dry-bulb temperature, and 2% showed a decrease; 26% showed an increase in average dew-point temperature and 10% a decrease; finally, 34% showed an increase in average wet-bulb temperature, and 5% showed a decrease. Geographically, increases in average dry-bulb temperature were most visible throughout Europe, in China and southeast Asia, the eastern United States, and southern Australia, and are typically in the range of 0.2 to 0.6 K per decade. Northern locations exhibited higher positive trends (above 1 K per decade). Dew-point temperature increases were most visible in eastern Europe, whereas decreases were experienced in the southern United States and South America. Regardless of the reasons for increases, the general approach of developing design conditions based on analysis of the recent record (25 years, in this case) was specifically adopted for updating the values in this chapter as a balance between accounting for long-term trends and the sampling variation caused by year-to-year variation. Although this does not necessarily provide the optimum predictive value for representing conditions over the next one or two decades, it at least has the effect of incorporating changes in climate and local conditions as they occur, as updates are conducted regularly using recent data. Meteorological services worldwide are considering the many aspects of this complex issue in the calculation of climate “normals” (averages, extremes, and other statistical summary information of climate elements typically calculated for a 30-year period at the end of each decade). Livezey et al. (2007) and WMO (2007) provide detailed analyses and recommendations in this regard. Extrapolating design conditions to the next few decades based on observed trends should only be done with attention to the particular climate element and the regional and temporal characteristics of observed trends (Livezey et al. 2007).
2017 ASHRAE Handbook—Fundamentals (SI) episodes last days, whereas the longest warm-season episodes were always shorter than 24 h. These results were seen at almost all locations, and are general for the continental United States. The only exception was Phoenix, where the longest cold-season episodes were less than 24 h. This is likely the result of the southern latitude and dry climate, which produces a large daily temperature range, even in the cold season.
3.4 OTHER SOURCES OF CLIMATIC INFORMATION Joint Frequency Tables of Psychrometric Conditions Design values in this chapter were developed by ASHRAE research project RP-1699 (Roth 2017). The frequency tables used to calculate the simple design conditions, and the joint frequency matrices used to calculate the coincident design conditions, are available in ASHRAE’s Weather Data Viewer 6.0 (WDView 6.0) (ASHRAE 2017). WDView 6.0 gives users full access to the frequency tables and joint frequency matrices for all 8118 stations in the 2017 ASHRAE Handbook—Fundamentals via a spreadsheet, and provides the following capabilities: • Select a station by WMO number or region/country/state/name or by proximity to a given latitude and longitude. • Retrieve design climatic conditions for a specified station, in SI or I-P units.
Episodes Exceeding the Design Dry-Bulb Temperature Design temperatures based on annual percentiles indicate how many hours each year on average the specific conditions will be exceeded, but do not provide any information on the length or frequency of such episodes. As reported by Hubbard et al. (2004), each episode and its duration for the locations in Table 9 during which the 2001 design conditions represented by the 99.6, 99, 0.4, 1, and 2% dry-bulb temperatures were exceeded (i.e., were more extreme) was tabulated and their frequency of occurrence analyzed. The measure of frequency is the average number of episodes per year or its reciprocal, the average period between episodes. Cold- and warm-season results are presented in Figures 5A and 5B, respectively, for Indianapolis, IN, as a representative example. The duration for the 10-year period between episodes more extreme than the 99.6% design dry bulb is 37 h, and 62 h for the 99% design dry bulb. For the warm season, the 10-year period durations corresponding to the 0.4, 1, and 2% design dry bulb, are about 10, 12, and 15 h, respectively. Although the results in Hubbard et al. (2004) varied somewhat among the locations analyzed, generally the longest cold-season
Fig. 5
Frequency and Duration of Episodes Exceeding Design Dry-Bulb Temperature for Indianapolis, IN
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• Display frequency vectors and joint frequency matrices in the form of numerical tables. • Display frequency distribution and the cumulative frequency distribution functions in graphical form. • Display joint frequency functions in graphical form. • Display the table of years and months used for the calculation. • Display hourly binned dry-bulb temperature data. • Calculate heating and cooling degree-days to any base, using the method of Schoenau and Kehrig (1990). The Engineering Weather Data CD (NCDC 1999), an update of Air Force Manual 88-29, was compiled by the U.S. Air Force 14th Weather Squadron. This CD contains several tabular and graphical summaries of temperature, humidity, and wind speed information for hundreds of locations in the United States and around the world. In particular, it contains detailed joint frequency tables of temperature and humidity for each month, binned at 0.5°C and 3 h local time-ofday intervals. This CD is available from NCDC: www.ncdc.noaa .gov/nespls/olstore.prodspecific?prodnum=5005. The International Station Meteorological Climate Summary (ISMCS) is a CD-ROM containing climatic summary information for over 7000 locations around the world (NCDC 1996). A table providing the joint frequency of dry-bulb temperature and wet-bulb temperature depression is provided for the locations with hourly observations. It can be used as an aid in estimating design conditions for locations for which no other information is available. The CD is available at gcmd.nasa.gov/records/GCMD_gov.noaa.ncdc.C00268 .html. A web version of this product is now available free of charge from NCDC at www7.ncdc.noaa.gov/CDO/cdoselect.cmd?data setabbv=SUMMARIES. This service is also available via gis.ncdc .noaa.gov/map/viewer. The monthly frequency distribution of dry-bulb temperatures and mean coincident wet-bulb temperatures for 134 Canadian locations is available from Environment Canada (1983-1987).
Degree Days and Climate Normals The 1981 to 2010 climate normals for over 6000 United States locations are available online (free of charge) from the National Climatic Data Center: gis.ncdc.noaa.gov/map/viewer/. Also, users may generate normals/averages for any chosen period (dynamic normals) at www7.ncdc.noaa.gov/CDO/normals. The Canadian Climate Normals for 1971 to 2000 are available from Environment Canada at climate.weatheroffice.ec.gc.ca (Environment Canada 2003). The Climatography of the United States No. 20 (CLIM20), monthly station climate summaries for 1971 to 2000 are climatic station summaries of particular interest to engineering, energy, industry, and agricultural applications (NCDC 2004). These summaries contain a variety of statistics for temperature, precipitation, snow, freeze dates, and degree-day elements for 4273 stations. The statistics include means, medians (precipitation and snow elements), extremes, mean number of days exceeding threshold values, and heating, cooling, and growing degree-days for various temperature bases. Also included are probabilities for monthly precipitation and freeze data. Information on this product can be found at www .ncdc.noaa.gov/oa/documentlibrary/pdf/eis/clim20eis.pdf. Note that this is for 1971 to 2000 and not for the 1981 to 2010 period (latest normals) noted previously. Heating and cooling degree-day and degree-hour data for 3677 locations from 115 countries were developed by Crawley (1994) from the Global Daily Summary (GDS) version 1.0 and the International Station Meteorological Climate Summary (ISMCS) version 4.0 data.
14.17 Typical Year Data Sets Software is available to simulate the annual energy performance of buildings requiring a 1-year data set (8760 h) of weather conditions. Many data sets in different record formats have been developed to meet this requirement. The data represent a typical year with respect to weather-induced energy loads on a building. No explicit effort was made to represent extreme conditions, so these files do not represent design conditions. The National Renewable Energy Laboratory’s (NREL) TMY3 data set (Wilcox and Marion 2008) contains data for 1020 U.S. locations. TMY3, along with the 1991-2010 National Solar Radiation Data Base (NSRDB) (NREL 2011), contains hourly solar radiation [global, beam (direct), and diffuse] and meteorological data for 1454 stations; TMY3 is available at rredc.nrel.gov/solar/old_data/nsrdb /1991-2005/tmy3/, and the NSRDB at www.ncdc.noaa.gov/land -based-station-data/solar-radiation/. These were produced using an objective statistical algorithm to select the most typical month from the long-term record. A more recent source of gridded weather, solar radiation, and environmental TMY data with a visual and dynamic interface is available from maps.nrel.gov/nsrdb-viewer. The solar radiation data are derived from satellite data, and the environmental data are downscaled from the MERRA reanalysis data set derived from a large climate forecasting model (Rienecker et al. 2011). The grid spacing for this source of data is about 4 km, and currently covers North America up to 50° N, as well as a part of South America down to 10° S. Various types of TMY data are available there, currently for the period 1998 to 2014, as well as each historical year within that time period, with anticipated annual updates. Canadian Weather Year for Energy Calculation (CWEC) files for 47 Canadian locations were developed for use with the Canadian National Energy Code, using the TMY algorithm and software (Environment Canada 1993). Files for 75 locations are now available. ASHRAE’s International Weather for Energy Calculations (IWEC2) data set (Huang et al. 2014) contains typical-year weather data for 3012 international locations outdoor of the United States and Canada. The IWEC2s were developed through ASHRAE RP1477, which used the same source of raw weather data (ISD; Lott et al. 2001) as used for the design condition tables in this chapter, but for a slightly earlier time period of 12 to 25 years ending in 2009. The IWEC2 data set is available on a DVD from the ASHRAE Climate Data Center at www.ashrae.org/resources--publications /bookstore/climate-data-center#iwec; individual files and country sets are also available online from commercial resellers.
Sequences of Extreme Temperature and Humidity Durations Colliver (1997) and Colliver et al. (1998) compiled extreme sequences of 1-, 3-, 5-, and 7-day duration for 239 U.S. and 144 Canadian locations based independently on the following five criteria: high dry-bulb temperature, high dew-point temperature, high enthalpy, low dry-bulb temperature, and low wet-bulb depression. For the criteria associated with high values, the sequences are selected according to annual percentiles of 0.4, 1.0, and 2.0. For the criteria corresponding to low values, annual percentiles of 99.6, 99.0, and 98.0 are reported. Although these percentiles are identical to those used to select annual heating and cooling design temperatures, the maximum or minimum temperatures within each sequence are significantly more extreme than the corresponding design temperatures. The data included for each hour of a sequence are solar radiation, dry-bulb and dew-point temperature, atmospheric pressure, and wind speed and direction. Accompanying information allows the user to go back to the source data and obtain sequences with different characteristics (e.g., different probability of occurrence, windy conditions, low or high solar radiation). These extreme sequences are available on CD (ASHRAE 1997).
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2017 ASHRAE Handbook—Fundamentals (SI)
These sequences were developed primarily to assist the design of heating or cooling systems having a finite capacity before regeneration is required or of systems that rely on thermal mass to limit loads. The information is also useful where information on the hourly weather sequence during extreme episodes is required for design.
Global Weather Data Source Web Page Because of growing demand for more comprehensive global coverage of weather data for HVAC applications around the world, ASHRAE sponsored research project RP-1170 (Plantico 2001) to construct a Global Weather Data Sources (GWDS) web page. Many national climate services and other climate data sources are making more information available over the Internet. The purpose of RP1170 was to provide ASHRAE membership with easy access to major sources of international weather data through one consolidated online system. This web page was later updated to better use the resources of the World Meteorological Organization (WMO) and NCDC. The GWDS web page is accessible at www.ncdc.noaa .gov/oa/ashrae/gwds-title.html.
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Observational Data Sets For detailed designs, custom analysis of the most appropriate long-term weather record is best. National weather services are generally the best source of long-term observational data. The National Climatic Data Center (NCDC), in conjunction with U.S. Air Force and Navy partners in Asheville’s Federal Climate Complex (FCC), developed the global Integrated Surface Data (Lott 2004; Lott et al. 2001) to address a pressing need for an integrated global database of hourly land surface climatological data. The database of over 20,000 stations contains hourly and some daily summary data from as early as 1900 (many stations beginning in the 1948-1973 timeframe), is operationally updated each day with the latest available data, and is now being further integrated with various data sets from the United States and other countries to further expand the spatial and temporal coverage of the data. For access to ISD, go to www.ncdc.noaa.gov /isd or, for a GIS interface, gis.ncdc.noaa.gov/map/viewer/. For a complete review of ISD and all of its products, go to www.ncdc.noaa .gov/isd. The National Solar Radiation Database (NSRDB) (www.ncdc .noaa.gov/land-based-station-data/solar-radiation/; maps.nrel.gov /nsrdb-viewer) and Canadian Weather Energy and Engineering Data Sets (CWEEDS) (Environment Canada 1993) provide long-term hourly data, including solar radiation values for the United States and Canada. A previous version of the NSRDB required a modified solar radiation model because of the implementation of automated observing systems that do not report traditional cloud elements. The current NSRDB Data Viewer (maps.nrel.gov/nsrdb-viewer), which is mentioned earlier in the section on Typical Year Data Sets, also contains both solar radiation and environmental data for every year from 1998 through 2014 covering North America up to 50° North, and South America down to 10° South. The solar radiation data are derived from GOES satellites and the environmental data are downscaled from MERRA reanalysis data set. Considerable information about weather and climate services and data sets is available elsewhere online. Information supplementary to this chapter may also be posted on the ASHRAE Technical Committee 4.2 website, the link to which is available from the ASHRAE website (www.ashrae.org).
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ASHRAE. 1997. Design weather sequence viewer 2.1. (CD-ROM). ASHRAE. 2017. Weather data viewer, version 6.0. (CD-ROM).
ASHRAE. 2013. Climatic data for building design standards. ANSI/ ASHRAE Standard 169-2013. Belcher, B.N., and A.T. DeGaetano. 2004. Integration of ASOS weather data into building energy calculations with emphasis on model-derived solar radiation (RP-1226). ASHRAE Research Project, Final Report. Colliver, D.G. 1997. Sequences of extreme temperature and humidity for design calculations (RP-828). ASHRAE Research Project, Final Report. Colliver, D.G., R.S. Gates, H. Zhang, and K.T. Priddy. 1998. Sequences of extreme temperature and humidity for design calculations. ASHRAE Transactions 104(1A):133-144. Colliver, D.G., R.S. Gates, T.F. Burkes, and H. Zhang. 2000. Development of the design climatic data for the 1997 ASHRAE Handbook—Fundamentals. ASHRAE Transactions 106(1). Crawley, D.B. 1994. Development of degree day and degree hour data for international locations. D.B. Crawley Consulting, Washington, D.C. Diamond, H., T. Karl, M. Palecki, C. Baker, J. Bell, R. Leeper, D. Easterling, J. Lawrimore, T. Meyers, M. Helfert, G. Goodge, and P. Thorne. 2013. U.S. Climate Reference Network after one decade of operations: Status and assessment. Bulletin of the American Meteorological Society 94(4): 485-498. Environment Canada. 1983-1987. Principal station data. PSD 1 to 134. Atmospheric Environment Service, Downsview, Ontario. Environment Canada. 1993. Canadian weather for energy calculations (CWEC files) user’s manual. Atmospheric Environment Service, Downsview, Ontario. Eskes, H., et al. 2015. Validation of reactive gases and aerosols in the MACC global analysis and forecast system. Geoscientific Model Development 8(11):3523-3543. www.geosci-model-dev.net/8/3523/2015/gmd-8-3523 -2015-discussion.html. FAO. 2011. Climate impact on agriculture. Food and Agriculture Organization of the United Nations, Rome. GHCN. 2015. Global historical climatology network—Daily. National Climatic Data Center, National Oceanic and Atmospheric Administration, Asheville, NC. www.ncdc.noaa.gov/oa/climate/ghcn-daily/ and ftp.ncdc .noaa.gov/pub/data/ghcn/daily/. GPCC. 2015. GPCC Full Data Reanalysis Version 7. Global Precipitation Climatology Centre, Deutscher Wetterdienst, Offenbach am Main, Germany. ftp.dwd.de/pub/data/gpcc/html/fulldata_v7_doi_download .html. Gueymard, C.A. 1987. An anisotropic solar irradiance model for tilted surfaces and its comparison with selected engineering algorithms. Solar Energy 38:367-386. Erratum, Solar Energy 40:175 (1988). Gueymard, C.A. 2008. REST2: High performance solar radiation model for cloudless-sky irradiance, illuminance and photosynthetically active radiation—Validation with a benchmark dataset. Solar Energy 82:272-285. Gueymard, C.A., and D. Thevenard. 2009. Monthly average clear-sky broadband irradiance database for worldwide solar heat gain and building cooling load calculations. Solar Energy 83:1998-2018. Harriman, L.G., D.G. Colliver, and H.K. Quinn. 1999. New weather data for energy calculations. ASHRAE Journal 41(3):31-38. Hedrick, R. 2009. Generation of hourly design-day weather data (RP-1363). ASHRAE Research Project, Final Report (Draft). Huang, Y.J., F.X. Su, D.H. Seo, and M. Krarti. 2014. Development of ASHRAE IWEC2 weather files from the Integrated Surface Hourly (ISH) data base of historical weather data for 3,012 international locations. ASHRAE Transactions 120(1):340-355. Hubbard, K., K. Kunkel, A. DeGaetano, and K. Redmond. 2004. Sources of uncertainty in the calculation of the design weather conditions in the ASHRAE Handbook of Fundamentals (RP-1171). ASHRAE Research Project, Final Report. Inness, A., et al. 2013. The MACC reanalysis: An 8 yr data set of atmospheric composition. Atmospheric Chemistry and Physics 13(8):4073-4109. www.atmos-chem-phys.net/13/4073/2013/acp-13-4073-2013-discussion .html. IPCC. 2007. Fourth assessment report: Summary for policy makers. International Panel on Climate Change, World Meteorological Organization, Geneva. www.ipcc.ch/pdf/assessment-report/ar4/wg1/ar4-wg1-spm.pdf. Iqbal, M. 1983. An introduction to solar radiation. Academic Press, Toronto. ISO. 2007. Hygrothermal performance of buildings—Calculation and presentation of climatic data—Part 6: Accumulated temperature differences (degree days). Standard 15927-6. International Organization for Standardization, Geneva.
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Licensed for single user. © 2017 ASHRAE, Inc.
Climatic Design Information Kasten, F., and T. Young. 1989. Revised optical air mass tables and approximation formula. Applied Optics 28:4735-4738. Lamming, S.D., and J.R. Salmon. 1996. Wind data for design of smoke control systems (RP-816). ASHRAE Research Project, Final Report. Lamming, S.D., and J.R. Salmon. 1998. Wind data for design of smoke control systems. ASHRAE Transactions 104(1A):742-751. Livezey, R.E., K.Y. Vinnikov, M.M. Timofeyeva, R. Tinker, and H.M. Van Den Dool. 2007. Estimation and extrapolation of climate normals and climatic trends. Journal of Applied Meteorology and Climatology 46: 1759-1776. Lott, J.N. 2004. The quality control of the integrated surface hourly database. 84th American Meteorological Society Annual Meeting, Seattle, WA. ams.confex.com/ams/pdfpapers/71929.pdf. Lott, J.N., R. Baldwin, and P. Jones. 2001. The FCC Integrated Surface Hourly Database, a new resource of global climate data. NCDC Technical Report 2001-01. National Climatic Data Center, Asheville, NC. ftp.ncdc.noaa.gov/pub/data/techrpts/tr200101/tr2001-01.pdf. Lowery, M.D., and J.E. Nash. 1970. A comparison of methods of fitting the double exponential distribution. Journal of Hydrology 10(3):259-275. Molod, A., L. Takacs, M. Suarez, and J. Bacmeister. 2015. Development of the GEOS-5 atmospheric general circulation model: Evolution from MERRA to MERRA2. Geoscientific Model Development 8:1339-1356. NCDC. 1996. International station meteorological climate summary (ISMCS). National Climatic Data Center, Asheville, NC. NCDC. 1999. Engineering weather data. National Climatic Data Center, Asheville, NC. NCDC. 2004. Monthly station climate summaries. In Climatography of the U.S. #20. National Climatic Data Center, Asheville, NC. NREL. 2011. National solar radiation database, 1991-2010 update: User’s manual. Technical Report NREL/TP-581-41364. National Renewable Energy Laboratory, Golden, CO. rredc.nrel.gov/solar/old_data/nsrdb /1991-2010/. Oke, T.R. 1987. Boundary layer climates, 2nd ed. Methuen, London. Perez, R., P. Ineichen, R. Seals, J. Michalsky, and R. Stewart. 1990. Modeling daylight availability and irradiance components from direct and global irradiance. Solar Energy 44(5):271-289. Plantico, M. 2001. Identify and characterize international weather data sources (RP-1170). ASHRAE Research Project, Final Report. Rienecker, M.M., et al. 2011. MERRA: NASA’s Modern-Era Retrospective Analysis for Research and Applications. Journal of Climate 24:36243648. journals.ametsoc.org/doi/full/10.1175/JCLI-D-11-00015.1. Qiu, X., M. Roth, H. Corbett-Hains, and F. Yang. 2016. Mesoscale climate modeling procedure development and performance evaluation (RP-1561). ASHRAE Transactions 122(2). Roth, M. 2017. Updating climatic design data in the 2017 ASHRAE Handbook—Fundamentals (RP-1699). ASHRAE Research Project RP-1699, Final Report (in preparation).
14.19 Schoenau, G.J., and R.A. Kehrig. 1990. A method for calculating degreedays to any base temperature. Energy and Buildings 14:299-302. Smith, A., N. Lott, and R. Vose. 2011. The integrated surface database: Recent developments and partnerships. Bulletin of the American Meteorological Society 92(6):704. Stephenson, D.G. 1965. Equations for solar heat gain through windows. Solar Energy 9(2):81-86. Thevenard, D. 2009. Updating the ASHRAE climatic data for design and standards (RP-1453). ASHRAE Research Project, Final Report. Thevenard, D., and C. Gueymard. 2013. Updating climatic design data in Chapter 14 of the 2013 Handbook of Fundamentals (RP-1613). ASHRAE Research Project RP-1613, Final Report. Thevenard, D., and K. Haddad. 2006. Ground reflectivity in the context of building energy simulation. Energy and Buildings 38(8):972-980. Thevenard, D., and M. Shephard. 2014. Temperature trends for locations listed in the tables of climatic design conditions in the 2013 ASHRAE Handbook—Fundamentals. ASHRAE Transactions 120(2). Thevenard, D., J. Lundgren, and R. Humphries. 2005. Updating the climatic design conditions in the ASHRAE Handbook of Fundamentals (RP-1273). ASHRAE Research Project, Final Report. Threlkeld, J.L. 1963. Solar irradiation of surfaces on clear days. ASHRAE Transactions 69:24. Wilcox, S., and W. Marion. 2008. Users manual for TMY3 data sets. Technical Report NREL/TP-581-43156. National Renewable Energy Laboratory, Golden, CO. www.nrel.gov/docs/fy08osti/43156.pdf. WMO. 2007. The role of climatological normals in a changing climate. Technical Document 1377. World Meteorological Organization, Geneva.
BIBLIOGRAPHY ASHRAE. 2006. Weather data for building design standards. ANSI/ ASHRAE Standard 169-2006. Environment Canada. 2013. Canadian 1971-2000 climate normals. Meteorological Service of Canada, Downsview, Ontario. climate.weather office.ec.gc.ca. NCDC. 2002. Monthly normals of temperature, precipitation, and heating and cooling degree-days. In Climatography of the United States #81. National Climatic Data Center, Asheville, NC. NCDC. 2002. Annual degree-days to selected bases (1971-2000). In Climatography of the United States #81. National Climatic Data Center, Asheville, NC. NCDC. 2003. Data documentation for data set 3505 (DSI-3505) integrated surface hourly (ISH) data. National Climatic Data Center, Asheville, NC. Thevenard, D., and R. Humphries. 2005. The calculation of climatic design conditions in the 2005 ASHRAE Handbook—Fundamentals. ASHRAE Transactions 111(1):457-466.
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APPENDIX: DESIGN CONDITIONS FOR SELECTED LOCATIONS (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry-bulb temperature, °C WB: Wet-bulb temperature, °C MCWB: Mean coincident wet-bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew-point temperature, °C MCDB: Mean coincident dry-bulb temperature, °C
Heating DB 99.6%
99%
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
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United States of America Alabama AUBURN UNIVERSITY REGIONAL BIRMINGHAM SHUTTLESWORTH INTL CAIRNS AAF DOTHAN REGIONAL HUNTSVILLE INTL MAXWELL AFB MOBILE REGIONAL MONTGOMERY REGIONAL NORTHEAST ALABAMA AP NORTHWEST ALABAMA REGIONAL TUSCALOOSA REGIONAL Alaska BRYANT AAF ELMENDORF AFB FAIRBANKS INTL JUNEAU INTL LAKE HOOD SEAPLANE BASE MERRILL FIELD TED STEVENS ANCHORAGE INTL Arizona CASA GRANDE MUNICIPAL DAVIS-MONTHAN AFB FLAGSTAFF PULLIAM AP LUKE AFB PHOENIX SKY HARBOR INTL PRESCOTT MUNICIPAL TUCSON INTL WINDOW ROCK AP YUMA INTL AP Arkansas BENTONVILLE MUNICIPAL CLINTON NATL DRAKE FIELD FORT SMITH REGIONAL GRIDER FIELD JONESBORO MUNICIPAL LITTLE ROCK AFB NORTH LITTLE ROCK MUNICIPAL ROGERS MUNICIPAL SMITH FIELD TEXARKANA REGIONAL California ALAMEDA BEALE AFB BOB HOPE AP BROWN FIELD MUNICIPAL CAMARILLO AP CAMP PENDLETON MCAS CASTLE AFB DESERT RESORTS REGIONAL EL TORO MCAS FRESNO YOSEMITE INTL FULLERTON MUNICIPAL HAWTHORN MUNICIPAL HAYWARD EXECUTIVE IMPERIAL COUNTY AP
32.616N 33.566N 31.267N 31.317N 34.644N 32.383N 30.688N 32.300N 33.967N 34.744N 33.212N
85.433W 86.745W 85.717W 85.450W 86.786W 86.350W 88.246W 86.408W 86.083W 87.600W 87.616W
237 188 92 114 190 52 66 62 173 165 46
-4.9 -6.2 -2.9 -2.6 -7.4 -3.5 -2.4 -4.3 -7.3 -6.8 -5.3
-2.5 -3.9 -1.2 -0.6 -5.1 -1.3 -0.5 -2.3 -5.2 -4.5 -3.0
34.1 35.4 35.3 35.8 35.2 36.4 34.5 36.0 34.1 35.7 36.3
23.0 23.7 24.6 24.3 23.9 24.5 24.9 24.5 23.6 24.1 24.1
32.9 34.0 34.1 34.4 33.8 35.3 33.5 34.8 32.9 34.3 34.8
23.3 23.6 24.5 24.1 23.7 24.6 24.7 24.4 23.6 24.0 24.3
32.3 32.8 33.0 33.3 32.6 34.1 32.5 33.7 32.2 33.0 33.6
23.1 23.5 24.2 23.9 23.4 24.6 24.5 24.2 23.5 23.7 24.1
25.3 25.8 26.9 26.6 25.8 26.8 26.7 26.4 25.6 26.0 26.3
31.3 31.4 31.5 32.0 31.3 32.9 31.4 32.6 31.6 32.0 32.3
24.8 25.3 26.3 25.9 25.3 26.3 26.2 25.8 25.0 25.5 25.8
30.4 30.8 30.9 31.1 30.8 32.3 30.7 31.7 31.0 31.4 31.6
23.8 24.4 26.0 25.2 24.4 25.2 25.5 24.8 23.8 24.5 24.9
19.2 19.8 21.6 20.6 19.8 20.5 20.9 20.0 19.0 19.8 20.1
27.5 28.2 28.6 28.5 28.2 29.0 28.5 28.9 28.6 28.7 28.5
22.9 23.7 25.1 24.6 23.8 24.8 25.0 24.3 22.9 23.9 24.1
18.2 19.0 20.5 19.9 19.1 19.9 20.3 19.3 18.0 19.1 19.1
26.8 27.7 27.8 28.1 27.6 28.7 28.0 28.4 27.8 28.1 28.0
61.266N 149.653W 61.253N 149.794W 64.804N 147.876W 58.357N 134.564W 61.178N 149.966W 61.217N 149.855W 61.169N 150.028W
118 65 132 5 27 42 37
-28.7 -26.3 -41.6 -15.2 -22.3 -23.7 -22.6
-25.5 -23.7 -39.2 -12.8 -19.8 -21.7 -20.3
23.8 23.4 27.4 23.3 23.3 22.9 21.9
15.7 14.8 16.0 15.3 15.3 15.2 15.0
22.0 22.0 25.6 21.2 21.5 21.4 20.2
14.9 14.3 15.5 14.5 14.5 14.6 14.1
20.1 20.1 23.8 19.2 19.9 20.0 18.9
13.9 13.5 14.7 13.6 13.7 13.9 13.5
16.5 16.0 17.3 16.1 16.1 16.2 15.8
22.6 21.3 25.0 21.8 21.9 21.4 20.6
15.4 15.2 16.4 15.2 15.2 15.4 15.0
20.8 19.8 23.3 19.7 20.1 19.9 19.1
13.5 14.0 14.6 14.0 13.6 14.0 13.6
9.8 10.0 10.5 10.0 9.8 10.0 9.7
18.6 16.2 18.6 16.3 17.4 17.1 17.2
12.7 13.0 13.6 13.3 12.8 13.0 12.9
9.3 9.4 9.9 9.5 9.2 9.4 9.3
17.0 15.8 17.9 15.8 16.8 16.8 16.5
32.950N 111.767W 32.167N 110.883W 35.144N 111.666W 33.550N 112.367W 33.428N 112.004W 34.652N 112.421W 32.131N 110.955W 35.658N 109.061W 32.650N 114.600W
446 824 2135 331 337 1537 777 2054 63
-0.1 0.2 -15.6 1.6 4.0 -7.9 -0.1 -17.7 5.7
1.6 1.9 -12.6 3.0 5.4 -6.2 1.4 -14.9 7.1
42.5 40.7 29.9 43.8 43.5 34.7 41.0 32.1 43.9
20.9 18.4 12.9 21.1 20.8 15.9 18.9 13.4 22.5
41.4 39.3 28.6 42.6 42.4 33.4 39.8 31.0 42.7
20.6 18.3 12.7 21.0 20.7 15.6 18.7 13.2 22.4
40.3 38.0 27.4 41.4 41.3 32.3 38.6 29.5 41.9
20.3 18.1 12.6 20.9 20.6 15.4 18.5 12.8 22.0
23.3 22.5 16.3 24.6 24.3 19.1 22.5 16.5 26.3
33.8 29.7 22.6 34.5 35.0 27.2 31.1 24.2 35.8
22.9 22.1 15.7 24.0 23.8 18.5 22.1 15.8 25.6
34.4 30.1 22.3 34.4 34.8 26.4 30.8 23.7 35.2
21.2 21.1 14.6 22.5 21.8 17.2 20.7 14.4 23.9
16.8 17.5 13.5 17.9 17.2 14.9 16.9 13.2 19.0
26.4 24.5 17.7 27.3 27.8 21.4 24.6 17.5 30.5
20.0 20.2 13.8 21.4 20.8 16.3 20.0 13.7 22.9
15.5 16.4 12.8 16.7 16.1 14.0 16.2 12.6 17.8
26.4 24.9 17.3 28.2 28.9 21.0 24.8 17.5 31.0
36.350N 34.727N 36.010N 35.333N 34.175N 35.831N 34.917N 34.835N 36.372N 36.191N 33.454N
94.217W 92.239W 94.169W 94.363W 91.935W 90.646W 92.150W 92.260W 94.107W 94.491W 94.007W
395 79 381 137 63 80 95 172 412 364 110
-12.2 -6.3 -12.0 -7.9 -5.3 -8.3 -7.7 -7.6 -12.3 -12.1 -4.6
-8.9 -4.2 -8.7 -5.4 -3.4 -6.3 -5.4 -4.9 -9.1 -8.8 -2.7
35.9 37.1 35.5 37.9 36.4 36.2 37.7 35.2 35.0 36.2 37.4
23.6 25.0 23.5 24.3 25.3 24.9 25.1 24.8 22.9 23.5 24.3
33.6 35.6 33.8 36.3 35.1 34.8 36.1 33.9 33.3 34.0 36.0
23.6 25.0 23.5 24.5 25.1 24.6 25.3 24.6 23.1 23.5 24.4
32.4 34.1 32.4 34.7 33.9 33.6 34.6 32.7 32.1 32.5 34.6
23.4 24.8 23.5 24.4 24.9 24.3 25.1 24.2 22.9 23.4 24.3
25.3 26.7 25.5 26.4 26.9 26.7 27.4 26.2 25.0 25.2 26.3
32.2 33.3 31.6 33.5 33.2 33.0 33.2 32.4 31.2 32.4 32.9
24.7 26.2 24.9 25.8 26.3 26.0 26.7 25.6 24.3 24.6 25.8
31.4 32.7 31.0 32.7 32.7 32.3 32.6 31.6 30.5 31.6 32.2
23.0 25.0 23.7 24.6 25.2 25.0 26.1 24.5 23.0 22.9 24.7
18.6 20.3 19.5 20.0 20.5 20.3 21.7 19.8 18.6 18.4 19.9
29.0 29.7 28.5 29.3 29.8 29.7 29.6 29.3 28.1 29.2 28.9
22.6 24.5 22.9 24.1 24.7 24.1 25.2 23.9 22.4 22.5 24.1
18.1 19.6 18.5 19.3 19.8 19.2 20.6 19.2 18.0 18.0 19.3
28.5 29.2 27.8 28.8 29.4 29.0 29.3 28.8 27.7 28.7 28.4
37.772N 122.298W 39.133N 121.433W 34.201N 118.358W 32.572N 116.979W 34.217N 119.083W 33.300N 117.350W 37.383N 120.567W 33.627N 116.159W 33.667N 117.733W 36.780N 119.719W 33.872N 117.979W 33.923N 118.334W 37.654N 122.115W 32.834N 115.579W
2 34 236 157 24 23 58 -36 117 102 29 19 13 -18
4.7 -0.5 3.8 3.9 3.0 0.0 -1.1 -0.5 6.2 -0.1 4.3 6.9 2.7 2.0
5.8 1.1 5.0 5.3 4.3 1.5 -0.1 1.3 7.4 1.2 6.0 7.7 3.8 3.3
27.3 38.4 36.4 32.0 30.1 33.3 39.2 44.2 33.3 39.8 35.0 31.1 30.6 44.1
17.8 21.3 19.9 17.8 16.8 18.8 20.9 22.3 19.9 21.3 19.4 17.1 18.5 22.6
25.1 36.8 34.5 29.5 28.0 31.1 37.7 42.8 31.6 38.3 32.9 28.9 28.0 42.8
17.2 20.6 19.3 18.1 17.3 18.5 20.1 22.0 19.7 20.5 19.3 17.2 17.9 22.5
23.3 35.1 32.8 27.8 26.6 29.0 36.5 41.7 30.0 36.9 31.3 27.3 25.8 41.9
16.6 19.9 19.1 18.1 17.3 18.4 19.6 21.8 19.3 20.0 19.0 17.5 17.4 22.3
19.2 22.8 22.6 21.7 20.6 21.9 22.3 26.5 22.1 23.0 22.4 21.0 19.8 27.2
25.9 35.5 32.0 27.7 26.2 28.6 35.6 36.1 30.1 36.3 30.7 26.5 27.6 36.2
18.4 21.8 21.7 20.9 19.9 21.2 21.5 25.7 21.4 22.0 21.6 20.3 18.9 26.4
24.6 34.1 30.8 26.2 25.3 27.7 34.6 36.0 29.1 35.0 29.5 25.5 26.1 35.7
16.9 18.2 19.3 19.8 18.3 19.8 18.0 23.9 19.1 18.5 19.4 19.0 17.0 25.0
12.1 13.2 14.5 14.8 13.2 14.5 13.1 18.7 14.1 13.5 14.2 13.8 12.1 20.1
20.5 27.9 25.4 24.1 23.1 24.9 23.5 31.6 26.2 29.8 26.1 23.6 22.0 31.3
16.1 17.3 18.7 19.0 17.6 18.9 17.4 22.8 18.3 17.2 18.8 18.1 16.2 23.9
11.4 12.4 13.9 14.0 12.7 13.7 12.5 17.4 13.4 12.5 13.6 13.0 11.5 18.7
19.9 26.5 24.9 23.3 22.1 24.4 22.6 31.7 25.4 28.9 25.4 22.4 20.7 31.6
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 542 sites, 1410 more on CD-ROM 11 sites, 28 more on CD-ROM 8.0 7.0 5.9 1342 1064 8.3 7.4 6.5 1450 1141 7.8 6.8 5.7 998 1321 8.8 7.9 7.0 964 1401 9.1 8.1 7.3 1696 1026 8.0 7.1 5.9 1087 1414 9.0 8.0 7.2 912 1398 8.4 7.4 6.4 1172 1307 7.4 6.3 5.5 1766 883 8.5 7.5 6.6 1665 1059 7.7 6.7 5.7 1361 1222 7 sites, 120 more on CD-ROM 8.6 6.5 5.2 5932 3 8.8 7.3 5.9 5736 7 7.6 6.4 5.4 7543 38 11.8 10.5 8.8 4654 2 8.2 7.1 5.8 5412 8 6.9 5.5 4.7 5534 7 9.3 8.3 7.4 5619 3 9 sites, 14 more on CD-ROM 9.2 7.9 6.7 832 1986 9.2 8.1 7.2 809 1790 11.1 9.3 8.2 3778 72 9.3 8.1 7.1 679 2245 8.4 7.3 6.1 507 2576 9.6 8.4 7.5 2303 567 9.6 8.4 7.5 778 1839 11.5 9.5 8.2 3553 170 9.4 8.3 7.4 365 2638 11 sites, 20 more on CD-ROM 8.8 7.8 7.0 2246 798 8.6 7.7 7.0 1609 1226 9.3 8.4 7.6 2201 792 9.1 8.0 7.1 1734 1184 8.7 7.8 7.1 1525 1235 10.1 8.7 7.8 1947 1088 8.2 7.1 6.0 1742 1166 8.2 7.4 6.6 1764 1076 9.9 8.6 7.7 2248 794 10.5 9.1 8.2 2197 827 8.4 7.5 6.6 1360 1316 55 sites, 60 more on CD-ROM 9.1 7.9 7.0 1393 89 10.4 8.7 7.5 1332 860 8.1 6.6 5.7 767 805 7.3 5.8 5.3 915 371 11.2 9.0 7.3 1011 234 7.4 6.3 5.5 1010 377 9.1 7.9 6.5 1324 979 9.0 8.0 7.0 607 2175 6.9 5.5 4.7 617 651 8.2 7.3 6.5 1231 1203 5.7 4.9 4.5 618 751 7.3 6.2 5.6 614 455 8.7 7.9 7.2 1412 163 11.6 9.8 8.4 521 2313
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station LEMOORE NAS LIVERMORE MUNICIPAL LOMPOC AP LONG BEACH AP LOS ANGELES INTL LOS ANGELES ONTARIO INTL MARCH AFB MCCLELLAN-PALOMAR AP MEADOWS FIELD MINETA SAN JOSE INTL MIRAMAR MCAS MODESTO CITY-COUNTY AP MOFFETT FEDERAL AIRFIELD MONTEREY REGIONAL MONTGOMERY FIELD NAPA COUNTY AP NORTH ISLAND NAS OAKLAND INTL PALM SPRINGS INTL POINT ARGUELLO POINT MUGU NAS PORTERVILLE MUNICIPAL REDDING MUNICIPAL RIVERSIDE MUNICIPAL SACRAMENTO EXECUTIVE SACRAMENTO INTL SACRAMENTO MATHER AP SACRAMENTO MCCLELLAN AFB SALINAS MUNICIPAL SAN BERNARDINO INTL SAN DIEGO INTL SAN FRANCISCO INTL SAN LUIS OBISPO CO REGIONAL SANTA BARBARA MUNICIPAL SANTA MARIA PUBLIC AP SONOMA COUNTY AP SOUTHERN CALIFORNIA LOGISTICS STOCKTON METROPOLITAN TRAVIS AFB VISALIA MUNICIPAL WILLIAM J FOX AP Colorado BUCKLEY AFB CENTENNIAL AP COLORADO SPRINGS MUNICIPAL DENVER INTL DENVER STAPLETON FORT COLLINS DOWNTOWN FORT COLLINS LOVELAND MUNI GRAND JUNCTION REGIONAL GREELEY-WELD COUNTY AP PUEBLO MEMORIAL Connecticut BRADLEY INTL HARTFORD-BRAINARD AP IGOR SIKORSKY MEMORIAL WATERBURY-OXFORD AP WINDHAM AP
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 39.6 21.5 38.1 20.9 36.9 20.4 37.1 19.8 34.9 18.9 32.7 18.2 27.6 16.1 25.2 16.1 23.4 15.8 33.0 19.0 31.0 19.0 29.2 18.7 28.8 17.3 26.9 17.7 25.4 17.9 37.9 21.0 36.4 20.2 34.8 19.9 38.5 19.9 37.1 19.4 35.5 19.0 28.8 16.7 27.1 17.7 25.5 17.8 39.4 21.3 38.0 20.7 36.7 20.1 33.0 18.9 31.0 18.5 28.9 18.0 32.6 18.8 30.8 18.8 28.9 18.6 38.5 21.0 36.8 20.2 35.2 19.5 31.2 18.6 28.7 18.1 27.0 17.8 25.9 15.5 23.1 15.1 21.7 14.9 32.4 18.6 30.3 18.3 28.5 18.3 32.8 18.8 30.3 18.2 28.0 17.7 29.2 17.8 27.4 18.4 26.0 18.8 28.6 17.9 26.2 17.2 23.9 16.8 44.1 21.5 42.9 21.3 42.0 21.1 22.0 N/A 20.0 N/A 18.6 N/A 27.5 16.2 25.8 16.9 24.1 17.2 38.0 21.1 37.4 20.7 36.2 20.0 40.7 20.2 38.9 19.5 37.3 18.9 37.9 20.7 36.5 20.2 34.9 19.7 37.8 21.0 36.0 20.2 34.2 19.6 38.0 21.3 36.4 20.7 34.8 20.0 38.5 20.4 36.8 19.7 34.9 19.2 39.0 21.1 37.4 20.4 35.5 19.7 28.4 16.7 25.9 16.1 23.9 15.9 39.4 20.9 37.9 20.8 36.3 20.4 28.7 18.2 27.0 18.6 25.7 18.7 28.1 17.0 25.5 16.6 23.5 16.4 31.7 17.8 29.0 17.3 27.3 17.1 28.3 17.4 26.5 17.6 25.0 17.3 29.0 16.6 26.7 16.3 24.8 16.0 34.8 19.1 32.7 18.7 30.7 18.0 38.2 18.5 36.9 18.2 35.6 17.7 38.4 21.1 36.6 20.5 34.9 20.0 37.2 19.8 35.0 19.2 32.9 18.7 37.8 22.1 37.1 21.6 35.9 21.1 39.4 18.8 37.9 18.0 36.7 17.5
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 23.2 36.1 22.2 35.5 20.9 34.2 19.9 32.4 18.5 24.3 17.6 23.2 22.2 28.4 21.4 27.2 21.1 25.2 20.4 24.3 23.0 34.0 22.2 32.8 22.3 34.3 21.5 33.2 21.2 25.1 20.5 24.1 23.0 36.1 22.0 35.2 20.5 29.9 19.7 28.4 21.9 28.6 21.3 27.6 22.1 35.8 21.2 34.5 20.1 28.1 19.3 26.6 17.2 22.2 16.6 21.2 21.8 28.2 21.0 26.8 20.2 30.0 19.3 28.2 21.8 25.4 21.1 24.8 19.2 25.6 18.4 24.1 25.9 36.8 25.1 36.2 N/A N/A N/A N/A 20.8 24.1 20.0 23.3 22.7 35.9 21.7 34.6 21.9 35.9 21.1 34.9 22.7 34.1 21.9 33.1 22.4 35.3 21.3 33.7 22.7 35.7 21.7 34.3 21.6 36.0 20.7 34.2 22.3 36.5 21.4 35.0 18.3 25.3 17.6 23.6 23.6 35.0 22.8 34.2 21.7 25.4 21.1 24.6 18.7 25.1 17.9 23.5 19.6 28.2 18.8 26.8 20.2 25.0 19.5 24.1 18.9 25.5 18.0 24.3 20.6 32.2 19.7 30.6 21.0 31.4 20.2 31.2 22.9 35.6 21.5 34.5 21.3 34.0 20.3 32.3 23.7 35.2 22.8 34.1 20.2 35.7 19.4 35.0
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 18.7 13.6 30.8 17.2 12.4 29.3 16.1 11.6 24.5 15.1 10.9 22.5 16.2 11.6 20.8 15.2 10.8 19.9 20.3 15.0 24.4 19.5 14.2 23.9 19.6 14.4 23.2 18.9 13.8 22.6 19.8 15.0 26.8 18.7 14.0 25.7 18.8 14.4 25.6 17.7 13.4 24.0 19.7 14.6 23.5 19.0 14.0 22.9 18.3 13.5 30.8 17.0 12.3 29.5 17.1 12.2 24.2 16.3 11.6 23.3 19.8 14.8 24.6 19.0 14.0 23.8 17.0 12.1 28.6 15.8 11.3 27.0 17.2 12.3 23.3 16.3 11.6 22.2 15.2 10.9 18.5 14.4 10.3 17.8 19.6 14.5 24.7 18.9 13.9 23.8 16.3 11.6 23.6 15.7 11.2 22.7 20.6 15.3 23.8 19.7 14.4 23.3 17.0 12.2 20.7 16.2 11.5 20.0 22.8 17.8 32.6 21.8 16.7 32.8 N/A N/A N/A N/A N/A N/A 19.7 14.5 22.4 18.6 13.4 21.9 17.6 12.8 29.9 17.0 12.3 29.3 17.5 12.7 26.6 16.3 11.8 25.6 18.9 14.1 27.5 17.7 13.1 26.1 17.5 12.6 28.6 16.4 11.7 26.4 17.8 12.8 29.1 17.0 12.1 28.1 16.3 11.6 24.1 15.5 11.1 24.3 17.3 12.4 26.4 16.3 11.6 26.0 15.9 11.3 19.8 15.1 10.8 19.1 20.0 15.4 28.4 19.0 14.4 28.5 20.3 15.0 23.7 19.7 14.4 23.2 16.3 11.6 20.2 15.6 11.1 19.6 16.4 11.7 21.8 15.8 11.3 21.2 18.3 13.2 22.0 17.6 12.6 21.2 16.3 11.7 20.6 15.5 11.1 19.9 16.1 11.5 24.3 15.1 10.8 23.0 18.1 14.5 25.0 16.6 13.2 25.9 18.1 13.0 29.3 16.5 11.7 26.8 17.0 12.2 23.1 16.0 11.4 22.3 19.9 14.8 29.6 18.6 13.6 29.2 14.5 11.2 27.0 12.9 10.1 27.0
36.333N 119.950W 37.693N 121.814W 34.667N 120.467W 33.812N 118.146W 33.938N 118.389W 34.056N 117.600W 33.900N 117.250W 33.128N 117.279W 35.434N 119.054W 37.359N 121.924W 32.867N 117.133W 37.624N 120.951W 37.406N 122.048W 36.588N 121.845W 32.816N 117.139W 38.210N 122.285W 32.700N 117.200W 37.721N 122.221W 33.822N 116.504W 34.580N 120.650W 34.117N 119.117W 36.029N 119.063W 40.518N 122.299W 33.952N 117.439W 38.507N 121.495W 38.696N 121.590W 38.567N 121.300W 38.667N 121.400W 36.664N 121.608W 34.095N 117.235W 32.734N 117.183W 37.620N 122.365W 35.237N 120.641W 34.426N 119.843W 34.899N 120.449W 38.504N 122.810W 34.583N 117.383W 37.889N 121.226W 38.267N 121.933W 36.317N 119.400W 34.741N 118.212W
71 120 27 10 30 289 468 100 149 16 145 22 12 50 127 4 8 2 125 32 4 135 152 245 5 7 30 24 23 353 5 2 61 3 74 35 879 8 19 90 713
99.6% -2.2 -0.8 0.2 5.3 7.1 3.6 0.1 6.1 0.3 1.9 3.9 -0.7 2.2 2.6 4.8 -1.6 7.1 2.5 5.1 7.6 3.8 -1.0 -2.0 2.5 -0.6 -0.9 -2.4 -1.0 1.0 1.1 7.2 4.1 1.1 1.6 0.7 -1.7 -2.5 -1.0 -1.1 -1.2 -6.0
99% -0.5 0.6 2.0 6.5 8.2 4.8 1.7 7.1 1.9 3.2 5.2 0.7 3.6 3.8 6.1 -0.1 7.9 3.8 6.4 8.6 5.0 0.8 -0.6 3.7 0.9 0.7 -1.1 0.6 2.4 2.5 8.3 5.3 2.4 2.8 2.0 -0.5 -0.9 0.4 0.6 0.2 -4.0
39.717N 104.750W 39.570N 104.849W 38.810N 104.688W 39.833N 104.658W 39.750N 104.867W 40.588N 105.042W 40.450N 105.017W 39.134N 108.540W 40.436N 104.632W 38.290N 104.498W
1726 1793 1884 1650 1612 1505 1529 1481 1432 1439
-17.5 -17.8 -17.0 -18.2 -18.6 -19.2 -17.9 -15.7 -22.1 -17.8
-13.8 -14.6 -13.9 -14.8 -14.9 -15.1 -14.7 -12.6 -17.8 -14.1
34.0 33.3 32.7 34.9 34.4 32.3 34.1 36.5 35.9 36.9
14.7 15.3 14.8 15.4 15.9 16.1 16.1 16.2 17.0 16.8
32.6 32.2 31.3 33.5 32.9 30.7 32.7 35.1 33.7 35.5
14.7 15.1 14.7 15.4 15.6 15.8 16.0 15.7 16.8 16.6
31.1 30.7 29.8 31.9 31.4 29.1 31.3 33.7 32.2 33.9
14.6 14.9 14.6 15.3 15.3 15.6 15.8 15.4 16.7 16.4
17.7 18.2 17.5 18.2 18.0 18.1 18.6 18.4 19.5 19.3
25.5 26.9 25.8 27.2 27.7 27.1 28.0 29.4 29.3 29.4
16.9 17.3 16.8 17.6 17.4 17.6 17.9 17.8 18.8 18.8
25.2 26.0 25.2 26.9 27.0 26.7 27.8 28.8 28.9 29.0
16.0 16.1 15.3 15.9 15.6 15.4 16.1 15.9 17.0 17.1
14.1 14.2 13.7 13.9 13.5 13.2 13.8 13.6 14.4 14.6
18.8 20.1 18.9 20.0 19.5 21.0 21.1 20.0 22.9 20.6
14.9 14.8 14.5 15.0 14.7 14.7 14.8 14.8 16.0 16.3
13.1 13.1 13.0 13.1 12.7 12.6 12.6 12.6 13.6 13.8
18.6 19.7 18.6 19.8 19.4 20.6 20.9 20.3 22.3 20.6
72.683W 58 72.651W 6 73.129W 2 73.133W 221 72.184W 75
-15.1 -13.2 -11.4 -16.0 -15.8
-12.3 -11.0 -9.1 -12.9 -12.6
33.0 32.7 31.1 30.9 32.1
22.9 22.8 22.8 22.7 22.9
31.3 31.2 29.4 28.7 30.3
22.1 22.3 22.1 21.7 22.2
29.8 29.5 28.0 27.4 28.8
21.3 21.5 21.4 20.8 21.4
24.6 24.8 24.5 24.1 24.5
30.3 29.8 28.5 28.5 29.4
23.7 24.0 23.8 23.2 23.7
28.8 28.5 27.2 26.9 27.9
22.9 23.2 23.2 22.7 22.9
17.7 18.0 18.0 17.9 17.8
26.8 26.8 26.4 26.1 26.4
22.2 22.6 22.7 22.2 22.4
17.0 17.3 17.4 17.4 17.3
26.1 26.2 25.7 25.5 25.8
41.938N 41.736N 41.158N 41.483N 41.742N
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 9.3 8.0 7.0 1276 1020 8.6 8.0 7.3 1484 464 9.1 8.3 7.6 1561 33 7.5 6.3 5.5 660 610 8.9 7.8 7.1 715 341 9.7 8.0 7.1 767 998 8.1 7.0 5.8 1053 880 6.0 5.4 4.8 902 298 8.2 7.0 5.8 1139 1288 8.7 8.0 7.2 1191 347 6.8 5.6 5.0 827 469 8.5 7.6 6.9 1312 906 8.4 7.6 6.9 1217 259 7.5 6.5 5.6 1795 28 7.0 5.8 5.3 814 474 9.5 8.5 7.8 1768 134 8.4 7.4 6.5 621 435 10.5 9.0 8.2 1496 91 10.3 8.9 8.0 421 2441 18.6 15.6 14.2 1892 13 10.4 8.5 7.4 1154 118 5.8 5.1 4.6 1400 947 11.2 8.8 7.5 1501 1049 8.5 7.2 5.9 804 954 9.0 8.0 7.1 1384 679 10.7 8.8 7.8 1387 758 9.2 7.6 6.1 1539 661 9.3 7.9 6.7 1291 882 9.3 8.3 7.6 1504 62 7.5 5.8 4.9 918 1006 7.6 6.8 5.8 651 397 12.7 11.4 10.4 1488 87 11.2 10.0 8.8 1235 163 8.5 7.3 5.9 1236 121 10.8 9.2 8.2 1482 64 7.7 6.7 5.6 1659 203 10.2 8.4 7.5 1478 1062 10.1 8.6 7.8 1360 774 12.9 12.0 11.2 1398 552 6.9 5.6 4.9 1394 912 13.3 12.1 11.1 1632 1050 10 sites, 30 more on CD-ROM 10.8 9.0 7.8 3221 382 11.0 9.3 8.2 3362 349 12.6 11.1 9.4 3398 281 12.1 10.6 8.9 3297 451 10.9 8.8 7.7 3148 401 8.9 7.5 6.1 3387 257 11.6 9.6 8.0 3442 347 10.5 8.7 7.6 3045 682 12.6 10.7 8.7 3655 357 12.8 11.1 9.1 3049 528 5 sites, 5 more on CD-ROM 10.1 8.6 7.8 3246 438 8.6 7.8 7.0 3056 479 10.9 9.2 8.3 2906 477 8.7 7.6 6.6 3564 260 8.6 7.6 6.8 3306 357
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station Delaware DOVER AFB NEW CASTLE AP Florida CECIL FIELD CRAIG MUNICIPAL DAYTONA BEACH INTL FT LAUDERDALE HOLLYWOOD INTL GAINESVILLE REGIONAL HOMESTEAD AFB JACKSONVILLE INTL JACKSONVILLE NAS KENNEDY SPACE CENTER MACDILL AFB MAYPORT NAF MELBOURNE INTL MIAMI EXECUTIVE MIAMI INTL NAPLES MUNICIPAL ORLANDO EXECUTIVE ORLANDO INTL ORLANDO SANFORD INTL PAGE FIELD PALM BEACH INTL PANAMA CITY BAY COUNTY INTL PENSACOLA INTL PENSACOLA NAS SARASOTA BRADENTON INTL SOUTHWEST FLORIDA INTL ST PETE-CLEARWATER INTL TALLAHASSEE REGIONAL TAMPA INTL TAYLOR FIELD TYNDALL AFB VENICE VERO BEACH REGIONAL Georgia ATHENS BEN EPPS AP ATLANTA HARTSFIELD-JACKSON ATLANTA REGIONAL AUGUSTA REGIONAL COLUMBUS AP DANIEL FIELD DEKALB-PEACHTREE AP DOBBINS AFB FULTON COUNTY AP HUNTER AAF LAWSON AAF LEE GILMER MEMORIAL MIDDLE GEORGIA REGIONAL MOODY AFB RICHARD B RUSSELL REGIONAL ROBINS AFB SAVANNAH HILTON HEAD INTL
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB 99.6%
99%
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
39.133N 75.467W 39.673N 75.601W
9 24
-9.6 -10.1
-7.4 -7.9
33.6 33.5
24.3 23.9
32.2 31.9
23.8 23.2
30.7 30.5
23.3 22.7
26.0 25.6
30.6 30.8
25.3 24.8
29.7 29.5
24.8 24.1
19.9 19.1
27.9 27.7
24.0 23.5
18.9 18.3
27.2 27.0
30.219N 30.336N 29.183N 26.072N 29.692N 25.483N 30.495N 30.233N 28.617N 27.850N 30.400N 28.101N 25.648N 25.791N 26.152N 28.545N 28.434N 28.780N 26.585N 26.685N 30.212N 30.478N 30.350N 27.401N 26.536N 27.911N 30.393N 27.962N 29.167N 30.067N 27.070N 27.651N
81.876W 81.515W 81.048W 80.154W 82.276W 80.383W 81.694W 81.667W 80.683W 82.517W 81.417W 80.644W 80.433W 80.316W 81.775W 81.333W 81.325W 81.244W 81.861W 80.099W 85.683W 87.187W 87.317W 82.559W 81.755W 82.688W 84.353W 82.540W 82.233W 85.583W 82.450W 80.420W
25 13 10 3 38 2 8 6 3 4 5 8 3 9 3 33 27 17 5 6 6 34 9 9 10 3 17 6 27 5 0 9
-1.0 0.4 2.0 8.8 -1.3 7.8 -1.4 1.2 4.0 3.8 1.8 3.7 7.6 9.3 6.5 3.9 3.5 2.7 6.0 7.0 -0.1 -1.1 -1.3 4.2 5.1 5.7 -3.2 4.2 -1.3 -0.3 5.4 3.8
1.1 2.3 4.2 11.1 0.7 10.1 0.4 3.1 6.2 6.1 3.9 6.1 9.8 11.5 8.6 6.2 5.7 4.9 8.1 9.2 2.1 1.0 0.6 6.7 7.4 7.5 -1.3 6.3 1.0 1.9 7.7 6.2
35.6 34.6 33.7 33.2 34.3 32.9 34.7 35.4 33.2 34.1 34.4 33.3 33.8 33.3 33.1 34.2 34.3 34.7 34.2 33.2 33.8 34.4 33.9 33.5 34.1 33.4 35.6 33.6 34.0 32.9 31.2 33.2
24.8 25.0 25.0 25.7 24.4 26.2 25.1 24.9 25.6 25.8 25.1 25.3 25.5 25.3 25.4 24.5 24.7 24.3 24.8 25.4 24.9 25.3 26.0 25.8 24.8 25.4 24.5 25.0 24.0 26.0 24.7 25.3
34.4 33.4 32.7 32.6 33.4 32.4 33.7 34.3 32.5 33.4 33.1 32.5 33.0 32.7 32.5 33.5 33.5 33.8 33.5 32.5 32.8 33.3 32.9 32.8 33.5 32.8 34.6 33.0 33.0 32.3 30.5 32.5
24.6 24.9 24.9 25.7 24.3 26.1 24.9 24.6 25.6 25.6 25.1 25.3 25.4 25.3 25.4 24.4 24.6 24.2 24.8 25.4 24.9 25.2 25.8 25.8 24.8 25.3 24.2 25.0 24.0 26.0 25.0 25.4
33.3 32.5 31.8 32.1 32.6 32.0 32.7 33.3 31.8 32.8 32.2 31.9 32.5 32.1 32.0 32.8 32.8 32.9 32.9 31.9 32.3 32.4 32.2 32.4 32.8 32.4 33.6 32.4 32.6 31.7 30.1 31.9
24.4 24.8 24.8 25.7 24.2 26.0 24.7 24.4 25.5 25.5 24.9 25.3 25.3 25.3 25.4 24.3 24.4 24.1 24.8 25.4 24.9 25.0 25.6 25.7 24.7 25.2 24.1 25.0 23.9 25.9 25.1 25.4
26.4 26.7 26.7 27.3 26.4 27.4 26.6 26.9 27.1 27.9 26.9 26.8 26.9 26.8 27.1 26.5 26.4 26.1 26.8 26.8 27.4 27.2 27.7 28.1 26.8 27.5 26.5 26.9 25.9 28.0 27.9 26.9
32.0 31.6 31.2 31.1 31.1 30.8 31.9 31.3 30.8 31.5 31.3 31.0 31.1 30.5 30.8 30.4 30.7 31.1 31.1 31.0 30.5 31.3 31.4 31.4 30.8 30.5 31.7 31.0 31.0 30.8 28.7 31.1
25.9 26.2 26.2 26.9 25.8 27.0 26.2 26.3 26.6 27.4 26.4 26.4 26.5 26.5 26.7 26.0 26.0 25.7 26.4 26.4 26.8 26.7 27.1 27.4 26.3 26.9 26.0 26.6 25.5 27.5 27.2 26.5
31.4 31.0 30.6 30.7 30.4 30.6 31.2 31.0 30.3 31.1 30.9 30.6 30.8 30.5 30.6 30.0 30.3 30.6 30.7 30.7 30.2 30.8 30.9 30.9 30.4 30.2 31.0 30.9 30.8 30.4 28.7 30.7
25.1 25.3 25.3 26.3 25.2 26.4 25.2 25.9 26.2 27.0 25.9 25.8 26.0 25.8 26.2 25.8 25.4 25.0 25.7 25.6 26.4 26.2 26.7 27.3 26.0 26.7 25.2 25.8 24.3 27.4 27.7 25.8
20.3 20.5 20.5 21.7 20.4 21.8 20.4 21.3 21.6 22.7 21.3 21.2 21.4 21.1 21.6 21.2 20.7 20.1 20.9 20.8 21.9 21.7 22.3 23.2 21.4 22.2 20.4 21.1 19.3 23.2 23.6 21.1
28.1 28.6 28.7 29.3 28.1 28.9 28.5 28.7 28.7 29.5 29.4 28.9 28.4 28.6 29.0 28.0 27.6 27.8 28.4 28.7 28.8 29.1 29.7 30.3 28.2 29.0 28.2 29.4 28.0 29.6 28.3 29.0
24.5 24.9 25.0 25.9 24.7 26.1 24.8 25.2 25.7 26.4 25.1 25.1 25.3 25.3 25.7 25.0 25.0 24.4 25.2 25.2 26.0 25.6 26.1 26.4 25.2 26.1 24.7 25.3 24.0 26.7 26.8 25.2
19.6 20.0 20.0 21.2 19.8 21.6 19.9 20.3 21.0 21.8 20.2 20.3 20.4 20.5 20.9 20.2 20.2 19.4 20.3 20.3 21.4 20.9 21.5 21.9 20.4 21.5 19.8 20.4 18.9 22.3 22.4 20.4
27.9 28.3 28.3 29.1 27.7 28.8 28.1 28.4 28.5 29.2 28.9 28.6 28.3 28.5 28.7 27.6 27.3 27.5 28.1 28.5 28.7 28.7 29.3 29.5 27.9 28.7 27.8 28.9 27.8 29.4 28.0 28.7
33.948N 33.640N 33.355N 33.364N 32.516N 33.467N 33.875N 33.917N 33.779N 32.017N 32.332N 34.272N 32.685N 30.967N 34.348N 32.633N 32.130N
83.328W 84.430W 84.567W 81.963W 84.942W 82.039W 84.302W 84.517W 84.521W 81.133W 84.989W 83.830W 83.653W 83.200W 85.161W 83.600W 81.210W
239 313 243 40 120 129 305 326 256 13 71 389 105 71 195 90 14
-5.2 -5.6 -7.0 -5.2 -3.4 -2.7 -6.0 -6.9 -6.0 -2.3 -5.3 -6.0 -4.5 -1.6 -7.0 -4.0 -2.5
-3.1 -3.0 -4.8 -3.3 -1.4 -1.1 -3.6 -4.1 -3.5 -0.1 -3.4 -3.3 -2.5 0.4 -4.8 -2.2 -0.6
35.2 34.4 34.1 36.3 35.8 36.0 34.5 34.0 34.5 35.3 36.0 33.6 35.9 35.6 35.4 36.1 35.3
23.7 23.4 23.2 24.4 23.7 23.6 23.1 23.5 23.5 25.2 24.4 23.0 24.1 24.7 23.6 24.4 25.0
33.9 33.1 33.0 34.9 34.6 34.4 33.2 32.8 33.3 34.1 34.8 32.5 34.7 34.6 34.0 34.9 34.0
23.3 23.2 23.1 24.2 23.5 23.1 23.0 23.4 23.3 24.9 24.3 22.9 23.9 24.6 23.3 24.3 24.9
32.6 32.0 32.1 33.7 33.5 33.2 32.3 31.8 32.3 32.9 33.6 31.3 33.5 33.7 32.8 33.7 32.9
23.1 22.9 23.0 24.0 23.4 23.0 22.6 23.1 23.0 24.8 24.2 22.5 23.7 24.3 23.2 24.0 24.6
25.4 25.2 25.3 26.3 25.6 25.3 24.9 25.2 25.2 27.3 26.8 24.7 26.1 26.7 25.6 26.5 26.8
31.6 31.3 30.8 32.6 31.8 31.7 31.1 30.9 31.3 31.3 32.0 30.2 32.2 32.5 32.0 32.5 31.9
24.9 24.6 24.7 25.7 25.2 24.9 24.4 24.6 24.7 26.6 26.1 24.1 25.6 26.1 25.0 25.8 26.2
30.8 30.3 30.0 31.7 31.1 31.1 30.2 30.2 30.5 31.1 31.7 29.4 31.5 31.9 31.3 31.6 31.3
23.7 23.5 23.8 24.7 24.1 23.8 23.0 23.6 23.6 26.3 25.5 23.0 24.6 25.2 23.9 25.1 25.5
19.1 19.0 19.3 19.9 19.2 18.9 18.4 19.1 19.0 21.7 20.9 18.6 19.8 20.5 19.2 20.4 20.8
27.7 27.4 27.3 28.5 27.8 27.1 26.3 27.3 27.5 28.6 28.7 26.5 28.4 28.8 27.8 28.6 28.6
23.2 23.0 23.0 24.2 23.6 23.0 22.8 23.0 23.0 25.3 24.8 22.6 24.0 24.7 23.2 24.3 25.0
18.5 18.4 18.3 19.2 18.7 18.1 18.1 18.5 18.3 20.5 20.0 18.2 19.2 19.8 18.4 19.4 20.1
27.0 26.8 26.4 28.0 27.3 26.6 26.1 26.8 27.0 28.3 28.1 26.1 27.8 28.3 27.3 28.0 28.1
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 2 sites, 3 more on CD-ROM 11.0 9.4 8.4 2478 663 11.1 9.3 8.3 2602 644 32 sites, 45 more on CD-ROM 8.5 7.5 6.7 647 1517 8.5 7.8 7.1 643 1525 9.1 8.0 7.2 404 1678 9.8 8.8 8.1 69 2561 8.2 7.3 6.3 638 1483 9.0 8.1 7.3 80 2349 9.0 8.0 7.2 724 1461 9.4 8.3 7.5 526 1811 8.4 7.5 6.6 292 1764 8.5 7.6 6.9 273 2029 9.6 8.4 7.5 559 1645 9.3 8.5 7.9 257 1929 9.2 8.4 7.8 91 2304 9.0 8.2 7.5 63 2543 8.5 7.6 6.8 152 2115 8.7 7.9 7.1 281 1968 9.1 8.1 7.4 294 1892 9.0 8.0 7.2 343 1857 8.3 7.5 6.7 148 2191 10.1 8.9 8.2 116 2299 8.3 7.5 6.7 690 1582 9.0 8.1 7.3 790 1503 9.2 8.2 7.4 819 1454 9.3 8.2 7.4 253 1932 9.1 8.1 7.2 170 2079 9.4 8.4 7.7 246 2040 8.1 7.2 6.1 829 1479 8.1 7.2 6.1 281 2005 8.0 6.9 5.6 579 1531 9.2 8.2 7.3 736 1469 12.5 10.6 8.9 265 1672 9.1 8.3 7.6 230 1922 19 sites, 27 more on CD-ROM 8.1 7.2 6.1 1531 1002 9.5 8.5 7.6 1467 1056 7.7 6.7 5.5 1668 877 8.4 7.4 6.3 1324 1152 8.2 7.3 6.3 1137 1311 7.4 6.5 5.6 1170 1294 8.3 7.4 6.3 1593 1009 8.5 7.5 6.5 1624 976 7.9 6.9 5.8 1559 981 8.6 7.5 6.6 894 1433 7.6 6.5 5.5 1284 1152 8.5 7.6 6.8 1670 913 8.2 7.2 6.0 1258 1196 8.0 6.8 5.7 796 1482 7.2 6.0 5.3 1683 987 8.7 7.5 6.4 1186 1238 8.5 7.5 6.8 957 1369
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station SW GEORGIA REGIONAL VALDOSTA REGIONAL Hawaii HILO INTL HONOLULU INTL KALAELOA KANEOHE MCAS Idaho BOISE AP CALDWELL INDUSTRIAL AP COEUR D'ALENE AP IDAHO FALLS REGIONAL LEWISTON-NEZ PERCE CO REGL MAGIC VALLEY REGIONAL POCATELLO REGIONAL Illinois ABRAHAM LINCOLN CAPITAL AURORA MUNICIPAL CHICAGO MIDWAY INTL CHICAGO O'HARE INTL CHICAGO ROCKFORD INTL DECATUR AP DUPAGE COUNT AP GLENVIEW NAS GREATER PEORIA REGIONAL QUAD CITY INTL QUINCY REGIONAL SCOTT AFB ST LOUIS DOWNTOWN AP U OF ILLINOIS WILLARD AP Indiana EVANSVILLE REGIONAL FORT WAYNE INTL GRISSOM AFB INDIANAPOLIS INTL MONROE COUNTY AP PURDUE UNIVERSITY AP SOUTH BEND INTL TERRE HAUTE INTL Iowa AMES MUNICIPAL ANKENY REGIONAL BOONE MUNICIPAL DAVENPORT MUNICIPAL DES MOINES INTL DUBUQUE REGIONAL SIOUX GATEWAY AP THE EASTERN IOWA AP WATERLOO MUNICIPAL Kansas COLONEL JAMES JABARA AP FORBES FIELD JOHNSON COUNTY EXECUTIVE LAWRENCE MUNICIPAL MANHATTAN REGIONAL MARSHALL AIRFIELD MCCONNELL AFB PHILIP BILLARD MUNICIPAL
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB Dehumidification DP/HR/MCDB 0.4% 1% 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB DP / HR / MCDB DP / HR / MCDB 36.0 24.4 34.8 24.3 33.7 24.1 26.5 32.4 25.9 31.6 25.1 20.3 28.4 24.4 19.5 27.9 35.9 24.8 34.7 24.6 33.7 24.3 26.8 32.2 26.2 31.6 25.3 20.7 28.4 24.9 20.2 28.1
99.6% -3.0 -2.4
99% -1.3 -0.6
19.719N 155.053W 12 21.324N 157.929W 2 21.317N 158.067W 10 21.450N 157.768W 7
16.4 17.0 15.8 17.8
17.1 18.1 17.0 18.8
29.8 31.9 32.3 29.4
23.3 23.2 23.0 23.5
29.3 31.4 31.6 29.0
23.2 23.0 22.9 23.4
28.8 30.9 31.1 28.5
23.0 22.8 22.8 23.2
24.7 25.1 25.2 25.0
27.8 29.2 29.4 27.5
24.4 24.5 24.6 24.5
27.5 28.8 29.0 27.4
23.9 23.8 23.9 24.1
18.8 18.7 18.8 19.0
26.2 27.2 27.5 26.5
23.4 23.2 23.0 23.5
18.2 17.9 17.8 18.4
25.8 26.8 26.9 26.4
43.567N 116.241W 43.650N 116.633W 47.767N 116.817W 43.516N 112.067W 46.375N 117.016W 42.482N 114.487W 42.920N 112.571W
858 740 703 1441 438 1265 1357
-12.6 -12.4 -14.5 -21.4 -10.6 -13.6 -18.9
-9.0 -9.1 -12.0 -17.9 -7.3 -11.1 -15.7
37.0 36.1 32.9 33.4 36.9 35.0 34.9
17.6 19.0 17.3 16.1 18.5 17.0 16.3
35.4 34.0 31.4 32.0 35.0 33.4 33.2
17.1 18.2 17.0 15.8 18.0 16.7 16.0
33.8 32.6 29.0 30.4 33.0 32.1 31.6
16.6 17.7 16.2 15.3 17.3 16.4 15.5
19.0 20.1 19.0 18.2 19.9 19.1 18.4
33.4 33.5 29.8 28.4 33.3 31.5 30.1
18.2 19.1 18.0 17.2 19.0 18.2 17.5
32.4 32.2 28.6 27.7 32.1 30.0 29.2
14.2 15.1 15.1 14.8 15.6 14.9 14.8
11.2 11.7 11.7 12.6 11.7 12.3 12.4
21.9 25.8 22.2 21.7 22.6 23.7 21.5
12.9 13.7 13.8 13.5 14.3 13.6 13.2
10.3 10.7 10.7 11.5 10.7 11.4 11.2
22.0 25.4 21.5 20.5 22.2 23.4 21.4
39.845N 41.770N 41.786N 41.995N 42.193N 39.834N 41.914N 42.083N 40.668N 41.465N 39.937N 38.550N 38.571N 40.040N
89.684W 88.481W 87.752W 87.934W 89.093W 88.866W 88.246W 87.817W 89.684W 90.523W 91.192W 89.850W 90.157W 88.278W
181 216 187 202 223 206 230 199 198 180 234 140 126 230
-17.2 -20.5 -17.5 -18.3 -20.8 -16.7 -19.2 -18.2 -18.3 -19.7 -17.6 -13.8 -13.0 -18.0
-13.9 -17.4 -14.4 -15.3 -17.6 -13.7 -16.4 -15.1 -15.4 -16.8 -14.7 -11.1 -10.7 -15.1
33.7 32.6 33.3 32.9 32.7 33.9 32.4 34.3 33.3 33.5 34.0 35.1 35.2 33.0
24.8 23.4 23.7 23.4 23.5 24.8 23.7 23.9 24.7 24.5 24.7 25.4 24.8 24.2
32.4 31.2 31.9 31.4 31.1 32.6 31.0 32.3 32.0 32.1 32.4 33.7 33.7 32.0
24.2 22.9 22.8 22.7 22.7 24.2 23.1 23.0 24.0 23.9 24.1 25.0 24.6 23.8
31.1 29.7 30.3 29.9 29.7 31.3 29.2 30.6 30.6 30.6 30.9 32.4 32.4 30.5
23.4 22.2 22.2 22.0 22.0 23.5 22.1 22.3 23.1 22.9 23.4 24.4 24.0 23.0
26.4 25.4 25.5 25.3 25.4 26.3 25.5 25.5 26.3 26.2 26.1 27.0 26.6 26.2
32.0 30.6 31.0 30.7 30.6 31.9 30.4 32.3 31.3 31.6 31.7 32.3 32.5 31.0
25.5 24.4 24.4 24.3 24.3 25.5 24.4 24.5 25.3 25.2 25.3 26.2 25.7 25.2
30.7 29.0 29.6 29.3 28.9 30.8 28.9 30.6 30.3 30.3 30.6 31.5 31.6 29.9
24.7 23.8 23.8 23.6 23.9 24.7 23.9 23.4 24.8 24.6 24.4 25.6 25.0 24.8
20.2 19.1 19.1 18.9 19.3 20.2 19.3 18.7 20.3 20.1 19.9 21.2 20.4 20.4
30.1 28.6 28.9 28.6 28.5 30.0 28.8 29.5 29.6 29.6 29.6 29.9 29.6 29.5
23.9 22.8 22.7 22.7 22.9 23.8 22.9 22.5 23.8 23.7 23.6 24.8 24.0 23.8
19.2 18.0 17.9 17.8 18.1 19.1 18.1 17.6 19.1 18.9 18.9 20.2 19.1 19.2
29.1 27.4 27.8 27.4 27.6 28.9 27.5 28.6 28.5 28.4 28.7 29.2 28.9 28.1
38.044N 40.971N 40.650N 39.732N 39.133N 40.412N 41.707N 39.452N
87.521W 85.206W 86.150W 86.279W 86.617W 86.937W 86.316W 87.309W
122 241 248 241 257 183 236 175
-12.8 -18.0 -18.5 -16.4 -15.6 -17.4 -17.6 -16.6
-9.7 -14.8 -14.7 -13.0 -12.3 -14.3 -14.3 -12.9
34.5 32.7 32.6 33.0 33.0 33.1 32.4 33.6
24.5 23.6 24.1 23.9 23.9 24.2 23.3 24.5
33.2 31.2 31.1 31.6 32.0 32.0 30.9 32.3
24.2 22.8 23.3 23.3 23.8 23.5 22.4 24.2
32.0 29.7 29.8 30.3 30.6 30.5 29.4 31.0
23.7 22.0 22.7 22.7 23.0 22.7 21.7 23.4
26.2 25.3 26.3 25.6 25.8 25.8 25.1 26.2
32.1 30.6 30.0 30.8 30.5 31.1 30.1 31.5
25.5 24.3 25.2 24.8 25.1 25.0 24.1 25.4
31.3 28.9 29.0 29.6 29.9 29.8 28.7 30.5
24.5 23.7 25.4 24.1 24.2 24.3 23.6 24.7
19.8 19.1 21.2 19.5 19.7 19.7 18.9 20.2
29.5 28.3 28.4 28.7 28.5 29.2 28.2 29.5
23.9 22.9 24.2 23.3 23.6 23.4 22.6 23.8
19.0 18.1 19.6 18.6 19.0 18.6 17.8 19.1
28.7 27.3 27.2 27.8 27.9 28.0 27.0 28.5
41.991N 41.691N 42.049N 41.614N 41.534N 42.398N 42.391N 41.883N 42.554N
93.619W 93.566W 93.848W 90.591W 93.653W 90.704W 96.379W 91.717W 92.401W
291 277 354 230 292 322 334 265 265
-21.4 -19.9 -20.9 -21.6 -20.1 -22.3 -21.7 -22.3 -23.1
-18.6 -17.2 -17.7 -18.6 -17.5 -19.3 -19.1 -19.2 -20.3
32.7 33.9 32.9 32.8 33.8 31.4 33.8 32.4 32.8
24.5 24.2 24.7 23.9 24.5 23.9 24.0 24.6 24.1
31.3 32.4 31.9 31.2 32.1 29.8 32.3 30.9 31.2
23.7 23.9 24.3 23.3 23.9 22.9 23.5 23.6 23.2
29.9 31.0 30.1 29.9 30.6 28.4 30.8 29.4 29.7
22.9 23.2 23.0 22.6 23.0 21.9 22.8 22.6 22.4
26.2 26.3 26.8 25.5 26.0 25.3 26.0 25.9 25.9
30.7 31.2 31.0 30.9 31.5 29.6 31.3 30.4 30.6
25.1 25.2 25.6 24.7 25.1 24.2 25.1 25.0 24.8
29.7 30.4 29.6 29.8 30.3 28.1 30.4 29.2 29.2
24.8 25.0 25.3 23.9 24.4 24.0 24.5 24.6 24.5
20.6 20.7 21.4 19.3 20.0 19.7 20.3 20.3 20.1
29.4 29.4 29.7 29.1 29.6 28.0 29.5 28.9 29.1
23.7 23.6 24.1 22.9 23.5 22.9 23.5 23.6 23.4
19.2 19.1 19.9 18.2 18.9 18.4 19.0 19.0 18.8
28.1 28.4 28.3 28.2 28.7 26.9 28.6 27.8 27.8
37.746N 38.950N 38.850N 39.008N 39.135N 39.050N 37.617N 39.073N
97.221W 95.664W 94.739W 95.212W 96.679W 96.767W 97.267W 95.626W
433 325 326 254 322 325 418 267
-14.1 -15.8 -15.4 -15.9 -16.7 -15.1 -13.2 -15.6
-11.7 -12.8 -12.7 -13.0 -13.6 -12.6 -10.6 -12.8
37.8 37.2 35.8 37.3 37.9 38.1 37.9 36.8
23.1 24.4 24.2 24.8 24.0 23.9 22.6 24.4
36.2 34.9 33.7 35.3 36.1 36.0 36.2 34.9
23.3 24.1 24.1 24.3 24.0 24.0 23.0 24.4
34.0 33.0 32.3 33.4 33.9 34.2 34.3 33.2
23.1 23.7 23.8 24.1 23.7 23.8 23.0 23.9
25.2 26.0 26.0 26.6 25.8 26.0 25.4 26.2
32.8 33.2 32.0 33.5 33.6 33.1 32.2 33.1
24.6 25.4 25.4 25.7 25.3 25.3 24.8 25.5
32.2 32.5 31.5 32.6 33.0 32.5 31.6 32.4
23.0 24.0 24.1 24.6 23.8 24.2 23.8 24.2
18.7 19.7 19.8 20.2 19.4 19.9 19.6 19.7
28.6 29.9 29.6 30.6 29.8 29.4 27.8 30.0
22.5 23.1 23.6 23.8 22.9 23.4 22.9 23.5
18.1 18.6 19.2 19.2 18.4 18.9 18.5 18.9
28.0 28.9 29.0 29.8 28.7 28.7 27.4 29.3
31.536N 84.194W 30.783N 83.277W
58 60
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 8.2 7.3 6.4 970 1415 7.4 6.3 5.5 821 1459 4 sites, 8 more on CD-ROM 7.5 6.7 5.7 0 1803 10.1 9.1 8.5 0 2587 8.1 7.3 6.6 0 2341 8.3 7.5 6.9 0 2328 7 sites, 14 more on CD-ROM 9.8 8.5 7.6 3008 559 9.9 8.6 7.6 3188 384 10.0 8.5 7.5 3819 176 12.2 10.9 9.3 4262 160 9.3 8.0 6.7 2802 482 12.5 11.0 9.3 3349 431 12.8 11.4 10.1 3856 244 14 sites, 48 more on CD-ROM 11.1 9.6 8.5 2960 636 11.6 10.3 8.9 3632 405 10.9 9.4 8.6 3250 587 11.0 9.4 8.5 3439 490 10.9 9.3 8.4 3661 437 11.1 9.6 8.7 2993 614 11.1 9.7 8.6 3572 417 9.0 8.0 7.2 3391 505 10.3 8.8 7.9 3185 587 10.7 9.0 8.1 3384 549 10.9 9.3 8.4 3054 622 10.2 8.8 7.8 2565 785 9.3 8.3 7.4 2538 796 12.3 11.0 9.6 3163 541 8 sites, 9 more on CD-ROM 9.2 8.1 7.3 2438 807 11.2 9.7 8.6 3321 459 11.6 10.1 8.7 3260 499 11.1 9.6 8.5 2916 614 8.7 7.8 7.0 2817 570 10.1 8.8 8.0 3070 554 10.7 9.1 8.3 3422 444 10.0 8.6 7.9 2872 602 9 sites, 46 more on CD-ROM 11.8 10.5 9.0 3681 449 10.2 8.8 7.9 3403 542 11.7 10.5 9.1 3602 488 11.9 10.7 9.2 3597 488 11.4 10.0 8.7 3415 592 11.5 10.2 8.9 3912 356 12.7 11.2 9.8 3741 508 11.9 10.6 9.1 3753 430 11.7 10.5 9.0 3898 427 10 sites, 22 more on CD-ROM 12.4 11.2 10.2 2502 907 11.7 10.6 9.2 2765 790 10.4 9.0 8.2 2687 785 11.2 9.8 8.6 2773 819 10.8 9.1 8.2 2858 826 10.9 9.2 8.3 2744 885 12.2 11.1 9.9 2380 969 10.4 8.9 8.1 2714 837
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station SALINA MUNICIPAL WICHITA EISENHOWER NATL Kentucky BLUE GRASS AP BOWLING GREEN AP BOWMAN FIELD CAMPBELL AAF CINCINNATI NORTHERN KY INTL HENDERSON CITY-COUNTY AP LAKE CUMBERLAND REGIONAL LOUISVILLE INTL Louisiana ALEXANDRIA ESLER REGIONAL ALEXANDRIA INTL BARKSDALE AFB BATON ROUGE METROPOLITAN LAFAYETTE REGIONAL LAKE CHARLES REGIONAL LAKEFRONT AP LOUIS ARMSTRONG NEW ORLEANS MONROE REGIONAL NEW ORLEANS NAS SHREVEPORT DOWNTOWN AP SHREVEPORT REGIONAL Maine AUBURN LEWISTON MUNICIPAL BANGOR INTL BRUNSWICK NAS PORTLAND INTL JETPORT SANFORD SEACOAST REGIONAL Maryland ANDREWS AFB BALTIMORE-WASHINGTON INTL THOMAS POINT Massachusetts BARNSTABLE MUNICIPAL BOSTON LOGAN INTL BUZZARDS BAY CHATHAM MUNICIPAL LAWRENCE MUNICIPAL MARTHAS VINEYARD AP NEW BEDFORD REGIONAL NORWOOD MEMORIAL PLYMOUTH MUNICIPAL SOUTH WEYMOUTH NAS WORCESTER REGIONAL Michigan BISHOP INTL DETROIT COLEMAN YOUNG INTL DETROIT METRO WAYNE COUNTY AP GERALD R FORD INTL GROSSE ILE MUNICIPAL JACKSON COUNTY AP KALAMAZOO BATTLE CREEK INTL LANSING CAPITAL REGION INTL MBS INTL MUSKEGON COUNTY AP OAKLAND COUNTY INTL
Lat
Long
Elev
38.800N 97.650W 387 37.648N 97.430W 403
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB Dehumidification DP/HR/MCDB 0.4% 1% 0.4% 1% 2% 0.4% 1% 99% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB DP / HR / MCDB DP / HR / MCDB -12.7 38.9 23.1 37.1 23.2 35.1 22.9 25.2 33.6 24.5 32.8 22.9 18.4 28.8 22.3 17.8 28.3 -11.0 38.4 22.9 36.5 23.1 34.7 23.1 25.4 32.6 24.8 32.1 23.4 19.2 28.7 22.7 18.3 27.9
Heating DB 99.6% -15.4 -13.5
38.041N 36.965N 38.228N 36.667N 39.043N 37.800N 37.054N 38.181N
84.606W 86.424W 85.664W 87.483W 84.672W 87.683W 84.615W 85.739W
299 161 165 175 265 118 283 149
-12.9 -10.9 -11.8 -11.1 -14.5 -12.9 -11.0 -11.7
-9.9 -8.0 -8.7 -7.9 -11.2 -9.8 -7.8 -8.6
33.1 34.5 34.1 34.8 33.1 34.1 34.5 34.5
23.4 23.9 24.0 24.5 23.5 24.6 23.7 24.0
31.8 33.1 32.9 33.4 31.7 32.9 32.9 33.2
23.1 23.9 23.7 24.3 23.0 24.4 23.3 23.9
30.5 32.0 31.8 32.3 30.4 32.3 32.2 32.0
22.6 23.6 23.3 24.0 22.5 24.0 23.0 23.4
25.2 25.8 25.8 26.3 25.3 26.3 25.5 26.0
30.7 31.6 31.3 31.8 30.6 32.9 32.2 31.8
24.5 25.2 25.2 25.7 24.5 25.6 24.7 25.3
29.6 30.8 30.7 31.0 29.4 31.8 31.0 31.1
23.5 24.1 24.2 25.0 23.7 24.1 23.0 24.3
19.0 19.4 19.4 20.5 19.1 19.3 18.4 19.6
28.1 28.8 28.4 28.5 28.1 30.5 28.1 29.4
22.8 23.5 23.6 24.2 23.0 23.6 22.7 23.5
18.2 18.7 18.8 19.5 18.3 18.7 18.1 18.7
27.2 28.2 28.0 28.1 27.1 29.9 27.7 28.5
31.395N 31.335N 32.500N 30.537N 30.205N 30.125N 30.049N 29.993N 32.516N 29.817N 32.543N 32.447N
92.291W 92.559W 93.667W 91.147W 91.988W 93.228W 90.029W 90.251W 92.041W 90.017W 93.745W 93.824W
36 26 51 20 12 3 3 1 24 1 55 77
-3.3 -2.8 -4.1 -2.0 -1.0 -0.8 2.2 0.7 -3.7 -0.4 -3.1 -3.4
-2.0 -1.2 -2.4 -0.2 1.0 1.0 4.0 2.5 -2.0 1.4 -1.6 -1.7
36.7 36.3 37.2 34.9 34.9 34.9 34.2 34.5 36.8 33.9 37.5 37.3
24.8 24.9 24.2 25.2 25.4 25.4 25.9 25.5 25.4 25.4 24.6 24.3
35.3 35.1 35.8 34.0 33.9 33.9 33.7 33.6 35.5 33.1 36.1 35.9
25.0 25.0 24.3 25.2 25.3 25.4 25.8 25.4 25.2 25.2 24.5 24.4
34.2 33.9 34.4 33.2 33.0 33.0 32.8 32.8 34.2 32.4 34.8 34.6
24.9 24.8 24.4 25.0 25.1 25.3 25.5 25.2 25.0 25.1 24.4 24.4
26.8 26.9 26.3 26.9 27.0 27.4 27.5 27.0 27.1 27.3 26.3 26.3
32.2 32.0 32.6 31.5 31.5 31.4 32.1 31.6 33.1 30.8 32.9 32.9
26.4 26.4 25.8 26.5 26.6 26.9 27.0 26.7 26.6 26.7 25.9 25.8
32.0 31.9 32.0 31.2 31.2 30.9 31.6 31.2 32.7 30.5 32.1 32.2
25.6 25.8 24.7 25.8 26.0 26.3 26.3 25.8 25.6 26.3 24.8 24.6
20.9 21.2 19.9 21.1 21.4 21.8 21.7 21.1 20.9 21.8 19.9 19.8
28.7 28.9 28.5 28.7 28.7 29.0 30.0 29.1 29.6 28.7 28.3 28.5
25.0 25.0 24.1 25.3 25.4 25.9 25.9 25.5 25.1 25.8 24.2 24.2
20.2 20.2 19.2 20.5 20.6 21.2 21.3 20.7 20.2 21.1 19.2 19.3
28.5 28.6 28.0 28.4 28.5 28.7 29.8 28.9 29.2 28.5 28.1 28.2
44.050N 44.798N 43.900N 43.650N 43.394N
70.283W 68.819W 69.933W 70.300W 70.708W
88 45 21 14 74
-21.1 -21.7 -19.0 -17.5 -21.1
-17.7 -18.8 -16.5 -14.9 -17.6
31.0 30.9 30.1 30.4 32.1
21.8 21.4 21.3 21.8 22.1
28.7 29.0 28.2 28.5 29.8
20.9 20.5 20.5 21.0 21.1
27.3 27.3 26.8 26.8 27.9
19.8 19.4 19.5 20.1 20.1
23.2 22.8 23.0 23.3 23.6
28.6 28.5 27.8 28.4 29.6
22.0 21.8 21.9 22.3 22.4
26.9 27.0 26.5 26.7 28.0
21.4 20.9 21.3 21.6 22.0
16.2 15.7 16.0 16.2 16.8
25.9 25.5 25.5 25.8 27.2
20.9 20.0 20.4 20.7 21.0
15.7 14.8 15.1 15.4 15.8
25.2 24.1 24.3 24.5 25.9
38.817N 76.867W 39.167N 76.683W 38.900N 76.440W
86 48 0
-9.4 -9.7 -7.7
-7.4 -7.6 -5.5
34.5 34.5 30.5
23.9 23.8 24.7
32.8 33.0 29.4
23.3 23.4 24.4
31.4 31.5 28.4
22.7 22.6 24.0
25.6 25.4 27.7
31.4 31.6 28.5
24.8 24.7 26.6
30.2 30.2 27.7
24.0 23.8 27.5
19.0 18.8 23.4
27.7 27.6 28.1
23.0 23.2 26.3
18.0 18.0 21.8
26.8 26.8 27.2
41.669N 42.361N 41.400N 41.688N 42.717N 41.393N 41.676N 42.191N 41.910N 42.150N 42.271N
70.280W 17 71.011W 4 71.030W 0 69.993W 21 71.124W 45 70.615W 21 70.958W 24 71.174W 15 70.729W 45 70.933W 49 71.873W 305
-12.1 -13.1 -10.5 -11.0 -15.0 -12.4 -13.0 -16.1 -14.4 -14.5 -16.4
-9.2 -10.3 -8.2 -8.1 -12.3 -9.8 -10.9 -12.8 -12.1 -12.0 -13.8
29.0 32.6 24.7 28.3 32.5 28.9 31.2 32.6 31.5 32.9 29.9
22.7 22.6 N/A 22.6 22.8 22.5 23.0 23.1 22.8 23.2 21.6
27.5 30.9 23.7 27.0 31.0 27.5 29.1 31.1 29.6 31.0 28.4
22.0 22.0 N/A 22.1 22.2 21.9 22.0 22.5 22.1 22.4 20.8
26.3 29.1 23.0 25.8 29.0 26.2 27.8 29.1 27.9 29.3 27.0
21.3 21.1 N/A 21.4 21.2 21.2 21.2 21.3 21.1 21.5 20.0
24.2 24.3 N/A 24.0 24.3 24.1 24.4 24.8 24.3 25.0 23.2
27.3 29.8 N/A 26.7 29.7 27.2 28.9 30.0 29.0 30.5 27.5
23.5 23.4 N/A 23.3 23.5 23.3 23.6 23.9 23.5 23.8 22.4
26.0 28.3 N/A 25.6 28.5 26.0 27.2 28.4 27.4 28.8 26.3
23.0 22.6 N/A 23.0 22.6 22.9 22.9 23.0 22.9 23.4 21.9
17.8 17.3 N/A 17.7 17.4 17.7 17.7 17.8 17.7 18.3 17.2
25.4 26.9 N/A 25.3 26.6 25.6 26.1 26.6 26.1 27.7 25.0
22.6 21.8 N/A 22.5 22.1 22.5 22.5 22.5 22.4 22.3 21.1
17.4 16.5 N/A 17.2 16.9 17.2 17.2 17.2 17.2 17.1 16.4
25.0 25.9 N/A 24.8 26.3 24.9 25.5 26.1 25.6 26.3 24.3
42.967N 42.409N 42.231N 42.883N 42.099N 42.267N 42.235N 42.780N 43.533N 43.171N 42.665N
83.749W 83.010W 83.331W 85.524W 83.161W 84.467W 85.552W 84.579W 84.080W 86.237W 83.418W
-17.9 -14.9 -16.0 -16.2 -14.1 -17.5 -16.4 -18.0 -17.6 -14.7 -17.2
-15.3 -12.4 -13.2 -13.7 -12.3 -14.8 -13.6 -15.2 -15.2 -12.5 -14.8
31.9 32.6 32.4 31.8 32.1 31.4 32.3 31.7 31.9 30.1 32.1
23.0 23.0 23.4 22.8 23.6 22.9 22.8 22.9 22.8 22.3 22.8
30.3 31.1 30.8 30.3 30.0 29.9 30.9 30.1 30.2 28.7 30.2
22.1 22.2 22.5 22.0 23.2 22.1 22.1 22.0 21.8 21.5 21.7
28.9 29.6 29.3 28.8 28.6 28.5 29.1 28.6 28.7 27.5 28.7
21.2 21.4 21.6 21.1 22.6 21.2 21.2 21.1 21.1 20.9 20.9
24.4 24.6 24.9 24.6 25.5 24.4 24.5 24.5 24.4 24.1 24.1
29.7 30.1 30.2 29.5 29.4 29.3 29.7 29.5 29.6 28.1 29.4
23.4 23.6 23.9 23.5 24.6 23.4 23.6 23.4 23.4 23.2 23.0
28.4 28.7 28.6 28.2 28.3 28.0 28.4 28.1 28.2 26.9 27.9
22.7 22.8 23.2 23.0 24.0 22.8 22.8 22.8 22.8 22.8 22.4
18.0 18.0 18.4 18.3 19.4 18.2 18.1 18.1 17.9 17.9 17.8
27.7 27.6 27.9 27.3 27.7 27.1 27.5 27.4 27.3 26.4 27.1
21.8 22.0 22.3 22.0 23.0 22.0 22.2 21.8 21.8 21.9 21.5
16.9 17.1 17.4 17.1 18.1 17.3 17.4 17.0 16.9 17.0 16.7
26.2 26.7 26.7 26.1 26.6 26.2 26.7 26.1 26.1 25.6 25.8
235 191 192 245 179 304 265 256 201 191 298
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 12.6 11.3 10.1 2666 948 12.7 11.5 10.5 2469 968 8 sites, 9 more on CD-ROM 9.3 8.2 7.4 2519 665 8.8 7.9 7.1 2211 819 8.3 7.4 6.4 2309 826 9.0 7.9 7.0 2096 897 9.7 8.5 7.6 2732 619 9.5 8.4 7.5 2485 776 8.0 6.9 5.6 2192 771 9.4 8.4 7.5 2254 891 12 sites, 16 more on CD-ROM 7.3 6.1 5.3 1104 1407 8.4 7.5 6.5 1028 1477 9.1 8.1 7.2 1243 1347 8.4 7.4 6.6 876 1512 9.1 8.2 7.4 802 1587 9.1 8.2 7.4 797 1583 11.7 10.3 8.9 596 1881 9.3 8.4 7.6 697 1667 8.6 7.6 6.7 1213 1392 8.3 7.3 6.3 764 1516 8.4 7.5 6.6 1207 1474 8.9 8.0 7.2 1163 1451 5 sites, 24 more on CD-ROM 9.3 8.3 7.3 4214 174 10.5 8.8 8.0 4230 200 10.5 8.9 7.9 3989 205 10.4 8.7 7.8 3850 212 9.4 8.3 7.3 4124 203 3 sites, 10 more on CD-ROM 10.8 9.2 8.2 2432 695 9.8 8.4 7.5 2495 704 16.8 14.1 11.9 2297 692 11 sites, 14 more on CD-ROM 11.0 9.5 8.5 3217 292 12.0 10.7 9.3 3062 431 19.7 17.1 15.1 3006 188 9.1 8.0 7.2 3110 274 9.1 8.1 7.3 3337 380 11.6 10.5 9.1 3209 251 10.3 8.8 7.9 3206 326 9.1 8.0 7.2 3411 337 10.4 8.9 8.0 3362 319 8.3 7.4 6.5 3240 359 11.6 10.3 8.9 3668 269 15 sites, 72 more on CD-ROM 10.7 9.1 8.2 3718 339 9.2 8.3 7.6 3321 491 11.2 9.6 8.5 3370 456 11.1 9.5 8.5 3636 368 9.4 8.4 7.5 3271 477 9.0 8.2 7.5 3664 323 9.7 8.5 7.7 3493 406 10.8 9.1 8.3 3763 322 11.0 9.4 8.5 3808 326 11.2 10.1 8.9 3647 299 10.8 9.2 8.3 3684 363
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station SELFRIDGE ANGB ST CLAIR COUNTY INTL WEST MICHIGAN REGIONAL WILLOW RUN AP Minnesota ANOKA COUNTY AP CRYSTAL AP DULUTH INTL FLYING CLOUD AP MANKATO MUNICIPAL MINNEAPOLIS-ST PAUL INTL ROCHESTER INTL SKY HARBOR AP SOUTH ST PAUL MUNICIPAL ST CLOUD REGIONAL ST PAUL DOWNTOWN AP Mississippi HATTIESBURG-LAUREL AP JACKSON INTL KEESLER AFB KEY FIELD MERIDIAN NAS MCCAIN FIELD TUPELO REGIONAL Missouri CAPE GIRARDEAU REGIONAL COLUMBIA REGIONAL CR WHEELER DOWNTOWN JEFFERSON CITY MEMORIAL JOPLIN REGIONAL KANSAS CITY INTL LAMBERT-ST LOUIS INTL SPIRIT OF ST LOUIS AP SPRINGFIELD-BRANSON REGIONAL Montana BERT MOONEY AP BILLINGS LOGAN INTL BOZEMAN YELLOWSTONE INTL GREAT FALLS GREAT FALLS INTL MALMSTROM AFB MISSOULA INTL Nebraska CENTRAL NEBRASKA REGIONAL EPPLEY AIRFIELD LINCOLN MUNICIPAL NORTH OMAHA AP OFFUTT AFB Nevada MCCARRAN INTL NELLIS AFB RENO TAHOE INTL New Hampshire CONCORD MUNICIPAL JAFFREY-SILVER RANCH AP
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 32.2 23.4 30.3 22.3 28.8 21.5 32.2 23.1 29.9 21.8 28.0 20.9 31.9 23.0 30.2 22.1 28.7 21.5 32.9 23.2 31.3 22.4 29.7 21.5
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 24.7 29.7 23.7 28.3 24.4 29.2 23.3 27.7 24.5 29.6 23.6 28.3 24.9 30.7 23.8 29.0
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 23.0 18.1 27.4 22.3 17.3 26.5 22.8 17.9 26.5 22.2 17.3 26.0 22.8 18.0 27.6 22.2 17.4 26.7 23.0 18.2 27.8 22.3 17.4 26.9
Lat
Long
Elev
42.608N 42.911N 42.746N 42.233N
82.818W 82.529W 86.097W 83.533W
177 195 210 237
99.6% -16.8 -17.8 -14.0 -17.2
99% -14.0 -15.0 -12.2 -14.2
45.150N 45.062N 46.837N 44.832N 44.217N 44.883N 43.904N 46.722N 44.857N 45.543N 44.932N
93.217W 93.351W 92.183W 93.471W 93.917W 93.229W 92.492W 92.043W 93.033W 94.051W 93.056W
278 262 437 277 311 266 398 186 250 308 213
-22.7 -23.1 -27.3 -23.7 -24.6 -23.7 -24.7 -23.6 -22.6 -27.1 -23.3
-20.3 -21.1 -24.4 -21.3 -22.3 -21.0 -22.0 -21.3 -20.2 -24.1 -21.0
32.3 32.6 29.0 32.7 32.2 32.7 30.9 30.1 32.7 31.9 32.6
23.8 23.0 21.0 23.4 23.2 22.9 23.0 22.2 22.8 22.5 23.2
30.9 31.2 27.3 31.2 30.2 31.0 29.3 27.9 31.2 30.2 31.1
23.1 22.2 19.5 22.6 22.2 22.2 22.1 20.7 22.0 21.5 22.5
28.8 29.2 25.7 29.4 28.6 29.4 27.9 27.0 29.2 28.6 29.0
22.0 21.1 18.5 21.6 21.4 21.2 21.3 20.0 21.0 20.4 21.4
25.6 24.9 22.5 25.3 25.0 25.0 24.9 24.1 25.0 24.5 25.2
29.9 30.5 27.3 30.8 29.7 30.7 29.1 28.2 30.1 29.7 30.3
24.3 23.7 21.1 24.1 23.7 23.8 23.6 22.7 23.8 23.3 24.0
28.5 29.0 25.6 29.0 28.3 28.9 27.6 26.3 28.6 28.3 28.9
24.0 22.9 20.8 23.7 23.0 23.1 23.5 22.7 23.6 22.9 23.6
19.5 18.3 16.3 19.2 18.5 18.5 19.2 17.8 19.0 18.3 18.9
28.3 28.6 25.5 28.9 28.0 28.6 27.7 26.2 28.1 27.8 28.6
22.8 22.1 19.5 22.5 22.3 22.0 22.3 21.9 22.4 21.6 22.4
18.2 17.4 15.0 17.8 17.7 17.3 17.8 17.0 17.6 16.9 17.5
26.9 27.5 24.1 27.7 27.2 27.3 26.5 25.4 26.9 26.4 27.4
31.467N 32.321N 30.417N 32.335N 32.550N 34.262N
89.333W 89 90.078W 101 88.917W 10 88.744W 90 88.567W 83 88.771W 110
-3.9 -4.6 -0.8 -5.2 -5.3 -6.9
-2.4 -2.8 1.5 -3.3 -3.0 -4.6
35.9 35.7 34.6 35.6 36.0 35.7
24.3 24.5 26.8 24.3 24.6 24.4
34.0 34.6 33.5 34.4 34.8 34.3
23.9 24.5 26.3 24.4 24.6 24.3
32.9 33.5 32.6 33.3 33.7 33.2
23.8 24.4 26.0 24.3 24.4 24.1
25.8 26.5 28.6 26.5 26.7 26.3
32.1 32.3 32.4 31.9 32.7 32.2
25.4 25.9 27.9 25.9 26.0 25.7
31.6 31.5 31.3 31.2 32.0 31.5
24.0 25.0 27.6 25.1 25.1 24.7
19.1 20.3 23.6 20.4 20.4 20.0
28.4 28.7 30.7 28.7 29.8 28.9
23.7 24.5 27.2 24.5 24.3 24.1
18.7 19.7 22.9 19.7 19.5 19.3
28.1 28.2 30.2 28.1 28.8 28.4
37.225N 38.817N 39.121N 38.591N 37.147N 39.297N 38.753N 38.657N 37.240N
89.571W 92.218W 94.597W 92.156W 94.502W 94.731W 90.374W 90.656W 93.390W
102 272 226 175 299 306 162 141 384
-12.2 -15.8 -14.5 -14.2 -12.9 -16.4 -13.6 -14.3 -13.9
-9.0 -12.7 -12.1 -11.3 -10.0 -13.6 -10.7 -11.3 -10.7
34.7 34.8 36.3 35.6 36.3 35.6 35.6 35.3 35.3
25.0 24.3 24.5 24.4 23.8 24.7 24.8 25.1 23.3
33.5 33.2 34.5 33.8 34.7 33.7 34.1 33.8 33.5
24.8 24.4 24.3 24.0 24.0 24.5 24.5 24.6 23.4
32.4 31.6 33.0 32.5 33.1 32.1 32.8 32.4 32.0
24.4 23.9 24.0 23.7 23.8 24.1 23.9 24.0 23.3
26.8 26.3 26.4 26.3 25.9 26.6 26.4 26.5 25.4
32.3 31.9 33.1 32.2 32.4 32.5 32.7 32.6 31.5
26.0 25.5 25.7 25.5 25.3 25.7 25.7 25.7 24.8
31.4 31.0 32.1 31.4 31.9 31.6 31.8 31.7 30.8
25.2 24.7 24.5 24.8 24.0 24.9 24.6 24.9 23.6
20.6 20.4 20.0 20.2 19.6 20.8 20.0 20.3 19.3
30.2 29.8 30.4 29.4 29.7 30.4 30.0 29.8 28.6
24.4 23.9 23.8 23.8 23.4 24.0 23.8 24.0 23.0
19.6 19.4 19.1 19.1 18.8 19.6 19.1 19.2 18.6
29.2 28.9 29.8 28.6 28.9 29.6 29.5 29.1 27.9
45.965N 112.501W 45.807N 108.542W 45.788N 111.161W 47.450N 111.383W 47.473N 111.382W 47.517N 111.183W 46.921N 114.093W
1678 1092 1349 1131 1117 1058 973
-27.4 -23.0 -26.2 -25.5 -26.8 -26.1 -19.7
-22.8 -19.6 -22.1 -21.6 -23.0 -22.5 -16.0
31.1 34.9 33.4 32.5 33.6 34.7 33.9
14.1 17.0 16.2 15.7 16.1 17.1 16.5
29.3 33.0 31.5 30.7 31.7 32.6 32.1
13.6 16.6 15.8 15.3 15.7 16.4 16.2
27.6 31.1 29.6 29.0 29.8 30.8 30.2
13.3 16.3 15.2 14.9 15.2 15.8 15.7
15.7 19.0 17.9 17.5 17.8 18.8 18.2
26.3 29.6 28.4 28.1 29.1 30.6 29.6
14.9 18.1 17.0 16.6 16.9 17.6 17.3
25.4 28.8 27.6 27.0 28.0 29.2 28.8
12.5 15.7 14.5 14.1 14.3 14.7 14.6
11.1 12.7 12.2 11.5 11.7 11.9 11.7
17.0 22.2 20.6 19.4 19.5 22.1 20.2
11.2 14.5 13.3 12.9 13.1 13.5 13.5
10.2 11.8 11.2 10.6 10.8 11.0 10.8
16.6 21.5 20.0 19.8 19.4 21.0 19.7
40.961N 41.310N 40.851N 41.367N 41.117N
561 299 363 406 319
-19.5 -19.6 -19.2 -21.2 -19.2
-16.7 -17.0 -16.7 -17.8 -16.8
35.4 34.9 35.9 34.5 35.0
23.3 24.4 23.8 23.9 24.7
33.6 33.2 34.1 32.7 33.0
22.9 24.1 23.6 23.7 24.3
31.9 31.6 32.4 31.1 31.7
22.2 23.2 23.1 22.8 23.6
25.3 26.3 25.8 25.4 26.7
31.8 31.9 32.5 31.7 31.8
24.5 25.3 25.0 24.6 25.7
31.0 30.9 31.7 30.7 30.8
23.5 24.7 24.0 23.6 25.2
19.6 20.5 19.7 19.3 21.2
29.0 30.0 29.9 29.0 29.5
22.5 23.7 23.0 22.7 24.1
18.4 19.2 18.6 18.3 19.8
28.2 28.9 28.8 28.4 28.7
36.072N 115.163W 665 36.250N 115.033W 570 39.484N 119.771W 1344
-0.2 -2.0 -10.5
1.3 -0.2 -7.8
42.5 42.9 35.8
19.7 19.6 16.3
41.3 41.8 34.3
19.3 19.3 15.8
40.1 40.5 32.9
18.8 18.8 15.1
22.5 22.4 17.7
35.6 34.9 31.9
21.7 21.7 16.8
34.7 34.8 31.1
18.9 19.0 12.8
14.8 14.8 10.8
27.4 26.7 22.0
17.5 17.5 11.0
13.6 13.4 9.6
28.5 28.5 22.2
43.195N 71.501W 106 42.805N 72.004W 317
-19.4 -19.0
-16.6 -16.6
32.2 30.6
21.8 21.1
30.5 28.8
21.0 20.3
29.0 27.4
20.3 19.5
23.6 22.8
29.4 27.4
22.7 22.0
27.9 26.2
22.0 21.4
16.9 16.7
25.6 24.4
21.1 21.0
16.0 16.3
24.9 24.0
98.314W 95.899W 96.748W 96.017W 95.917W
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 9.7 8.6 7.9 3563 368 8.3 7.4 6.6 3778 254 11.4 9.9 8.5 3454 368 11.0 9.5 8.5 3523 407 11 sites, 87 more on CD-ROM 10.3 8.8 7.9 4183 343 9.4 8.5 7.7 4222 402 11.1 9.6 8.6 5159 121 9.8 8.6 7.8 4119 448 12.0 10.8 9.3 4281 349 10.7 9.2 8.4 4154 434 12.6 11.4 10.4 4351 293 12.7 11.2 9.6 4756 172 8.4 7.5 6.6 4095 428 10.6 9.0 8.1 4681 269 10.4 9.0 8.2 4147 418 6 sites, 13 more on CD-ROM 7.4 6.2 5.4 1159 1263 8.3 7.3 6.4 1262 1286 8.3 7.4 6.6 796 1556 8.3 7.4 6.5 1306 1201 7.4 5.8 5.1 1302 1240 8.4 7.5 6.8 1617 1117 9 sites, 16 more on CD-ROM 9.8 8.6 7.8 2329 841 10.7 9.2 8.3 2732 709 9.9 8.7 8.1 2537 924 9.3 8.2 7.3 2541 802 11.2 9.8 8.7 2247 927 11.4 10.2 8.9 2793 764 10.4 8.9 8.0 2446 931 9.2 8.2 7.4 2596 777 10.5 9.1 8.3 2460 781 7 sites, 18 more on CD-ROM 9.4 8.3 7.4 5053 44 12.3 11.0 9.7 3743 360 9.5 8.1 6.8 4540 132 14.0 12.2 11.0 4281 185 14.0 12.2 11.0 4198 187 13.8 12.1 10.9 3819 268 9.4 8.3 7.3 4085 187 5 sites, 37 more on CD-ROM 12.9 11.5 10.3 3384 586 11.9 10.7 9.2 3347 647 12.2 10.9 9.4 3316 662 10.4 8.6 7.9 3323 607 11.1 9.3 8.3 3306 648 3 sites, 14 more on CD-ROM 11.5 10.1 8.8 1090 1976 12.0 10.7 9.1 1151 1919 11.5 9.5 8.4 2756 484 4 sites, 11 more on CD-ROM 9.4 8.3 7.4 3923 265 7.3 6.1 5.4 4044 211
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station MANCHESTER-BOSTON REGIONAL PORTSMOUTH INTL New Jersey ATLANTIC CITY INTL MCGUIRE AFB MILLVILLE MUNICIPAL MONMOUTH EXECUTIVE NEWARK LIBERTY INTL TETERBORO AP TRENTON MERCER AP New Mexico ALAMOGORDO-WHITE SANDS AP ALBUQUERQUE INTL CANNON AFB CLOVIS MUNICIPAL FOUR CORNERS REGIONAL HOLLOMAN AFB ROSWELL INTL AIR CENTER WHITE SANDS New York ALBANY INTL AMBROSE LIGHT BUFFALO NIAGARA INTL CHAUTAUQUA COUNTY AP DUTCHESS COUNTY AP ELMIRA CORNING REGIONAL GREATER BINGHAMTON AP GREATER ROCHESTER INTL GRIFFIS AIRFIELD JOHN F KENNEDY INTL LAGUARDIA AP LONG ISLAND MACARTHUR AP NIAGARA FALLS INTL ONEIDA COUNTY AP PLATTSBURGH INTL REPUBLIC AP STEWART INTL SYRACUSE HANCOCK INTL WESTCHESTER COUNTY AP North Carolina ALBERT J ELLIS AP ASHEVILLE REGIONAL CHARLOTTE DOUGLAS INTL FAYETTEVILLE REGIONAL HICKORY REGIONAL NEW RIVER MCAS PIEDMONT TRIAD INTL PITT-GREENVILLE AP POPE AFB RALEIGH-DURHAM INTL SEYMOUR-JOHNSON AFB SIMMONS AAF SMITH REYNOLDS AP WILMINGTON INTL North Dakota BISMARCK MUNICIPAL GRAND FORKS AFB GRAND FORKS INTL
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB Dehumidification DP/HR/MCDB 0.4% 1% 0.4% 1% 2% 0.4% 1% 99% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB DP / HR / MCDB DP / HR / MCDB -13.8 32.9 22.1 31.4 21.4 29.8 20.7 24.1 29.7 23.1 28.3 22.4 17.2 26.5 21.6 16.4 25.7 -13.5 32.0 22.5 30.0 21.7 28.2 20.9 24.0 29.5 23.0 27.8 22.4 17.2 26.4 21.6 16.3 25.5
Heating DB
42.933N 71.438W 43.083N 70.817W
69 31
99.6% -16.6 -16.3
39.449N 40.017N 39.367N 40.183N 40.683N 40.850N 40.277N
18 40 21 49 2 3 56
-11.0 -11.3 -11.3 -11.4 -10.7 -11.2 -11.2
-8.6 -8.9 -8.8 -8.9 -8.4 -8.8 -8.9
33.5 33.8 33.4 32.8 34.6 33.7 33.7
24.0 24.1 23.9 23.3 23.6 23.4 23.5
31.8 32.4 31.9 31.3 32.9 32.2 32.2
23.2 23.6 23.3 22.6 22.8 22.7 22.9
30.3 30.9 30.5 29.1 31.3 30.7 30.7
22.6 22.9 22.7 21.9 22.2 22.0 22.4
25.5 25.8 25.5 24.9 25.3 25.1 25.1
30.9 30.8 30.6 30.5 31.3 30.7 31.1
24.8 25.1 24.8 24.0 24.5 24.2 24.3
29.4 30.1 29.4 29.0 29.8 29.3 29.5
24.0 24.7 24.1 22.9 23.7 23.3 23.3
18.9 19.8 19.0 17.8 18.5 18.1 18.2
27.7 28.1 27.5 27.1 27.8 27.1 27.3
23.4 23.7 23.5 22.5 23.0 22.7 22.7
18.2 18.6 18.4 17.3 17.7 17.4 17.6
27.0 27.4 27.0 26.8 27.1 26.6 26.7
32.840N 105.991W 35.042N 106.616W 34.383N 103.317W 34.433N 103.083W 36.744N 108.229W 32.850N 106.100W 33.308N 104.508W 32.383N 106.483W
1280 1619 1309 1285 1675 1267 1112 1244
-6.3 -7.6 -10.2 -11.0 -13.4 -7.4 -7.9 -7.6
-4.0 -5.6 -7.7 -8.0 -11.0 -5.5 -5.8 -5.3
37.7 35.3 37.0 36.3 35.5 37.7 38.4 37.2
17.5 15.5 17.3 17.8 15.3 17.1 18.0 17.6
37.0 34.0 35.3 34.8 34.0 36.4 36.9 35.8
17.6 15.4 17.4 17.7 15.0 16.9 18.2 17.7
35.1 32.8 33.9 33.0 32.8 35.1 35.7 34.6
17.3 15.3 17.6 17.7 14.9 16.9 18.1 17.7
21.5 18.4 21.4 20.8 18.1 20.3 21.5 21.0
30.3 27.5 27.1 29.3 27.5 29.2 30.4 30.8
20.7 18.0 20.6 20.2 17.5 19.9 20.9 20.5
29.4 27.0 27.2 29.0 27.3 28.9 29.9 30.1
19.0 16.4 20.2 18.6 16.1 18.5 19.5 18.8
16.1 14.2 17.5 15.8 14.1 15.6 16.3 15.9
23.9 20.0 22.0 23.2 19.5 21.7 23.4 22.3
18.0 15.6 19.1 17.7 15.0 17.6 18.8 18.1
15.2 13.5 16.3 14.9 13.1 14.7 15.6 15.2
23.6 20.5 21.6 22.5 19.7 22.1 23.1 22.4
42.743N 40.460N 42.941N 42.150N 41.627N 42.159N 42.207N 43.117N 43.234N 40.639N 40.779N 40.794N 43.108N 43.145N 44.650N 40.734N 41.500N 43.111N 41.067N
73.809W 73.830W 78.736W 79.250W 73.884W 76.892W 75.980W 77.677W 75.412W 73.762W 73.880W 73.102W 78.938W 75.384W 73.467W 73.417W 74.100W 76.104W 73.708W
95 0 218 525 51 291 486 164 158 3 3 26 178 217 71 25 150 126 116
-17.9 -10.1 -15.8 -17.5 -16.3 -17.8 -17.8 -16.2 -21.0 -9.9 -9.8 -11.2 -16.2 -20.4 -22.6 -10.9 -15.8 -18.2 -12.6
-15.2 -7.8 -13.7 -15.1 -13.1 -15.1 -15.3 -13.8 -17.4 -7.8 -7.6 -9.0 -13.9 -17.1 -19.3 -8.3 -12.7 -15.0 -10.3
31.6 28.9 30.2 28.0 33.0 32.0 29.6 31.4 31.1 32.2 33.6 31.4 31.0 30.6 30.4 32.2 32.3 31.8 32.0
22.5 N/A 21.8 20.9 23.0 22.0 21.0 22.8 22.2 22.8 23.3 23.0 22.5 22.5 21.7 23.2 22.3 22.8 23.0
30.0 27.2 28.8 27.3 31.4 30.2 28.0 29.8 29.5 30.3 32.0 29.8 29.5 29.0 28.3 30.2 30.5 30.2 30.2
21.7 N/A 21.1 20.3 22.4 21.1 20.2 21.7 21.2 22.1 22.5 22.2 21.7 21.4 20.8 22.2 21.9 21.8 22.1
28.6 25.8 27.5 26.0 29.8 28.7 26.6 28.3 28.1 28.8 30.5 28.4 28.0 27.7 26.8 28.7 28.9 28.7 28.8
21.0 N/A 20.5 19.4 21.6 20.4 19.5 20.9 20.4 21.6 21.9 21.7 20.9 20.6 20.1 21.8 21.0 21.0 21.4
24.1 N/A 23.7 22.4 24.7 23.5 22.6 24.0 23.7 24.8 24.9 24.8 24.0 23.9 23.3 24.8 24.3 24.0 24.5
29.3 N/A 27.6 26.7 30.4 29.2 27.1 29.2 28.9 28.9 30.5 28.7 28.7 28.4 27.9 29.1 29.4 29.8 29.3
23.2 N/A 22.8 21.4 23.9 22.6 21.7 23.0 22.7 24.1 24.2 24.0 23.1 22.9 22.3 24.1 23.4 23.0 23.7
27.8 N/A 26.8 25.4 29.0 28.0 25.9 27.7 27.5 27.7 29.0 27.4 27.5 27.0 26.6 27.7 28.2 28.1 27.9
22.4 N/A 22.4 21.2 22.9 21.9 21.1 22.4 22.0 23.6 23.3 23.6 22.5 22.5 21.8 23.7 22.8 22.2 23.0
17.3 N/A 17.6 16.9 17.7 17.2 16.8 17.4 17.0 18.4 18.1 18.5 17.6 17.6 16.6 18.5 17.8 17.2 18.0
26.6 N/A 26.1 25.2 27.3 26.3 24.7 26.8 26.5 26.8 27.2 26.5 26.5 26.1 25.8 26.6 26.8 27.1 26.4
21.6 N/A 21.4 20.0 22.3 21.0 20.3 21.4 21.1 22.9 22.6 22.9 21.8 21.4 20.9 22.9 22.2 21.3 22.4
16.5 N/A 16.5 15.7 17.1 16.2 15.9 16.4 16.0 17.7 17.4 17.7 16.8 16.5 15.7 17.7 17.2 16.2 17.4
25.8 N/A 25.3 24.0 26.7 25.4 23.9 25.6 25.4 25.9 26.7 25.8 25.7 25.2 24.9 25.8 26.0 26.0 25.8
34.833N 35.432N 35.224N 34.991N 35.743N 34.708N 36.097N 35.633N 35.174N 35.892N 35.344N 35.133N 36.134N 34.268N
77.617W 82.538W 80.955W 78.880W 81.382W 77.440W 79.943W 77.383W 79.009W 78.782W 77.965W 78.933W 80.222W 77.900W
29 645 222 57 348 8 271 8 67 127 33 74 296 12
-6.4 -9.2 -5.8 -5.1 -6.7 -4.8 -7.1 -6.2 -6.1 -6.4 -5.2 -5.3 -6.6 -4.0
-4.0 -6.9 -3.7 -2.9 -4.5 -2.9 -5.1 -3.9 -3.9 -4.3 -3.0 -3.2 -4.2 -2.3
34.8 31.1 34.5 35.7 33.5 33.9 33.6 35.0 36.2 35.0 35.9 35.9 33.4 34.1
25.0 21.8 23.7 24.4 22.6 25.6 23.4 24.5 24.6 24.2 24.6 24.3 23.2 25.5
33.0 29.8 33.2 34.1 32.2 32.6 32.4 33.8 34.8 33.6 34.3 34.5 32.4 32.9
24.5 21.4 23.4 23.8 22.4 25.1 23.0 24.0 24.3 23.9 24.3 24.0 22.8 25.0
32.4 28.7 32.0 32.9 31.0 31.5 31.2 32.7 33.2 32.3 33.0 33.0 31.2 31.7
24.2 20.9 23.0 23.6 22.0 24.7 22.6 23.6 23.8 23.5 23.9 23.6 22.4 24.6
26.6 23.2 25.1 26.0 24.4 26.9 24.9 26.1 26.6 25.7 26.7 26.0 24.7 26.7
32.6 28.4 31.4 32.0 29.9 31.5 30.9 32.5 31.6 32.1 31.3 32.0 30.5 31.6
25.8 22.7 24.6 25.4 23.8 26.2 24.2 25.3 25.8 25.1 26.1 25.5 24.1 26.1
31.4 27.5 30.5 31.2 29.1 30.8 29.7 31.1 31.0 31.2 30.6 31.2 29.6 30.6
25.0 21.8 23.4 24.3 22.9 25.8 23.1 24.1 25.3 24.0 25.8 24.6 22.9 25.4
20.1 17.8 18.7 19.4 18.4 21.1 18.4 19.0 20.7 19.2 21.2 19.8 18.3 20.7
29.4 25.1 27.4 27.7 26.4 29.3 27.2 28.1 28.3 28.2 27.9 28.0 26.7 28.8
24.0 21.2 22.9 23.9 22.3 24.9 22.6 23.8 24.7 23.4 25.1 24.0 22.5 24.9
19.0 17.2 18.1 18.9 17.8 20.0 17.9 18.6 19.9 18.5 20.2 19.1 17.8 20.0
28.4 24.5 26.8 27.4 25.9 28.5 26.6 27.8 27.9 27.5 27.5 27.5 26.3 28.3
46.783N 100.757W 503 47.967N 97.400W 278 47.943N 97.184W 257
-27.5 -28.9 -29.7
-24.5 -26.4 -27.0
33.9 32.0 31.7
21.1 22.3 22.0
31.9 30.1 29.9
20.5 21.2 20.9
30.0 28.5 28.4
19.8 20.3 19.9
23.6 24.7 23.8
30.0 28.7 29.2
22.3 23.2 22.6
29.0 27.8 28.0
21.6 23.6 22.1
17.3 19.0 17.3
27.6 26.9 27.3
20.2 22.1 20.7
15.8 17.4 15.8
26.0 25.7 25.9
74.567W 74.600W 75.067W 74.133W 74.169W 74.061W 74.816W
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 8.8 7.9 7.0 3448 407 10.4 8.9 7.9 3553 309 7 sites, 7 more on CD-ROM 11.2 9.6 8.4 2670 575 10.5 8.9 8.0 2694 596 9.1 8.1 7.3 2692 587 11.1 9.6 8.5 2807 511 11.2 9.8 8.7 2574 706 9.2 8.3 7.5 2731 606 9.1 8.1 7.3 2731 587 8 sites, 20 more on CD-ROM 10.1 8.4 7.4 1609 1061 12.7 11.1 9.2 2198 799 13.1 11.6 10.3 2066 801 14.3 12.3 10.9 2252 693 11.3 9.8 8.4 2928 549 11.1 9.2 8.0 1787 1021 11.9 10.2 8.5 1719 1098 8.4 7.2 5.9 1637 1006 19 sites, 26 more on CD-ROM 10.7 9.1 8.2 3596 355 19.3 16.9 15.2 2694 400 12.1 10.8 9.2 3589 322 9.6 8.4 7.7 3992 169 8.3 7.5 6.4 3335 403 9.1 8.1 7.1 3737 261 9.3 8.3 7.6 3907 224 11.3 9.6 8.5 3592 317 10.3 8.7 7.7 3884 269 12.3 11.1 9.8 2654 563 12.1 10.9 9.4 2496 713 10.9 9.2 8.4 2888 466 11.8 10.6 9.1 3664 323 9.3 8.4 7.7 3894 262 9.2 8.2 7.3 4186 208 11.0 9.3 8.4 2799 506 11.0 9.3 8.5 3300 396 10.9 9.2 8.2 3608 347 10.2 8.5 7.6 3039 429 14 sites, 55 more on CD-ROM 8.8 7.8 6.9 1639 965 10.2 8.6 7.7 2264 476 8.2 7.3 6.2 1689 939 9.1 8.0 7.1 1503 1102 7.8 6.6 5.7 1919 766 9.1 8.1 7.2 1409 1069 9.0 8.0 7.2 1963 808 8.4 7.3 6.2 1663 1037 8.5 7.5 6.5 1565 1130 8.4 7.5 6.6 1782 936 8.6 7.6 6.7 1492 1118 8.1 7.0 5.8 1523 1130 7.9 6.8 5.7 1898 816 9.3 8.3 7.5 1313 1122 6 sites, 10 more on CD-ROM 12.1 10.8 9.3 4696 294 12.6 11.2 9.9 5173 233 12.0 10.8 9.3 5202 232
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station HECTOR INTL MINOT AFB MINOT INTL Ohio AKRON-CANTON REGIONAL CINCINNATI MUNICIPAL LUNKEN CLEVELAND HOPKINS INTL FAIRFIELD COUNTY AP FINDLAY AP JMCOX DAYTON INTL MANSFIELD LAHM REGIONAL OHIO STATE UNIVERSITY AP PORT COLUMBUS INTL RICKENBACKER INTL TOLEDO EXPRESS AP WRIGHT-PATTERSON AFB YOUNGSTOWN-WARREN REGIONAL Oklahoma HENRY POST AAF LAWTON-FORT SILL REGIONAL RICHARD L JONES JR AP STILLWATER REGIONAL TINKER AFB TULSA INTL VANCE AFB WILEY POST AP WILL ROGERS WORLD AP Oregon AURORA STATE AP CORVALLIS MUNICIPAL EUGENE AP MCMINNVILLE MUNICIPAL PORTLAND INTL PORTLAND-HILLSBORO AP ROBERTS FIELD ROGUE VALLEY INTL SALEM MUNICIPAL Pennsylvania ALLEGHENY COUNTY AP ALTOONA-BLAIR COUNTY AP BUTLER COUNTY AP CAPITAL CITY AP ERIE INTL HARRISBURG INTL LEHIGH VALLEY INTL NORTHEAST PHILADELPHIA AP PHILADELPHIA INTL PITTSBURGH INTL READING REGIONAL WASHINGTON COUNTY AP WILKES-BARRE SCRANTON INTL WILLOW GROVE NAS Rhode Island T F GREEN AP
Lat
Long
Elev
46.925N 96.811W 274 48.417N 101.350W 508 48.255N 101.273W 508
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C Cooling DB/MCWB 0.4% 1% 2% 99% DB / MCWB DB / MCWB DB / MCWB -25.5 32.2 22.4 30.6 21.3 29.1 20.4 -27.4 32.5 20.5 30.6 20.0 28.7 19.1 -25.1 32.5 20.7 30.5 20.1 28.7 19.0
Heating DB 99.6% -28.2 -30.2 -27.8
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB Dehumidification DP/HR/MCDB 0.4% 1% 0.4% 1% WB / MCDB WB / MCDB DP / HR / MCDB DP / HR / MCDB 24.1 29.6 23.0 28.6 22.4 17.7 27.9 21.1 16.2 26.7 22.9 29.0 21.6 27.6 21.1 16.8 26.0 19.8 15.4 24.8 23.1 28.9 21.7 27.7 21.2 16.9 26.6 19.8 15.4 25.3
40.917N 39.103N 41.405N 39.756N 41.014N 39.906N 40.820N 40.078N 39.991N 39.817N 41.589N 39.833N 41.254N
81.433W 84.419W 81.853W 82.657W 83.669W 84.219W 82.518W 83.078W 82.881W 82.933W 83.801W 84.050W 80.674W
368 149 235 265 244 305 395 276 247 227 204 251 360
-16.0 -13.2 -15.2 -16.5 -17.1 -16.5 -16.9 -15.3 -14.7 -14.0 -16.9 -16.0 -16.0
-13.2 -10.1 -12.2 -12.5 -13.9 -13.1 -13.9 -12.3 -11.7 -11.1 -13.9 -12.5 -13.3
31.4 33.7 31.9 32.5 32.4 32.4 31.1 32.5 32.9 33.7 32.9 32.8 31.2
22.6 23.9 23.1 23.3 23.0 23.2 22.7 23.0 23.2 23.5 23.4 23.5 22.5
30.0 32.3 30.4 31.3 31.0 31.1 29.8 31.2 31.7 32.5 31.3 31.5 29.8
21.9 23.5 22.3 22.9 22.3 22.6 21.9 22.6 22.6 23.0 22.4 22.9 21.5
28.6 31.0 29.0 29.9 29.5 29.7 28.5 29.9 30.4 31.2 29.9 30.1 28.5
21.0 23.0 21.6 22.0 21.5 21.8 21.2 21.8 21.9 22.4 21.6 22.1 20.8
24.1 25.6 24.5 25.0 24.7 24.7 24.2 24.7 24.8 26.5 25.0 25.0 23.8
29.1 31.1 29.6 30.5 30.2 30.2 29.1 30.2 30.6 30.1 30.5 30.4 29.0
23.2 24.9 23.7 24.1 23.7 24.0 23.3 23.9 24.0 25.3 24.0 24.2 22.9
27.8 30.1 28.4 29.0 28.6 28.9 27.8 28.9 29.3 29.9 29.0 29.4 27.7
22.5 24.0 22.9 23.1 22.9 23.0 22.7 22.9 23.1 25.9 23.3 23.4 22.2
18.0 19.3 18.1 18.4 18.2 18.4 18.3 18.2 18.4 21.8 18.5 18.8 17.7
26.9 28.2 27.3 27.5 28.0 27.8 26.9 27.6 27.4 28.9 28.0 27.5 26.4
21.8 23.3 22.1 22.5 22.2 22.3 21.9 22.4 22.4 23.9 22.4 22.6 21.4
17.2 18.4 17.2 17.8 17.4 17.6 17.4 17.6 17.6 19.3 17.5 17.9 16.7
25.6 27.4 26.4 26.9 26.9 26.9 26.0 26.8 26.8 27.2 26.7 26.7 25.3
34.650N 34.558N 36.039N 36.162N 35.417N 36.199N 36.333N 35.534N 35.389N
98.400W 98.417W 95.984W 97.089W 97.383W 95.887W 97.917W 97.647W 97.601W
362 326 195 300 394 198 398 395 392
-9.1 -8.8 -9.8 -11.0 -10.1 -10.3 -12.1 -10.1 -9.7
-6.6 -6.7 -7.5 -7.9 -7.5 -7.6 -9.1 -7.4 -7.2
39.7 40.3 38.7 39.2 38.0 38.0 38.8 38.0 38.1
22.5 22.6 24.1 23.5 23.0 24.0 22.8 23.0 23.2
37.9 38.8 37.2 37.5 36.4 36.4 37.4 36.8 36.6
22.6 22.7 24.6 23.9 23.1 24.2 22.9 23.1 23.3
36.4 37.4 35.6 36.0 34.8 34.9 35.9 35.1 35.0
22.5 23.0 24.5 23.8 23.3 24.2 23.0 23.1 23.3
25.2 25.5 26.4 26.1 25.6 26.2 25.3 25.2 25.5
32.8 33.5 34.2 34.1 32.3 33.5 33.2 32.8 32.8
24.6 24.9 25.8 25.5 25.0 25.6 24.7 24.6 24.9
32.4 32.9 33.5 33.4 31.8 32.9 32.5 32.2 32.3
23.2 23.1 24.1 23.9 23.9 24.1 23.3 22.9 23.5
18.8 18.5 19.5 19.5 19.6 19.5 19.0 18.5 19.2
27.8 28.3 29.6 30.1 28.5 29.8 27.8 28.7 28.8
22.6 22.7 23.7 23.0 23.0 23.6 22.7 22.5 22.9
18.0 18.1 19.0 18.4 18.6 18.8 18.3 18.0 18.5
27.4 28.0 29.3 29.0 27.8 29.3 27.5 28.1 28.1
45.249N 122.769W 44.500N 123.283W 44.128N 123.221W 45.195N 123.134W 45.596N 122.609W 45.541N 122.949W 44.256N 121.139W 42.381N 122.872W 44.905N 123.001W
60 76 108 49 6 62 928 395 63
-3.8 -4.0 -5.1 -3.5 -3.9 -5.2 -14.9 -5.1 -4.6
-2.2 -2.5 -2.9 -2.1 -1.6 -3.0 -11.1 -3.3 -2.6
33.0 33.7 33.2 33.4 32.9 33.2 34.0 37.2 33.4
19.5 19.4 19.3 19.3 19.7 19.9 16.3 19.5 19.4
31.2 32.1 31.1 31.3 30.6 31.1 32.3 35.3 31.2
19.2 18.9 18.7 18.9 19.1 19.4 15.9 18.7 18.8
28.8 29.7 29.0 29.0 28.6 28.8 30.6 33.5 29.1
18.4 18.0 18.1 18.2 18.4 18.5 15.2 18.1 18.1
21.0 20.5 20.5 20.5 20.9 21.2 17.6 20.5 20.4
30.0 31.4 30.8 30.7 30.4 30.9 31.2 34.5 31.3
20.0 19.5 19.6 19.6 20.0 20.1 16.7 19.6 19.6
29.0 30.4 29.3 29.6 29.0 29.4 29.9 33.0 29.5
17.6 16.2 16.7 16.9 17.3 17.6 12.7 15.7 16.5
12.7 11.6 12.1 12.1 12.4 12.7 10.2 11.7 11.8
23.9 24.2 23.7 23.0 24.0 24.7 19.4 23.4 23.0
17.0 15.0 15.8 16.0 16.5 16.4 11.6 14.6 15.6
12.2 10.8 11.3 11.4 11.8 11.8 9.5 10.9 11.2
23.0 23.7 22.5 22.4 23.0 22.9 19.2 23.2 22.4
40.355N 40.296N 40.777N 40.217N 42.080N 40.196N 40.651N 40.082N 39.868N 40.485N 40.367N 40.133N 41.334N 40.200N
79.922W 78.320W 79.950W 76.851W 80.183W 76.772W 75.449W 75.011W 75.231W 80.214W 75.967W 80.283W 75.727W 75.150W
380 451 380 104 223 95 119 31 3 367 105 361 284 110
-14.0 -14.2 -16.1 -11.0 -14.4 -11.3 -12.9 -10.4 -9.7 -14.6 -12.3 -16.3 -15.2 -10.8
-11.7 -12.0 -12.9 -8.8 -12.0 -9.1 -10.5 -7.9 -7.4 -12.0 -9.9 -12.9 -12.5 -8.6
31.8 31.3 31.2 33.5 30.4 33.6 32.8 34.0 34.2 31.8 33.7 31.3 31.7 33.3
22.3 22.1 22.2 23.3 22.8 24.0 23.2 23.8 23.8 22.2 23.5 21.8 22.1 23.3
30.3 29.7 29.2 31.9 29.0 32.0 31.3 32.5 32.7 30.3 32.1 29.7 30.1 31.9
21.6 21.5 21.4 22.5 22.1 23.2 22.5 23.1 23.3 21.5 22.9 21.1 21.2 22.6
29.0 28.3 27.9 30.5 27.7 30.4 29.8 31.2 31.3 29.1 30.5 28.2 28.6 30.5
20.9 20.8 20.6 22.0 21.4 22.6 21.8 22.5 22.6 20.9 22.2 20.4 20.5 21.9
23.9 23.7 23.7 24.7 24.1 25.5 24.8 25.5 25.6 23.9 25.1 23.4 23.7 25.1
29.1 28.9 28.7 30.6 28.2 31.0 30.3 31.4 31.4 29.2 31.1 28.5 29.0 30.9
23.0 22.9 22.8 24.0 23.2 24.6 24.0 24.6 24.8 23.1 24.2 22.5 22.8 24.2
28.0 27.6 27.6 29.4 27.3 29.7 28.8 30.1 30.0 27.8 29.4 27.7 27.7 29.4
22.4 22.2 22.3 22.9 22.7 23.9 23.1 23.8 24.0 22.3 23.1 22.1 22.1 23.3
17.9 17.8 17.8 17.9 17.9 19.0 18.1 18.7 18.9 17.8 18.0 17.5 17.4 18.3
26.5 26.5 26.7 27.0 27.0 28.3 27.2 28.0 28.1 26.5 27.6 26.8 25.9 28.1
21.5 21.3 21.3 22.4 21.8 23.0 22.4 22.9 23.3 21.5 22.6 21.1 21.3 22.5
16.9 16.9 16.7 17.3 16.9 17.9 17.4 17.7 18.1 16.9 17.5 16.5 16.5 17.4
25.6 25.4 25.4 26.5 25.9 27.2 26.5 26.9 27.3 25.6 27.0 25.5 25.2 27.1
41.722N 71.433W
18
-12.9
-10.4
32.3
22.9
30.4
22.1
28.8
21.3
24.6
29.6
23.8
27.8
23.1
17.9
26.7
22.4
17.2
25.9
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 12.6 11.3 10.3 4856 306 13.2 11.7 10.5 5193 197 12.5 11.1 9.8 4886 239 13 sites, 21 more on CD-ROM 10.4 8.9 8.1 3329 391 9.1 8.2 7.3 2642 624 11.0 9.4 8.5 3223 437 9.0 7.9 7.1 3027 453 11.2 9.7 8.6 3284 448 11.1 9.6 8.5 3045 529 10.8 9.2 8.4 3394 372 9.9 8.6 7.7 3028 516 10.3 8.7 7.7 2881 586 10.5 8.8 7.8 2792 632 10.9 9.3 8.3 3361 449 10.0 8.7 7.8 2979 516 9.4 8.4 7.6 3422 323 9 sites, 33 more on CD-ROM 11.6 10.2 9.0 1752 1261 11.7 10.4 9.0 1754 1319 8.8 8.0 7.2 1952 1158 11.1 9.9 8.8 2006 1151 11.9 10.7 9.4 1882 1109 11.0 9.7 8.7 1917 1171 12.5 11.2 10.0 2178 1089 12.2 11.2 10.3 1929 1173 12.4 11.2 10.2 1900 1120 9 sites, 29 more on CD-ROM 8.1 7.1 5.8 2478 208 8.8 7.9 7.2 2392 218 8.7 7.7 7.0 2591 157 9.2 7.9 6.9 2586 167 10.6 8.8 7.8 2351 248 8.3 7.4 6.2 2662 153 9.3 8.4 7.5 3634 132 8.2 6.8 5.5 2368 480 9.3 8.1 7.2 2502 189 14 sites, 18 more on CD-ROM 9.1 8.2 7.4 2994 466 10.4 8.7 7.7 3267 344 8.1 7.1 5.9 3389 312 9.1 8.1 7.2 2783 592 10.9 9.4 8.6 3353 371 11.5 10.1 8.5 2794 622 10.0 8.5 7.6 3046 473 9.9 8.5 7.8 2607 663 11.1 9.4 8.4 2458 754 10.2 8.7 7.8 3073 436 10.2 8.6 7.8 2872 556 8.5 7.5 6.5 3310 300 9.0 8.0 7.2 3337 363 8.7 7.5 6.6 2742 574 1 site, 10 more on CD-ROM 10.8 9.2 8.3 3045 425
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station South Carolina CHARLESTON INTL COLUMBIA METROPOLITAN FLORENCE REGIONAL FOLLY ISLAND GREENVILLE-SPARTANBURG INTL SHAW AFB South Dakota ELLSWORTH AFB RAPID CITY REGIONAL SIOUX FALLS REGIONAL Tennessee CHATTANOOGA AP MCGHEE TYSON AP MCKELLAR-SIPES REGIONAL MEMPHIS INTL MILLINGTON MUNICIPAL NASHVILLE INTL TRI-CITIES REGIONAL Texas ABILENE REGIONAL AL MANGHAM JR REGIONAL AMARILLO RICK HUSBAND INTL ANGELINA COUNTY AP AUSTIN-BERGSTROM INTL BROWNSVILLE INTL CORPUS CHRISTI INTL CORPUS CHRISTI NAS DALLAS EXECUTIVE DALLAS FORT WORTH INTL DALLAS HENSLEY FIELD NAS DALLAS LOVE FIELD DEL RIO INTL DRAUGHON-MILLER CENTRAL TEXAS DYESS AFB ABILENE EASTERWOOD FIELD EL PASO INTL FORT WORTH ALLIANCE AP FORT WORTH NAS FT WORTH MEACHAM INTL GALVESTON SCHOLES INTL GEORGE BUSH INTERCONTINENTAL GEORGETOWN MUNICIPAL HOOKS MEMORIAL HOUSTON ELLINGTON AP JACK BROOKS REGIONAL KILLEEN MUNICIPAL LACKLAND AFB LAREDO INTL LAUGHLIN AFB LONGVIEW EAST TEXAS REGIONAL LUBBOCK INTL MCALLEN INTL MCGREGOR EXECUTIVE MCKINNEY NATL MIDLAND INTL NEW BRAUNFELS MUNICIPAL PORT ARANSAS
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
99.6%
99%
80.040W 12 81.118W 69 79.724W 45 79.890W 3 82.221W 287 80.467W 74
-2.5 -4.8 -4.4 -0.2 -5.6 -4.6
-0.6 -2.8 -2.6 1.7 -3.6 -2.7
34.6 36.2 35.6 30.8 34.6 35.6
25.6 24.1 24.8 25.5 23.1 24.3
33.4 34.8 34.2 30.1 33.2 34.1
25.3 23.8 24.4 25.5 23.0 24.1
32.4 33.6 33.0 29.5 31.9 32.9
25.0 23.6 24.1 25.4 22.6 23.9
27.1 25.8 26.3 27.0 25.0 26.2
31.7 32.1 32.5 29.4 31.0 32.2
26.6 25.4 25.7 26.5 24.4 25.6
31.0 31.4 31.6 29.0 30.0 31.2
25.9 24.3 24.7 26.3 23.3 24.7
21.3 19.5 19.8 21.7 18.8 19.9
29.1 28.0 28.6 29.0 26.8 28.5
25.3 23.8 24.1 25.7 22.8 24.1
20.5 18.8 19.1 20.9 18.1 19.1
28.6 27.4 27.9 28.5 26.4 27.8
44.150N 103.100W 999 44.043N 103.054W 963 43.578N 96.754W 435
-22.3 -22.7 -24.1
-19.3 -19.7 -21.4
35.1 35.9 33.0
19.1 18.9 23.2
33.0 33.7 31.4
18.7 18.7 22.7
31.2 31.7 29.8
18.4 18.3 21.8
21.8 21.7 25.2
29.9 29.7 30.4
20.8 20.8 24.2
28.8 29.1 29.4
19.2 19.3 23.7
15.8 15.8 19.6
25.3 25.6 28.7
18.1 18.2 22.6
14.7 14.7 18.2
24.3 24.3 27.6
35.031N 35.818N 35.593N 35.056N 35.350N 36.119N 36.473N
32.899N 33.942N 34.185N 32.680N 34.884N 33.967N
85.201W 83.986W 88.917W 89.987W 89.867W 86.689W 82.404W
205 293 132 77 98 183 457
-6.9 -8.3 -9.1 -6.9 -7.5 -9.2 -10.3
-4.6 -5.8 -6.7 -4.6 -4.9 -6.7 -7.8
35.0 33.7 35.0 35.9 37.4 34.8 32.4
23.6 23.3 24.7 25.0 26.8 23.8 22.1
33.7 32.4 33.8 34.7 36.1 33.5 31.1
23.5 23.1 24.7 24.7 26.0 23.7 21.8
32.5 31.3 32.7 33.6 34.1 32.4 30.0
23.1 22.7 24.4 24.4 25.2 23.3 21.6
25.4 25.0 26.6 26.6 28.3 25.6 23.9
31.6 30.9 32.4 33.1 35.1 31.6 29.4
24.9 24.4 25.9 26.0 27.4 25.0 23.3
30.7 29.9 31.6 32.2 33.6 30.9 28.7
23.8 23.3 25.0 24.8 26.4 23.9 22.3
19.1 18.8 20.4 20.1 22.1 19.2 17.9
27.5 27.4 29.9 29.9 31.8 28.3 26.1
23.2 22.8 24.3 24.2 26.0 23.3 21.7
18.5 18.2 19.5 19.3 21.6 18.5 17.3
27.0 26.9 29.0 29.4 31.4 27.7 25.5
32.411N 99.682W 31.578N 94.709W 35.230N 101.704W 31.236N 94.754W 30.183N 97.680W 25.914N 97.423W 27.774N 97.512W 27.683N 97.283W 32.681N 96.868W 32.898N 97.019W 32.733N 96.967W 32.852N 96.856W 29.378N 100.927W 31.150N 97.417W 32.417N 99.850W 30.589N 96.365W 31.811N 106.376W 32.973N 97.318W 32.767N 97.450W 32.819N 97.361W 29.273N 94.859W 29.980N 95.360W 30.679N 97.679W 30.068N 95.556W 29.617N 95.167W 29.951N 94.021W 31.083N 97.683W 29.383N 98.583W 27.544N 99.463W 29.367N 100.783W 32.385N 94.712W 33.666N 101.823W 26.176N 98.240W 31.485N 97.316W 33.190N 96.591W 31.948N 102.209W 29.709N 98.046W 27.830N 97.050W
546 108 1099 88 146 7 13 6 201 171 150 134 305 208 545 93 1194 209 185 209 2 29 240 46 10 5 256 210 151 330 111 992 31 180 179 872 197 0
-6.6 -3.9 -12.2 -2.8 -2.9 3.7 1.5 2.9 -3.9 -4.8 -5.8 -4.1 -0.3 -3.9 -7.2 -2.0 -3.9 -6.0 -4.4 -5.3 2.5 -0.5 -3.0 -1.3 1.2 -0.2 -3.2 -1.3 2.3 -0.7 -3.6 -8.9 3.7 -4.0 -6.2 -6.4 -2.1 3.1
-4.1 -2.5 -9.0 -1.3 -1.0 5.9 3.5 5.3 -2.2 -2.5 -2.7 -1.8 1.4 -2.2 -4.8 -0.1 -2.0 -3.7 -2.2 -2.9 4.2 1.1 -2.0 0.5 2.8 1.7 -1.7 0.6 4.0 1.3 -2.1 -6.6 5.6 -2.4 -3.9 -4.2 -0.3 5.3
37.9 37.3 37.0 37.2 37.9 35.4 36.0 33.8 38.7 38.5 37.6 38.6 38.9 37.8 39.1 37.8 38.4 39.1 39.2 38.6 33.3 36.4 37.2 37.0 36.0 35.0 37.8 38.0 40.4 40.2 37.9 37.5 38.0 38.7 38.4 38.4 37.9 30.1
21.4 24.3 18.6 24.5 23.4 25.5 25.3 26.3 23.4 23.5 24.2 23.7 22.2 23.3 22.2 24.1 17.9 23.3 22.8 23.3 26.2 24.7 22.7 24.3 25.6 25.4 23.3 23.4 23.6 22.2 23.9 19.0 24.4 23.6 23.5 19.3 23.2 25.5
36.7 36.0 35.4 35.8 36.8 34.7 35.1 33.1 37.4 37.3 36.4 37.4 37.8 37.2 37.9 36.6 37.1 37.8 37.9 37.4 32.7 35.3 36.1 35.7 34.9 34.0 37.2 37.2 39.1 38.9 36.6 36.1 37.2 37.5 37.3 37.0 37.0 29.7
21.4 24.3 18.8 24.7 23.5 25.5 25.4 26.3 23.7 23.6 24.1 24.0 22.3 23.3 22.1 24.2 17.6 23.6 23.1 23.6 26.2 24.7 22.7 24.4 25.8 25.5 23.4 23.4 23.5 22.5 24.0 19.4 24.7 23.7 23.9 19.5 23.3 25.7
35.5 34.0 34.0 34.6 35.7 34.1 34.3 32.7 36.2 36.0 35.1 36.2 36.9 36.0 36.7 35.5 35.9 36.5 36.7 36.2 32.2 34.3 34.7 34.1 33.9 33.2 36.0 36.1 38.0 37.7 35.1 34.8 36.3 36.3 36.0 35.8 35.9 29.5
21.5 24.2 18.8 24.6 23.6 25.5 25.3 26.3 23.6 23.7 23.9 23.8 22.3 23.3 22.0 24.1 17.5 23.4 23.2 23.5 26.2 24.7 22.7 24.6 25.8 25.5 23.4 23.4 23.5 22.5 24.1 19.4 24.6 23.6 23.8 19.6 23.3 25.6
24.2 26.1 21.8 26.6 25.8 27.2 27.2 28.0 25.6 25.8 26.1 26.1 25.3 25.5 24.9 26.5 21.1 25.7 25.7 25.8 27.7 26.7 25.0 26.8 27.9 27.5 25.3 26.7 26.4 26.0 26.0 22.6 26.8 25.8 25.8 22.8 25.6 26.9
31.7 32.3 30.2 32.4 31.6 31.1 31.9 31.1 33.2 33.0 33.4 33.6 31.9 32.5 33.1 32.5 30.2 33.6 33.5 33.1 31.0 31.6 31.4 31.2 31.6 31.6 33.2 31.2 33.1 32.2 32.5 31.0 32.6 33.3 33.2 31.1 31.6 28.8
23.6 25.7 21.1 26.2 25.5 26.8 26.8 27.5 25.2 25.4 25.5 25.5 24.9 25.1 24.2 25.9 20.6 25.2 25.2 25.3 27.3 26.3 24.5 26.3 27.4 26.9 25.0 26.0 25.9 25.3 25.6 21.9 26.4 25.4 25.4 22.2 25.2 26.8
31.2 31.9 29.8 31.9 31.3 30.9 31.4 30.8 32.7 32.6 32.9 32.9 31.5 32.2 32.5 31.6 29.7 33.0 32.9 32.7 30.6 31.1 30.9 30.9 31.1 31.1 32.8 30.6 32.6 32.2 32.0 30.4 31.7 33.1 32.7 30.5 31.1 28.7
22.3 24.7 19.5 25.2 24.6 26.3 26.2 27.3 23.8 24.1 24.1 24.1 23.8 23.8 22.9 25.1 19.1 23.7 23.7 24.0 27.1 25.6 23.0 26.0 27.3 26.4 23.0 25.8 25.0 24.7 24.4 20.3 25.8 24.0 24.0 20.8 24.2 26.4
18.1 20.0 16.3 20.5 20.0 21.8 21.6 23.2 19.1 19.4 19.4 19.3 19.4 19.1 18.8 20.5 16.0 19.0 18.9 19.3 22.9 20.9 18.3 21.5 23.1 21.8 18.3 21.7 20.5 20.5 19.6 17.0 21.2 19.2 19.3 17.2 19.6 21.8
26.6 27.9 24.0 28.2 27.3 28.3 28.5 29.3 27.9 28.7 29.8 29.1 27.4 27.6 27.2 28.4 22.9 28.6 28.5 28.6 29.1 28.2 26.6 28.0 29.0 29.3 28.0 27.8 28.2 28.5 27.9 24.9 27.9 28.4 28.2 24.5 26.8 28.4
21.7 24.0 18.9 24.8 24.2 25.9 25.8 26.7 23.0 23.5 23.4 23.6 23.2 23.0 22.3 24.6 18.4 23.0 23.0 23.3 26.3 25.2 22.8 25.2 26.4 25.9 22.7 25.0 24.3 23.7 24.0 19.7 25.3 23.0 23.4 20.2 23.9 26.1
17.5 19.2 15.7 20.1 19.4 21.2 21.1 22.3 18.2 18.7 18.6 18.7 18.6 18.2 18.2 19.9 15.3 18.2 18.1 18.6 21.8 20.4 18.0 20.4 21.8 21.3 18.0 20.6 19.6 19.2 19.1 16.3 20.5 18.2 18.6 16.6 19.2 21.5
26.3 27.6 23.5 28.1 26.9 28.2 28.3 29.1 27.5 28.2 29.0 28.7 27.2 27.2 26.9 27.9 23.0 27.9 28.1 28.0 29.2 28.0 26.5 27.8 28.5 28.8 27.8 27.6 28.1 27.7 27.7 24.3 27.8 27.9 27.8 24.4 26.7 28.3
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 6 sites, 13 more on CD-ROM 9.2 8.2 7.4 1025 1322 8.7 7.6 6.6 1351 1223 8.7 7.9 7.0 1331 1174 14.7 11.7 10.3 1043 1198 8.6 7.6 6.9 1676 912 8.8 7.8 6.9 1353 1148 3 sites, 21 more on CD-ROM 15.8 13.4 11.6 3904 377 15.8 13.7 11.7 3927 366 12.2 11.0 9.4 4172 411 7 sites, 6 more on CD-ROM 8.0 7.1 6.1 1718 994 9.2 7.9 6.8 1962 847 8.8 8.0 7.1 1909 966 9.2 8.2 7.4 1602 1259 8.1 7.2 6.1 1703 1181 8.5 7.6 6.8 1939 964 8.4 7.4 6.1 2321 576 51 sites, 101 more on CD-ROM 11.7 10.7 9.3 1358 1382 8.3 7.3 6.2 1191 1367 13.5 11.9 10.9 2246 817 8.0 7.2 6.3 1017 1505 9.3 8.4 7.6 908 1669 11.8 10.7 9.3 283 2263 12.3 11.2 10.3 459 2014 11.7 10.7 9.3 392 2084 10.1 8.7 7.8 1173 1586 11.7 10.6 9.2 1201 1599 9.2 8.4 7.6 1206 1513 10.3 9.1 8.3 1128 1680 9.3 8.3 7.5 705 1963 11.2 10.1 9.0 1099 1537 12.0 10.9 9.4 1394 1469 9.2 8.3 7.5 871 1722 12.2 10.5 8.7 1278 1401 10.5 9.2 8.4 1332 1509 11.0 9.7 8.7 1162 1654 10.4 9.1 8.3 1245 1543 11.4 10.1 9.0 550 1858 8.9 8.1 7.3 744 1735 9.4 8.4 7.6 1082 1493 7.9 7.1 6.1 810 1676 8.9 8.1 7.3 648 1776 9.6 8.6 7.9 738 1638 9.9 8.9 8.1 1049 1589 9.3 8.4 7.5 755 1821 10.9 9.5 8.8 435 2474 10.7 9.2 8.3 655 2061 9.1 8.1 7.3 1178 1460 12.9 11.6 10.5 1809 1064 11.1 10.2 9.2 296 2501 10.5 9.2 8.4 1159 1552 10.6 9.1 8.2 1395 1414 12.0 10.8 9.3 1419 1329 10.6 9.2 8.3 839 1701 17.2 14.4 12.0 442 1716
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station RANDOLPH AFB REESE AFB LUBBOCK ROBERT GRAY AFF SABINE PASS SAN ANGELO REGIONAL SAN ANTONIO INTL SAN MARCOS REGIONAL STINSON MUNICIPAL VALLEY INTL VICTORIA REGIONAL WACO REGIONAL WICHITA FALL REGIONAL WILLIAM P HOBBY AP Utah HILL AFB LOGAN-CACHE AP PROVO MUNICIPAL SALT LAKE CITY INTL ST GEORGE MUNICIPAL Vermont BURLINGTON INTL Virginia DANVILLE REGIONAL DAVISON AAF DINWIDDIE COUNTY LANGLEY AFB LESSBURG EXECUTIVE LYNCHBURG REGIONAL MANSAS REGIONAL NEWPORT NEWS WILLIAMSBURG NORFOLK INTL NORFOLK NAS OCEANA NAS QUANTICO MCAF RICHMOND INTL ROANOKE-BLACKSBURG REGIONAL RONALD REAGAN WASHINGTON NATL SHENANDOAH VALLEY REGIONAL VIRGINIA TECH MONTGOMERY EXEC WASHINGTON DULLES INTL Washington ARLINGTON MUNICIPAL BELLINGHAM INTL BREMERTON NATL FAIRCHILD AFB FELTS FIELD GRAY AFF KING COUNTY INTL AP MCCHORD AFB OLYMPIA REGIONAL PAINE FIELD PEARSON FIELD SANDERSON FIELD SEATTLE-TACOMA INTL SOUTHWEST WASHINGTON REGIONAL SPOKANE INTL TACOMA NARROWS AP TRI-CITIES AP
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 37.9 23.2 37.0 23.2 35.9 23.2 38.3 19.4 36.6 19.6 35.1 19.6 38.0 22.8 37.2 22.9 36.0 22.9 31.7 25.2 30.8 25.3 30.3 25.3 38.7 21.1 37.5 21.0 36.3 21.0 37.4 23.0 36.4 23.1 35.4 23.2 37.7 23.4 37.1 23.4 35.9 23.4 38.4 23.1 37.4 23.3 36.3 23.1 37.1 25.2 36.2 25.2 35.4 25.2 36.6 24.7 35.5 24.8 34.5 24.8 38.5 23.5 37.4 23.7 36.2 23.8 39.9 22.6 38.4 22.6 37.0 22.7 35.5 25.0 34.5 25.1 33.6 25.0
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 25.8 31.2 25.4 30.8 22.9 30.7 22.2 30.4 25.4 32.2 24.9 31.6 27.0 29.7 26.7 29.5 24.0 31.7 23.4 31.2 25.6 31.1 25.2 30.6 25.7 32.4 25.3 32.2 26.0 31.6 25.5 31.2 27.2 31.9 26.8 31.6 26.9 31.1 26.5 30.8 25.8 32.8 25.5 32.4 25.3 33.2 24.8 32.7 26.9 31.7 26.5 31.2
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 24.7 20.2 27.0 24.1 19.5 26.7 20.8 17.6 25.9 20.0 16.6 25.4 23.9 19.5 27.2 23.1 18.6 26.7 26.4 21.8 28.8 25.8 21.1 28.7 22.0 18.0 26.4 21.4 17.3 26.0 24.4 20.0 26.8 24.0 19.5 26.6 24.0 19.3 28.3 23.3 18.5 27.9 24.6 20.0 27.6 24.1 19.4 27.1 26.3 21.8 28.2 26.0 21.4 28.2 25.9 21.4 27.9 25.6 20.9 27.8 24.2 19.5 27.8 23.8 19.0 27.6 23.2 18.7 28.2 22.6 18.0 27.7 25.8 21.2 28.5 25.3 20.4 28.2
29.533N 98.262W 33.600N 102.050W 31.067N 97.833W 29.670N 94.050W 31.352N 100.495W 29.544N 98.484W 29.891N 97.864W 29.339N 98.472W 26.228N 97.654W 28.861N 96.930W 31.619N 97.228W 33.979N 98.493W 29.638N 95.282W
222 1017 309 1 584 241 182 174 10 35 152 310 13
99.6% -2.0 -9.6 -3.5 0.2 -5.5 -1.2 -2.5 -0.8 2.8 -0.4 -4.1 -7.4 0.9
99% 0.0 -7.0 -1.3 2.3 -3.4 0.6 -1.0 1.0 5.0 1.4 -2.1 -5.2 2.6
41.117N 111.967W 41.787N 111.853W 40.219N 111.723W 40.778N 111.969W 37.100N 113.600W
1460 1358 1371 1288 895
-13.0 -21.0 -13.6 -12.3 -3.8
-11.0 -17.6 -11.2 -9.8 -2.4
34.4 34.9 34.8 36.6 41.2
16.0 16.5 16.9 16.9 19.0
33.0 33.5 32.9 35.2 39.7
15.7 16.1 16.7 16.6 18.4
31.8 32.2 32.1 33.8 38.0
15.5 15.8 16.5 16.3 17.9
18.1 18.4 19.1 19.0 20.6
29.3 29.6 30.6 30.7 34.2
17.4 17.7 18.3 18.4 20.1
29.0 29.0 29.6 30.2 33.7
14.5 15.3 15.2 15.7 17.4
12.3 12.8 12.8 13.1 13.8
22.0 21.1 23.9 22.3 24.3
13.0 13.8 13.9 14.4 16.1
11.1 11.6 11.8 12.0 12.8
22.5 20.8 23.7 22.6 25.3
44.468N 73.150W 101
-21.9
-18.8
31.3
21.8
29.6
21.0
28.1
20.2
23.5
28.7
22.5
27.4
21.8
16.7
25.9
20.9
15.7
25.1
36.573N 38.717N 37.183N 37.083N 39.078N 37.321N 38.721N 37.132N 36.903N 36.937N 36.817N 38.504N 37.505N 37.317N 38.848N 38.264N 37.208N 38.941N
79.336W 77.183W 77.500W 76.360W 77.558W 79.207W 77.515W 76.493W 76.192W 76.289W 76.033W 77.305W 77.320W 79.974W 77.034W 78.896W 80.408W 77.464W
174 22 59 3 119 287 59 13 9 5 7 3 50 358 3 366 650 88
-7.6 -9.4 -8.7 -6.2 -9.9 -9.0 -11.3 -6.4 -5.1 -4.4 -5.6 -8.3 -7.5 -8.4 -7.8 -11.2 -11.7 -10.4
-5.7 -7.1 -7.0 -3.9 -7.7 -6.8 -8.8 -4.3 -3.0 -2.4 -3.4 -6.4 -5.5 -6.4 -5.8 -8.7 -8.9 -8.1
34.4 35.5 36.7 33.8 35.0 33.4 33.9 34.7 34.3 34.7 33.8 33.7 35.1 33.4 34.8 34.0 32.1 34.2
23.6 24.3 25.0 24.6 24.6 23.2 23.5 25.0 24.9 25.1 25.1 24.6 24.3 22.6 24.2 23.2 22.6 23.7
33.0 33.9 34.9 32.5 33.5 32.0 32.6 33.3 32.9 33.3 32.5 32.4 33.7 32.0 33.3 32.7 30.3 32.7
23.4 23.7 24.5 24.2 23.9 22.8 23.3 24.5 24.4 24.6 24.6 24.2 23.9 22.2 23.7 23.0 21.8 23.2
32.0 32.5 33.0 31.3 32.4 30.6 31.3 32.2 31.6 32.2 31.2 31.0 32.3 30.7 31.9 31.5 28.8 31.3
23.0 23.2 23.8 23.7 23.5 22.2 22.7 24.0 23.9 24.2 24.1 23.7 23.4 21.7 23.1 22.6 21.4 22.6
25.4 26.2 26.9 26.3 26.2 24.7 25.1 26.3 26.2 26.6 26.3 26.4 25.8 24.1 25.8 25.7 24.4 25.3
31.7 32.5 33.1 30.5 32.5 30.5 31.3 32.4 31.4 32.0 31.6 31.6 31.9 30.2 31.7 30.6 28.7 31.6
24.7 25.3 26.1 25.7 25.4 24.0 24.4 25.7 25.6 25.8 25.7 25.5 25.2 23.5 25.2 24.9 23.6 24.6
30.6 31.2 32.5 29.9 31.3 29.5 30.4 31.2 30.6 31.0 30.4 30.6 31.0 29.4 30.8 29.9 27.7 30.2
23.6 24.4 25.2 25.2 24.2 23.0 23.0 24.8 24.8 25.1 25.0 24.9 24.3 22.3 24.3 24.1 22.9 23.5
18.8 19.3 20.5 20.4 19.3 18.4 17.9 19.9 19.9 20.2 20.0 19.9 19.3 17.8 19.2 19.8 19.1 18.5
28.0 29.0 30.0 27.9 28.3 26.9 27.8 28.9 28.3 29.0 28.8 29.6 28.2 26.3 28.4 28.0 25.9 27.6
22.9 23.6 24.1 24.5 23.8 22.4 22.5 24.1 24.2 24.3 24.2 24.0 23.6 21.7 23.6 23.0 22.4 22.9
18.0 18.5 19.1 19.5 18.9 17.7 17.4 19.0 19.1 19.3 19.1 18.8 18.5 17.1 18.4 18.6 18.5 17.8
27.1 28.0 28.7 27.5 28.0 26.3 27.4 28.2 27.6 28.4 28.0 28.5 27.5 25.8 27.8 26.8 25.5 27.0
48.161N 122.159W 48.794N 122.537W 47.483N 122.767W 47.633N 117.650W 47.683N 117.321W 47.083N 122.583W 47.530N 122.301W 47.150N 122.483W 46.973N 122.903W 47.908N 122.280W 45.621N 122.657W 47.238N 123.141W 47.444N 122.314W 46.117N 122.894W 47.622N 117.528W 47.268N 122.576W 46.267N 119.117W
42 45 135 750 595 91 6 98 57 185 9 83 113 6 717 89 124
-7.0 -7.1 -5.2 -14.9 -13.1 -6.7 -3.4 -6.4 -6.5 -3.9 -4.4 -5.4 -3.6 -5.1 -15.0 -2.6 -12.1
-4.2 -4.6 -3.0 -11.6 -9.9 -4.2 -1.4 -4.2 -4.3 -1.7 -2.7 -3.4 -1.4 -2.9 -11.5 -0.6 -8.7
27.8 26.3 29.8 33.2 34.7 30.9 29.8 30.2 30.7 26.7 32.6 30.8 29.6 31.1 33.8 28.7 37.4
18.9 18.5 18.4 16.7 18.3 18.6 18.5 18.4 18.9 17.5 19.1 18.4 18.4 19.6 17.1 18.0 20.8
26.3 24.6 27.6 31.7 32.8 28.4 27.7 28.0 28.5 24.6 30.6 28.3 27.6 28.0 32.0 26.8 35.7
18.0 17.7 17.5 16.3 17.6 17.9 17.8 17.6 18.2 16.8 18.8 17.9 17.7 18.6 16.5 17.1 19.9
24.1 22.9 26.0 29.8 31.1 26.5 26.1 26.2 26.5 22.8 28.3 26.1 25.6 26.3 30.1 24.8 33.7
17.2 16.8 16.8 15.8 17.0 17.1 17.1 16.9 17.4 16.1 18.1 17.1 17.1 17.7 15.9 16.5 19.3
19.7 19.3 19.2 18.3 19.7 19.6 19.4 19.3 19.9 18.5 20.6 19.6 19.4 20.4 18.4 18.9 22.1
26.8 25.3 28.1 29.2 31.8 28.6 28.2 28.1 29.2 25.0 30.0 28.6 28.0 28.8 30.5 27.0 34.3
18.6 18.3 18.2 17.4 18.6 18.6 18.5 18.4 18.9 17.5 19.7 18.5 18.4 19.3 17.6 17.9 20.9
25.2 23.6 26.4 28.3 30.3 26.7 26.2 26.5 27.2 23.2 28.8 26.9 26.1 27.2 29.3 25.1 33.1
17.0 16.8 16.0 15.0 15.7 16.4 16.2 16.3 16.5 16.1 17.2 16.2 16.2 17.2 14.4 16.1 17.8
12.2 12.1 11.5 11.7 11.9 11.8 11.5 11.7 11.8 11.7 12.3 11.6 11.7 12.3 11.2 11.5 13.0
23.3 22.1 22.1 20.0 22.1 20.4 20.8 21.0 21.8 20.3 23.5 21.0 21.2 23.9 20.3 20.3 26.3
16.1 16.0 14.4 13.6 14.2 15.9 15.3 15.4 15.7 15.0 16.2 15.3 15.4 16.2 13.3 15.1 17.0
11.5 11.4 10.4 10.7 10.8 11.4 10.9 11.0 11.2 10.9 11.5 11.0 11.1 11.5 10.4 10.8 12.3
21.8 20.7 20.4 19.0 21.9 20.3 20.8 20.2 21.0 19.8 22.7 20.3 20.3 22.4 20.0 19.5 25.2
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 9.6 8.6 7.9 808 1724 12.2 10.8 9.2 1768 1017 10.7 9.3 8.4 1001 1614 15.5 12.1 10.6 793 1478 11.0 9.5 8.6 1214 1468 9.2 8.3 7.6 767 1790 11.0 9.5 8.6 901 1684 8.5 7.7 7.1 708 1888 12.4 11.1 10.2 321 2280 10.9 9.4 8.5 647 1804 10.8 9.3 8.6 1116 1613 12.1 11.0 9.7 1552 1414 9.2 8.3 7.6 633 1790 5 sites, 10 more on CD-ROM 10.6 9.1 8.3 3321 529 8.9 7.5 5.9 3998 282 10.8 9.0 7.8 3318 432 11.1 9.3 8.2 3037 701 11.9 10.3 8.7 1652 1512 1 site, 6 more on CD-ROM 10.4 8.9 8.1 4015 292 18 sites, 35 more on CD-ROM 8.3 7.3 6.3 2027 800 9.6 8.1 6.7 2349 754 8.0 7.0 5.7 2046 894 10.7 9.1 8.2 1897 880 10.0 8.4 7.3 2462 748 7.8 6.9 5.8 2329 618 9.6 8.3 7.2 2681 597 9.1 8.2 7.5 1893 910 10.9 9.2 8.4 1758 953 11.3 9.7 8.5 1651 1049 11.0 9.4 8.4 1812 886 9.0 7.9 7.0 2294 764 9.3 8.3 7.5 2019 869 10.3 8.6 7.6 2201 688 10.4 8.9 8.1 2167 882 7.9 6.8 5.6 2482 627 9.0 7.9 6.9 2666 437 9.6 8.4 7.5 2551 660 20 sites, 28 more on CD-ROM 9.3 8.1 7.0 3001 34 11.3 9.3 8.2 2966 31 8.7 7.6 6.7 3134 54 10.9 9.2 8.1 3773 223 8.9 7.8 6.7 3401 258 8.3 7.2 6.0 2869 85 8.2 7.3 6.2 2419 149 9.1 7.9 7.0 2867 74 8.4 7.4 6.4 2987 61 10.9 9.0 7.8 2884 45 7.3 5.9 5.3 2453 212 9.1 8.2 7.3 3051 56 9.1 8.1 7.3 2614 109 7.7 6.7 5.7 2684 101 11.5 9.8 8.5 3682 256 8.7 7.8 6.9 2662 81 11.1 9.3 8.2 2747 453
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station WALLA WALLA REGIONAL WEST POINT YAKIMA AIR TERMINAL West Virginia HUNTINGTON TRI-STATE AP MID-OHIO VALLEY REGIONAL YEAGER AP Wisconsin APPLETON INTL AUSTIN STRAUBEL INTL CENTRAL WISCONSIN AP CHIPPEWA VALLEY REGIONAL DANE COUNTY REGIONAL FOND DU LAC COUNTY AP GENERAL MITCHELL INTL KENOSHA REGIONAL LA CROSSE MUNICIPAL MANITOWOC COUNTY AP SHEBOYGAN SHEBOYGAN COUNTY MEMORIAL WAUSAU DOWNTOWN AP WITTMAN REGIONAL Wyoming CASPER-NATRONA COUNTY INTL CHEYENNE REGIONAL
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C Cooling DB/MCWB 0.4% 1% 2% 99% DB / MCWB DB / MCWB DB / MCWB -8.2 37.0 19.1 34.8 18.3 32.8 17.7 0.6 21.4 16.1 20.1 15.7 19.0 15.3 -10.1 35.8 19.0 34.1 18.5 32.2 17.7
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB Dehumidification DP/HR/MCDB 0.4% 1% 0.4% 1% WB / MCDB WB / MCDB DP / HR / MCDB DP / HR / MCDB 20.2 33.8 19.2 32.4 15.8 11.7 23.2 14.4 10.7 22.6 17.0 19.6 16.2 18.8 15.8 11.2 17.9 15.1 10.7 17.2 20.2 32.8 19.2 31.7 15.7 11.6 24.3 14.4 10.7 23.5
46.095N 118.287W 355 47.660N 122.440W 3 46.568N 120.543W 324
99.6% -11.7 -1.4 -13.2
38.365N 82.555W 251 39.345N 81.439W 253 38.379N 81.590W 277
-11.8 -13.0 -11.9
-9.0 -10.2 -9.0
33.2 32.5 32.9
23.2 23.2 22.8
31.8 31.2 31.6
22.9 22.7 22.6
30.5 29.9 30.3
22.4 22.0 22.1
25.2 24.9 24.8
30.3 30.1 29.9
24.4 24.0 24.0
29.3 28.9 28.9
23.7 23.2 23.3
19.1 18.6 18.7
27.5 27.3 27.0
22.9 22.5 22.6
18.2 17.8 17.9
26.7 26.5 26.1
44.267N 44.479N 44.783N 44.867N 43.141N 43.769N 42.955N 42.595N 43.879N 44.133N 43.750N 43.769N 44.929N 43.984N
280 209 389 270 264 246 204 227 199 198 176 227 366 238
-21.1 -22.2 -23.9 -25.1 -21.3 -21.3 -18.3 -19.1 -22.9 -20.2 -18.8 -20.3 -24.4 -21.2
-17.9 -19.4 -21.8 -22.2 -18.5 -18.3 -15.8 -16.5 -20.1 -17.5 -16.2 -17.7 -21.7 -18.7
31.3 31.1 30.3 32.2 31.9 31.6 32.0 32.4 33.0 29.3 28.4 31.4 30.9 31.4
24.0 23.1 22.4 22.8 23.4 23.1 23.6 23.5 23.9 22.1 21.9 23.3 22.1 23.1
29.4 29.4 28.7 30.4 30.3 29.9 30.2 30.8 31.3 27.7 26.4 29.2 29.1 29.7
22.7 22.1 21.4 21.6 22.4 22.0 22.4 22.9 22.8 21.3 21.4 21.9 20.8 22.0
27.9 28.0 27.3 28.9 28.8 28.2 28.6 28.9 29.9 26.4 24.8 27.7 27.7 28.0
21.6 21.2 20.1 20.7 21.7 21.1 21.4 21.8 21.9 20.2 21.0 21.1 19.8 21.2
25.4 24.6 23.5 24.4 25.0 24.6 24.8 25.0 25.4 23.8 24.6 24.4 23.6 24.6
29.6 29.3 28.5 29.7 30.1 29.5 30.2 30.5 30.8 28.2 26.3 29.4 28.4 29.4
24.2 23.5 22.2 23.3 23.8 23.5 23.8 24.0 24.3 22.6 23.4 23.3 22.5 23.5
28.1 27.8 27.0 28.3 28.5 28.1 28.5 28.8 29.5 26.5 24.9 27.8 27.1 28.0
24.0 23.0 22.0 22.7 23.3 22.8 23.0 23.0 23.8 22.4 24.2 22.8 22.1 22.9
19.5 18.2 17.5 18.0 18.6 18.1 18.2 18.2 19.1 17.5 19.5 18.0 17.5 18.1
27.5 27.5 27.3 27.5 28.3 27.8 27.8 27.8 28.7 26.8 25.2 27.5 26.1 27.5
22.8 22.0 20.9 21.6 22.3 22.2 22.1 22.4 22.6 21.3 23.0 22.0 21.0 22.2
18.1 17.1 16.3 16.8 17.5 17.4 17.2 17.6 17.7 16.3 18.1 17.1 16.4 17.3
26.4 26.5 26.0 26.4 27.0 26.9 26.9 27.1 27.5 25.4 24.3 26.2 25.2 26.7
42.898N 106.474W 1619 41.150N 104.817W 1868
-22.5 -20.1
-18.4 -16.4
34.5 31.9
15.3 14.5
32.9 30.5
15.0 14.2
31.3 28.9
14.7 14.0
17.3 17.0
28.2 25.2
16.6 16.4
27.7 24.9
14.2 14.7
12.3 13.2
19.2 18.8
13.0 13.8
11.4 12.4
18.9 18.4
49.730N 111.450W 51.120N 114.010W 51.080N 114.220W 53.570N 113.520W 53.310N 113.580W 53.670N 113.470W 56.650N 111.210W 55.180N 118.880W 52.450N 113.760W 49.700N 112.770W 50.030N 110.720W 52.180N 113.890W 51.100N 114.370W
817 1099 1235 671 723 688 369 669 860 910 715 905 1201
-29.3 -28.5 -27.3 -28.9 -32.6 -30.1 -34.7 -35.9 -32.0 -27.8 -28.6 -32.0 -31.4
-25.8 -24.9 -24.2 -25.8 -29.2 -27.0 -31.9 -31.4 -28.3 -24.5 -25.2 -28.2 -27.7
31.5 28.5 28.0 28.4 27.8 27.8 28.7 27.7 28.1 31.7 33.1 27.9 26.9
18.0 16.1 15.2 18.0 17.9 17.9 17.7 16.8 18.2 16.9 17.0 17.3 15.6
29.6 26.6 26.0 26.6 25.9 26.0 26.7 25.7 26.1 29.7 31.1 26.0 24.9
17.3 15.5 14.5 17.1 17.2 16.9 16.6 15.9 17.3 16.3 16.4 16.4 14.6
27.7 24.7 24.1 24.9 24.3 24.2 25.0 24.0 24.4 27.7 29.2 24.3 23.2
16.8 14.9 14.0 16.1 16.1 15.9 15.7 15.0 16.3 16.0 16.1 15.5 14.2
20.2 17.8 17.4 19.2 19.5 19.1 19.2 18.1 19.6 19.1 18.9 18.6 16.9
28.1 25.3 24.0 26.5 25.8 25.9 25.8 25.4 25.9 27.3 28.3 25.8 24.3
19.0 16.7 16.2 18.1 18.3 18.0 18.1 17.0 18.3 18.0 17.9 17.5 15.8
26.6 24.4 23.0 24.8 24.3 24.4 24.1 23.8 24.5 26.2 27.3 24.2 22.9
17.5 15.0 15.1 16.6 17.1 16.5 17.0 15.4 17.2 16.3 15.8 15.9 14.0
13.8 12.2 12.5 12.8 13.4 12.8 12.7 11.9 13.7 12.9 12.2 12.6 11.5
23.8 21.1 20.1 22.6 23.0 22.3 21.0 20.6 23.2 22.6 22.0 22.2 20.2
16.2 13.7 13.7 15.4 15.9 15.4 15.7 14.2 15.9 14.9 14.6 14.8 12.9
12.7 11.2 11.4 11.8 12.3 11.9 11.7 11.0 12.5 11.9 11.3 11.7 10.7
22.5 19.5 18.6 21.3 21.6 20.9 20.0 19.2 21.7 21.3 21.0 20.8 18.5
49.030N 122.360W 49.240N 121.760W 49.350N 124.160W 49.720N 124.900W 48.420N 123.230W 49.210N 123.810W 48.430N 123.440W 49.490N 123.300W 50.700N 120.450W 49.960N 119.380W 48.570N 123.530W 49.460N 119.600W 49.210N 122.690W 49.330N 123.260W 53.880N 122.680W 49.110N 123.300W 49.560N 119.650W
59 19 13 26 19 8 3 7 345 433 366 344 5 14 691 11 454
-7.7 -7.4 -1.0 -4.7 -1.3 -1.4 -2.6 -2.8 -19.5 -17.3 -5.8 -13.4 -7.1 -1.8 -30.2 -3.4 -14.0
-5.1 -4.9 0.5 -2.7 1.2 0.2 -0.7 -0.9 -15.6 -13.7 -3.6 -10.7 -4.7 -0.1 -25.4 -1.2 -10.8
29.8 30.1 23.5 26.8 22.7 23.6 22.1 24.7 34.1 33.1 27.7 32.9 30.2 24.5 28.0 22.2 32.9
19.6 20.3 19.1 17.8 N/A 18.2 15.9 19.1 18.2 18.2 17.4 18.7 19.8 N/A 16.4 N/A 17.7
27.7 28.2 22.3 24.8 20.7 22.2 20.4 23.1 31.9 31.2 25.7 31.0 28.2 23.2 25.9 21.2 31.0
18.9 19.6 18.5 17.1 N/A 17.7 15.3 18.2 17.6 17.5 16.6 18.0 19.1 N/A 15.6 N/A 17.2
25.8 26.4 21.2 23.1 19.1 21.0 19.0 21.8 29.7 29.1 24.0 29.2 26.2 22.1 24.0 20.2 29.1
18.0 19.0 17.8 16.5 N/A 17.2 14.8 17.8 16.8 16.7 16.0 17.3 18.3 N/A 14.6 N/A 16.6
20.5 21.6 19.8 18.6 N/A 18.7 16.7 20.0 19.2 19.3 18.9 19.6 20.7 N/A 17.3 N/A 19.3
28.4 28.1 22.6 24.9 N/A 21.7 20.4 23.2 31.2 30.0 25.6 30.5 28.2 N/A 25.9 N/A 29.3
19.4 20.5 19.0 17.8 N/A 18.1 16.0 19.1 18.3 18.2 17.8 18.7 19.6 N/A 16.2 N/A 18.3
26.5 26.7 21.6 23.3 N/A 21.1 19.2 22.1 29.6 28.6 23.8 28.9 26.6 N/A 24.1 N/A 28.0
17.4 19.2 18.7 16.2 N/A 17.4 15.2 18.7 15.2 15.6 16.0 15.7 17.8 N/A 14.1 N/A 15.9
12.5 14.0 13.5 11.5 N/A 12.5 10.8 13.5 11.3 11.7 11.9 11.6 12.8 N/A 10.9 N/A 11.9
24.4 25.4 21.7 20.5 N/A 20.9 17.9 22.2 21.6 21.7 22.8 23.0 23.8 N/A 19.2 N/A 22.3
16.4 18.1 17.8 15.5 N/A 16.7 14.6 17.8 14.2 14.6 15.1 14.8 16.8 N/A 13.2 N/A 14.7
11.8 13.0 12.8 11.0 N/A 11.9 10.4 12.8 10.5 10.9 11.2 10.9 12.0 N/A 10.3 N/A 11.0
22.9 23.8 20.7 19.7 N/A 20.3 17.2 21.1 21.0 20.9 21.2 22.4 22.1 N/A 18.3 N/A 21.5
88.517W 88.137W 89.667W 91.488W 89.345W 88.491W 87.904W 87.938W 91.253W 87.667W 87.690W 87.851W 89.628W 88.557W
Canada Alberta BOW ISLAND CALGARY INTL CANADIAN OLYMPIC PARK UPPER EDMONTON CITY CENTRE AWOS EDMONTON INTL EDMONTON NAMAO AWOS FORT MCMURRAY CS GRANDE PRAIRIE LACOMBE CDA 2 LETHBRIDGE CDA MEDICINE HAT RCS RED DEER SPRINGBANK British Columbia ABBOTSFORD AGASSIZ RCS BALLENAS ISLAND COMOX DISCOVERY ISLAND ENTRANCE ISLAND ESQUIMALT HARBOUR HOWE SOUND PAM ROCKS KAMLOOPS KELOWNA MALAHAT PENTICTON PITT MEADOWS CS POINT ATKINSON PRINCE GEORGE SANDHEADS CS SUMMERLAND CS
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 10.9 9.0 8.0 2672 519 16.4 13.9 11.7 2737 5 10.1 8.4 7.2 3247 309 3 sites, 12 more on CD-ROM 7.5 6.6 5.6 2443 635 8.2 7.2 6.2 2714 524 7.7 6.6 5.4 2446 599 14 sites, 42 more on CD-ROM 11.1 9.7 8.5 4006 338 10.6 9.0 8.1 4217 267 10.3 8.8 7.9 4613 201 9.0 8.1 7.3 4357 326 9.6 8.5 7.7 3935 356 10.5 9.0 8.2 3985 334 11.1 9.7 8.7 3709 385 11.1 9.6 8.6 3742 349 10.4 8.8 8.1 3903 456 10.8 9.3 8.4 4207 197 18.4 15.1 12.7 4027 189 10.8 9.2 8.3 4149 248 9.1 8.1 7.2 4458 253 10.3 8.9 8.0 4079 317 2 sites, 21 more on CD-ROM 14.4 12.6 11.4 4060 261 14.9 12.8 11.5 3920 197 100 sites, 665 more on CD-ROM 13 sites, 114 more on CD-ROM 12.8 11.1 9.7 4780 111 12.0 10.3 9.0 5109 37 10.2 8.8 7.7 5039 41 10.0 8.6 7.5 5235 73 10.2 8.9 7.7 5818 25 10.2 8.9 7.7 5589 37 8.7 7.4 6.5 6247 42 11.3 9.8 8.4 5930 26 9.4 8.1 6.9 5732 23 13.3 11.9 10.5 4504 114 10.8 9.3 8.1 4721 179 8.9 8.1 7.2 5727 23 11.1 9.5 8.3 5720 4 27 sites, 71 more on CD-ROM 9.1 7.6 6.5 2911 81 10.1 7.9 6.2 2857 116 15.9 13.6 12.0 2651 57 13.4 11.5 9.7 3089 57 16.2 12.7 9.8 2775 10 14.2 12.5 11.2 2693 58 9.8 8.5 7.4 3042 6 17.8 15.6 13.2 2649 77 10.3 9.0 8.1 3533 278 7.9 6.5 5.5 3894 140 6.6 5.7 4.8 3254 96 10.4 9.1 8.1 3426 228 5.4 4.6 4.0 2990 78 13.5 11.5 9.8 2433 103 9.8 8.4 7.4 5136 22 13.6 12.0 10.6 2768 29 8.2 6.5 5.3 3503 254
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station VANCOUVER HARBOUR CS VANCOUVER INTL VERNON VICTORIA GONZALES CS VICTORIA HARTLAND CS VICTORIA INTL VICTORIA UNIVERSITY CS WEST VANCOUVER WHITE ROCK CAMPBELL SCI YOHO PARK Manitoba WINNIPEG INTL New Brunswick FREDERICTON INTL MONCTON INTL SAINT JOHN Newfoundland and Labrador ST JOHN'S INTL Northwest Territories YELLOWKNIFE Nova Scotia HALIFAX STANFIELD INTL SHEARWATER SYDNEY Nunavut IQALUIT CLIMATE Ontario BEAUSOLEIL BELLE RIVER ERIEAU GUELPH TURFGRASS INSTITUTE HAMILTON INTL LONDON INTL NORTH BAY OTTAWA INTL PETERBOROUGH TRENT U PORT WELLER REGION OF WATERLOO INTL SAULT STE MARIE SUDBURY THUNDER BAY CS TIMMINS TORONTO BILLY BISHOP TORONTO BUTTONVILLE TORONTO PEARSON INTL TRENTON WINDSOR Prince Edward Island CHARLOTTETOWN Québec BAGOTVILLE BIG TROUT LAKE JONQUIERE LA BAIE LAC SAINT-PIERRE L'ACADIE L'ASSOMPTION LENNOXVILLE
Lat
Long
49.300N 123.120W 49.190N 123.180W 50.220N 119.190W 48.410N 123.320W 48.530N 123.460W 48.650N 123.430W 48.460N 123.300W 49.350N 123.190W 49.020N 122.780W 51.440N 116.340W
Elev 3 4 482 61 154 20 60 170 13 1602
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C Cooling DB/MCWB 0.4% 1% 2% 99% DB / MCWB DB / MCWB DB / MCWB -0.9 25.8 17.9 24.4 17.2 23.1 16.6 -3.4 25.1 18.5 23.6 17.9 22.4 17.2 -13.2 32.9 18.4 30.8 17.8 28.7 17.1 -0.7 24.6 16.9 22.2 16.0 20.5 15.2 -1.6 28.5 18.7 26.6 17.9 24.8 17.3 -2.3 26.7 17.7 24.6 16.9 22.9 16.2 -0.6 26.9 18.1 25.0 17.4 23.3 16.8 -3.2 27.0 18.4 25.2 17.9 23.6 17.3 -2.9 24.8 18.8 23.2 18.1 22.0 17.4 -26.6 25.5 13.7 23.4 13.0 21.3 12.2
Heating DB 99.6% -2.9 -6.1 -16.5 -3.0 -3.7 -4.0 -2.7 -5.8 -5.3 -29.9
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 18.7 24.2 18.0 23.5 19.2 23.9 18.4 22.8 19.7 29.4 18.7 28.3 17.7 23.0 16.7 21.0 20.0 26.5 19.0 24.9 18.2 25.5 17.3 23.7 19.3 25.1 18.3 23.4 19.6 25.4 18.7 23.9 19.7 23.6 18.7 22.3 14.7 22.7 13.7 21.4
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 16.3 11.6 21.6 15.6 11.1 20.6 17.2 12.3 21.9 16.4 11.7 21.0 16.7 12.6 21.5 15.6 11.7 20.6 15.6 11.1 19.0 14.9 10.6 18.1 17.6 12.8 22.4 16.7 12.1 21.2 15.1 10.7 20.2 14.4 10.3 19.6 17.2 12.3 21.1 16.3 11.6 20.1 17.3 12.6 22.7 16.5 12.0 21.6 18.1 13.1 22.0 17.2 12.3 20.7 11.9 10.6 15.8 10.9 9.9 15.0
49.910N 97.240W 239
-32.0
-29.7
30.3
21.3
28.6
20.3
27.0
19.4
23.0
28.4
21.7
26.9
21.2
16.3
26.5
19.8
14.9
24.8
45.870N 66.540W 21 46.110N 64.680W 71 45.320N 65.890W 109
-23.2 -22.2 -22.3
-20.6 -19.7 -19.5
29.9 28.7 26.2
21.0 20.9 18.8
28.1 27.0 24.6
19.9 19.8 17.9
26.4 25.4 23.0
19.0 18.9 17.0
22.3 22.2 20.2
27.8 26.7 24.1
21.3 21.2 19.2
26.2 25.2 22.7
20.4 20.7 18.8
15.2 15.5 13.8
24.8 24.4 21.9
19.6 19.8 18.0
14.3 14.6 13.1
23.7 23.5 20.6
47.620N 52.750W 141
-15.1
-13.0
24.8
19.1
23.3
18.3
21.9
17.7
20.6
23.2
19.6
22.0
19.7
14.7
22.2
18.6
13.7
21.0
62.460N 114.440W 206
-40.5
-38.3
25.4
16.0
23.7
15.2
22.2
14.6
17.2
22.7
16.3
21.8
15.1
11.0
19.2
14.0
10.2
18.4
44.880N 63.500W 145 44.630N 63.500W 44 46.170N 60.050W 62
-18.2 -17.2 -17.7
-16.1 -15.1 -15.2
27.9 26.2 27.6
20.4 19.5 20.3
26.2 24.6 25.9
19.4 18.6 19.4
24.6 23.0 24.2
18.6 17.9 18.6
21.8 21.1 21.6
25.7 23.9 25.8
20.8 20.1 20.5
24.2 22.7 24.2
20.6 20.1 20.1
15.5 14.9 14.9
23.4 22.1 23.7
19.7 19.2 19.2
14.7 14.0 14.0
22.3 21.2 22.4
63.750N 68.540W
34
-35.6
-34.1
18.0
12.6
14.8
11.2
12.6
9.8
13.1
17.4
11.4
14.5
10.5
7.9
14.5
9.4
7.3
13.2
44.850N 42.300N 42.250N 43.547N 43.170N 43.030N 46.360N 45.320N 44.350N 43.250N 43.460N 46.480N 46.630N 48.370N 48.570N 43.630N 43.860N 43.680N 44.120N 42.280N
79.870W 82.700W 81.900W 80.215W 79.940W 81.150W 79.420W 75.670W 78.300W 79.220W 80.380W 84.510W 80.800W 89.330W 81.380W 79.400W 79.370W 79.630W 77.530W 82.960W
183 184 178 325 238 278 370 114 216 79 322 192 348 199 295 77 198 173 86 190
-24.7 -14.2 -14.8 -21.4 -18.2 -18.2 -27.4 -24.2 -22.4 -12.8 -20.6 -24.6 -27.7 -28.9 -33.2 -15.6 -19.7 -18.0 -21.7 -15.5
-21.1 -11.9 -12.6 -18.4 -15.5 -15.6 -24.5 -21.3 -19.1 -10.8 -17.4 -21.5 -24.6 -26.2 -29.8 -13.1 -16.8 -15.5 -18.6 -13.0
29.9 31.5 26.8 29.8 30.6 30.2 27.9 30.6 30.8 29.1 30.6 28.4 29.0 29.2 29.6 28.4 31.6 31.3 29.1 32.0
23.3 24.0 22.9 21.7 22.5 22.4 20.1 21.8 21.3 22.8 21.9 21.1 20.2 20.2 20.0 21.6 22.4 22.4 22.1 23.1
28.1 29.8 25.9 28.0 29.0 28.6 26.3 29.0 28.9 27.6 28.9 26.6 27.2 27.4 27.6 26.8 29.7 29.5 27.7 30.4
22.1 23.3 22.2 20.9 21.7 21.7 19.2 20.8 20.5 22.1 21.2 20.0 19.0 19.0 18.5 21.1 21.3 21.5 21.3 22.3
26.5 28.2 25.0 26.6 27.4 27.2 24.9 27.4 27.3 26.2 27.2 25.1 25.6 25.6 25.8 25.3 28.0 27.8 26.3 29.0
21.3 22.4 21.7 20.0 20.9 20.7 18.4 20.0 19.5 21.4 20.3 19.0 18.0 18.1 17.8 20.5 20.5 20.6 20.5 21.5
24.4 26.3 24.5 23.1 23.8 23.7 21.8 23.3 22.7 24.3 23.3 22.3 21.5 21.6 21.4 23.3 23.6 23.7 23.5 24.5
28.1 29.1 25.6 27.6 28.5 28.2 25.8 28.4 28.5 27.1 28.4 26.5 26.7 26.8 27.2 26.4 29.5 29.2 27.4 29.9
23.3 25.0 23.6 22.1 22.8 22.8 20.7 22.2 21.7 23.4 22.4 21.1 20.5 20.3 20.2 22.5 22.5 22.7 22.5 23.5
26.6 28.0 24.9 26.3 27.3 27.1 24.1 26.8 27.0 26.0 27.1 25.1 25.0 25.1 25.4 25.3 27.9 27.9 26.2 28.4
23.3 25.7 24.2 21.8 22.4 22.3 20.5 21.7 20.9 23.5 21.8 20.8 19.9 19.6 19.4 22.4 21.7 22.0 22.2 22.9
18.5 21.4 19.5 17.2 17.6 17.6 15.9 16.5 15.9 18.5 17.1 15.8 15.2 14.7 14.7 17.2 16.7 17.0 17.0 18.1
26.4 27.9 25.3 25.4 26.2 26.3 23.5 26.0 25.3 26.0 25.6 24.7 23.5 24.9 24.2 25.1 26.6 26.7 25.9 27.5
22.2 24.1 23.2 20.8 21.4 21.3 19.5 20.7 20.0 22.5 20.9 19.8 18.9 18.4 18.3 21.5 20.7 21.0 21.2 22.0
17.3 19.4 18.3 16.1 16.6 16.5 14.9 15.6 15.1 17.4 16.2 14.8 14.3 13.6 13.7 16.3 15.7 16.0 16.1 17.1
25.1 26.7 24.5 24.4 25.2 25.2 22.7 24.8 24.5 25.0 24.8 23.4 22.6 23.2 22.9 24.2 25.4 25.7 24.9 26.3
46.290N 63.120W
49
-20.2
-17.7
26.9
20.8
25.4
19.7
24.1
18.8
21.8
25.4
20.8
24.0
20.5
15.3
23.9
19.6
14.4
22.9
48.330N 53.820N 48.430N 48.300N 46.180N 45.290N 45.810N 45.370N
159 223 136 152 16 44 21 181
-29.8 -36.7 -30.2 -30.2 -24.2 -24.3 -25.6 -25.4
-27.1 -34.4 -27.2 -27.7 -21.3 -21.7 -22.5 -22.2
29.4 27.0 29.0 29.1 27.9 30.1 30.5 29.4
19.4 18.7 19.8 19.7 21.1 21.6 21.9 21.5
27.4 25.3 27.1 27.1 26.4 28.6 28.8 27.9
18.5 17.7 18.9 19.0 20.3 21.0 20.9 20.7
25.6 23.5 25.3 25.2 25.2 27.1 27.2 26.5
17.8 16.9 18.2 18.3 19.6 20.2 20.1 19.8
21.1 20.3 21.7 21.6 22.6 23.6 23.5 23.1
26.4 24.7 26.3 26.4 26.0 27.8 28.3 27.3
20.0 19.0 20.7 20.5 21.7 22.5 22.4 22.1
25.1 23.4 24.8 24.9 25.0 26.4 26.7 26.1
19.2 18.7 20.2 20.1 21.4 22.1 21.9 21.8
14.3 13.9 15.1 15.1 16.1 16.9 16.6 16.8
23.2 22.3 24.0 23.8 24.7 26.1 25.9 25.4
18.3 17.4 19.1 19.1 20.6 21.2 20.9 20.7
13.4 12.8 14.2 14.1 15.3 15.9 15.6 15.7
22.2 21.0 22.7 22.6 23.9 24.9 24.8 24.3
71.000W 89.900W 71.140W 70.920W 72.920W 73.350W 73.430W 71.820W
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 N/A N/A N/A 2683 71 10.8 9.2 8.0 2911 46 7.7 5.8 4.9 3771 205 12.4 10.6 9.3 2872 22 9.3 8.0 6.8 2857 95 9.1 7.6 6.5 3007 26 5.6 4.9 4.3 2772 36 4.9 4.1 3.4 2999 76 6.2 5.1 4.1 2788 31 10.6 9.5 8.7 6496 1 1 site, 41 more on CD-ROM 12.7 11.2 10.0 5758 159 3 sites, 19 more on CD-ROM 9.9 8.6 7.7 4585 143 12.6 11.0 9.6 4665 111 12.4 10.7 9.4 4692 33 1 site, 41 more on CD-ROM 15.8 13.4 12.1 4794 33 1 site, 45 more on CD-ROM 9.3 8.3 7.4 8159 37 3 sites, 34 more on CD-ROM 12.6 10.9 9.5 4253 111 12.0 10.4 9.2 4183 72 12.5 11.0 9.7 4496 88 1 site, 60 more on CD-ROM 13.2 11.4 10.0 9468 1 21 sites, 75 more on CD-ROM 6.2 5.4 4.8 4394 212 12.9 11.3 9.9 3346 423 12.7 11.2 9.9 3612 282 8.5 7.5 6.8 4434 153 12.2 10.4 9.2 3926 249 10.7 9.4 8.4 3943 239 10.0 8.7 7.8 5151 126 10.1 8.9 7.9 4483 241 6.2 5.4 4.8 4323 177 14.1 12.4 10.9 3495 310 11.5 10.1 8.9 4241 184 10.2 8.8 7.8 4941 90 10.1 8.9 8.0 5214 124 9.5 8.4 7.5 5525 81 8.4 7.8 6.5 5982 88 13.3 11.7 10.4 3682 240 9.4 8.4 7.5 4018 264 12.2 10.6 9.3 3837 304 10.5 9.1 8.0 4139 208 11.4 10.0 9.0 3421 438 1 site, 6 more on CD-ROM 11.8 10.2 9.1 4572 108 22 sites, 81 more on CD-ROM 12.0 10.5 9.4 5607 102 9.1 8.0 7.1 7349 52 10.4 9.3 8.4 5508 98 10.2 9.0 8.0 5680 71 13.4 12.0 10.7 4572 199 10.3 8.8 7.6 4409 229 8.4 7.3 6.4 4604 211 9.0 7.9 6.9 4585 151
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station MONT-JOLI MONT-ORFORD MONTREAL MCTAVISH MONTREAL MIRABEL INTL MONTREAL ST-HUBERT MONTREAL TRUDEAU INTL NICOLET POINTE-AU-PERE INRS QUEBEC CITY JEAN LESAGE INTL SAINTE-FOY U LAVAL SHERBROOKE ST-ANICET 1 STE-ANNE-DE-BELLEVUE 1 TROIS-RIVIERES VARENNES Saskatchewan MOOSE JAW CS PRINCE ALBERT REGINA RCS SASKATOON INTL SASKATOON KERNEN FARM Yukon Territory WHITEHORSE
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 26.9 19.8 25.1 18.7 23.6 17.8 25.1 18.6 23.5 17.8 22.0 17.1 30.1 21.9 28.6 20.9 27.2 20.1 29.6 21.9 28.0 20.8 26.5 19.9 30.1 22.1 28.6 21.1 27.2 20.3 30.0 22.1 28.5 21.1 27.1 20.2 28.8 22.3 27.2 21.3 25.9 20.4 23.0 18.6 21.5 17.6 20.2 16.6 28.7 20.8 27.3 20.1 25.8 19.2 29.1 20.6 27.5 19.6 26.0 18.7 28.8 21.1 27.3 20.2 25.9 19.4 30.6 22.7 29.1 21.8 27.6 20.9 30.0 21.8 28.5 21.0 27.0 20.2 27.3 21.3 26.1 20.8 25.0 20.2 30.3 21.8 28.6 20.9 27.1 20.0
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 20.7 25.3 19.6 23.9 20.6 23.1 19.3 21.5 23.3 28.2 22.2 26.6 23.1 27.9 22.0 26.4 23.5 28.1 22.5 26.7 23.2 28.1 22.2 26.7 23.5 27.3 22.4 25.9 19.6 22.1 18.3 20.8 22.7 27.0 21.5 25.5 22.4 26.9 21.3 25.3 22.6 26.9 21.5 25.6 24.2 28.7 23.1 27.4 23.4 28.0 22.4 26.5 22.9 25.7 22.0 24.8 23.5 28.1 22.4 26.7
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 18.9 13.8 23.6 17.8 12.9 22.4 19.7 16.0 21.8 18.6 15.0 20.5 21.6 16.4 26.1 20.7 15.5 25.1 21.4 16.2 26.0 20.4 15.2 24.7 22.0 16.7 26.1 21.0 15.7 25.1 21.6 16.3 26.0 20.7 15.4 25.1 22.2 16.9 25.9 21.2 15.9 24.6 18.5 13.3 21.4 17.2 12.3 20.1 21.2 15.9 25.3 20.1 14.9 24.0 21.0 15.8 24.6 20.0 14.9 23.5 21.1 16.2 25.3 20.0 15.2 23.9 22.7 17.5 26.9 21.7 16.5 25.7 22.0 16.7 25.9 21.0 15.8 24.6 21.9 16.5 24.7 21.0 15.7 24.0 21.9 16.6 26.0 21.0 15.7 24.9
99.6% -23.7 -28.5 -22.0 -26.0 -23.5 -23.1 -25.6 -21.8 -24.5 -24.2 -27.3 -24.6 -23.8 -24.2 -23.5
99% -21.4 -25.2 -19.2 -22.9 -20.8 -20.3 -22.8 -19.2 -21.9 -21.6 -24.2 -21.7 -20.9 -21.6 -20.9
577 428 577 504 510
-30.1 -35.7 -32.3 -34.1 -33.5
-27.4 -32.4 -29.5 -31.1 -30.6
31.8 28.9 31.0 30.2 30.6
18.6 18.9 19.6 18.9 17.7
29.7 27.0 28.9 28.2 28.6
18.1 18.0 18.5 18.1 16.9
27.7 25.4 27.1 26.4 26.8
17.4 17.0 17.7 17.4 16.1
21.7 20.3 21.7 20.7 20.5
27.1 26.6 27.8 27.3 27.0
20.2 19.1 20.2 19.5 19.2
26.0 25.2 26.2 25.9 24.9
19.9 18.0 19.7 18.5 18.4
15.7 13.6 15.5 14.2 14.1
24.7 23.2 24.9 23.9 23.7
18.4 16.9 18.1 17.3 17.0
14.3 12.7 14.0 13.2 12.9
22.7 21.8 23.3 22.4 22.0
60.710N 135.070W 706
-39.7
-34.6
25.7
14.2
23.4
13.3
21.3
12.5
14.9
23.7
13.9
22.0
11.3
9.0
16.2
10.3
8.5
15.7
41.415N 19.721E
38
-2.8
-1.3
34.2
22.9
33.1
23.3
32.0
23.2
27.9
30.1
26.6
29.3
27.2
23.1
29.0
26.0
21.5
28.3
36.276N 6.620E 36.691N 3.215E 35.624N 0.621W
690 25 90
-0.5 1.8 2.2
0.5 2.9 3.8
38.9 35.2 34.3
20.2 22.4 21.1
37.0 33.5 32.5
20.2 22.5 21.3
35.0 32.0 31.0
19.8 22.5 21.5
22.5 25.6 24.8
33.1 30.7 29.5
21.7 24.8 24.2
32.2 29.8 28.6
19.6 24.1 23.2
15.6 19.1 18.2
25.7 28.0 27.2
18.8 23.2 22.9
14.8 18.0 17.8
25.3 27.4 26.9
31.324S 27.446S 32.832S 34.822S 34.559S 37.934S 31.795S 27.386S 27.450S 32.904S 24.856S 31.571S 26.841S 27.766S 31.712S
489 62 704 20 6 22 74 131 53 26 1246 597 456 200 17
-0.2 4.2 -0.7 -0.1 4.2 -1.2 2.5 4.9 1.8 -0.8 -1.1 -2.1 3.1 -0.8 0.4
1.3 5.9 0.9 1.2 5.8 0.0 3.8 6.5 3.8 1.0 0.2 -0.6 4.8 1.4 2.2
34.9 36.9 35.8 33.9 31.2 31.1 34.5 36.1 37.2 34.2 33.1 38.0 36.3 39.2 35.1
21.5 24.4 19.7 22.6 23.2 21.2 23.1 24.0 24.1 23.2 18.4 19.7 23.3 23.5 24.2
33.1 35.8 34.1 32.2 30.0 29.1 33.0 35.1 36.0 33.0 31.7 36.6 34.9 37.5 33.6
21.3 24.5 19.6 22.1 23.0 20.4 22.6 23.9 24.3 22.8 18.7 19.6 23.3 23.3 23.6
31.8 34.2 33.0 30.9 28.8 27.2 31.8 34.1 34.7 31.8 30.1 35.2 33.4 35.9 32.1
21.0 24.2 19.4 21.7 22.4 19.9 22.2 23.8 24.2 22.4 18.7 19.2 23.0 23.0 23.1
24.9 27.1 22.6 24.7 25.2 23.0 25.6 26.5 27.1 25.7 22.2 22.3 26.3 26.4 26.5
30.8 32.5 31.3 30.3 29.1 27.5 31.5 32.6 32.8 31.2 28.2 33.5 32.4 33.4 32.0
23.9 26.6 21.8 23.9 24.4 22.2 24.7 26.0 26.5 24.7 21.6 21.6 25.5 25.6 25.6
29.4 31.9 30.5 29.2 28.2 26.2 30.3 32.1 32.0 29.8 27.2 32.6 31.6 32.8 30.8
23.2 25.9 20.0 23.1 24.0 21.9 23.8 24.9 25.7 24.1 20.6 18.9 24.7 24.5 25.0
19.1 21.4 16.0 17.9 18.9 16.6 18.9 20.3 21.1 19.1 17.8 14.7 20.8 20.0 20.1
27.7 30.4 27.3 27.3 28.0 24.7 29.2 30.1 30.2 28.8 24.5 27.0 30.5 30.1 29.8
22.2 25.1 19.0 22.2 23.1 21.0 23.0 24.2 25.0 23.1 20.0 18.0 23.9 23.8 24.1
17.9 20.4 15.1 16.9 17.8 15.7 17.9 19.4 20.2 18.0 17.2 13.9 19.8 19.1 19.0
26.7 29.5 26.5 26.3 27.1 23.8 28.0 29.3 29.6 27.6 24.0 26.7 29.7 28.9 28.7
40.133N 44.467E 1140
-13.1
-10.7
36.1
21.5
34.8
21.0
33.2
20.4
22.7
34.1
21.7
33.1
19.0
15.8
30.9
17.9
14.8
29.2
34.945S 27.570S 27.383S 35.307S 33.900S 28.164S 27.933S 32.100S 34.917S 37.867S
3.9 5.4 5.8 -3.3 3.8 6.0 9.6 1.8 4.7 1.9
5.0 6.7 7.2 -2.1 4.8 7.8 10.8 3.2 5.8 3.0
36.2 32.9 30.9 34.0 32.5 29.2 30.5 36.4 37.7 35.0
18.3 22.8 22.6 17.9 20.0 23.3 23.0 20.0 19.0 18.9
33.9 31.4 29.7 31.8 30.3 28.5 29.3 34.7 35.2 32.0
17.9 22.6 23.0 17.3 20.2 23.1 22.7 19.7 18.6 18.5
31.8 30.2 28.9 29.7 28.5 27.9 28.3 32.8 33.0 29.2
17.5 22.2 22.5 16.8 20.0 22.7 22.5 19.2 18.0 18.0
21.5 25.1 25.2 20.2 23.2 25.2 25.2 23.1 21.5 21.1
28.6 29.8 28.3 27.3 27.9 27.7 27.7 30.4 30.7 28.4
20.6 24.4 24.6 19.3 22.4 24.6 24.6 21.7 20.6 20.2
28.1 28.7 27.7 26.3 26.9 27.1 27.1 29.6 29.9 27.5
19.4 23.9 24.2 18.4 21.7 24.4 24.6 21.0 19.1 19.1
14.2 18.8 19.1 14.2 16.4 19.4 19.6 15.7 14.0 13.9
24.1 26.8 27.1 21.9 25.1 26.8 26.4 25.5 24.2 23.4
18.1 23.2 23.7 17.2 21.0 23.8 23.8 19.4 17.6 18.0
13.0 18.0 18.5 13.2 15.7 18.6 18.7 14.2 12.7 13.0
23.3 26.4 26.7 21.1 24.6 26.4 25.9 24.2 23.3 22.4
48.610N 45.310N 45.510N 45.680N 45.520N 45.470N 46.230N 48.510N 46.800N 46.780N 45.440N 45.120N 45.430N 46.350N 45.720N
68.210W 52 72.240W 846 73.580W 73 74.040W 82 73.420W 27 73.740W 36 72.660W 8 68.470W 5 71.380W 60 71.290W 91 71.690W 241 74.290W 49 73.930W 39 72.520W 6 73.380W 18
50.330N 105.540W 53.210N 105.670W 50.430N 104.670W 52.170N 106.700W 52.150N 106.550W
Licensed for single user. © 2017 ASHRAE, Inc.
Albania TIRANA RINAS
Algeria CONSTANTINE BOUDIAF INTL DAR EL BEIDA ORAN INTL
Argentina CORDOBA INTL CORRIENTES INTL EL PLUMERILLO EZEIZA INTL JORGE NEWBERY INTL MAR DEL PLATA PARANA POSADAS INTL RESISTENCIA INTL ROSARIO SALTA AP SAN JUAN SAN MIGUEL DE TUCUMAN SANTIAGO DEL ESTERO SAUCE VIEJO
64.208W 58.762W 68.793W 58.536W 58.416W 57.573W 60.480W 55.971W 59.056W 60.785W 65.486W 68.418W 65.105W 64.310W 60.812W
Armenia YEREVAN EREBUNI
Australia ADELAIDE INTL BRISBANE ARCHERFIELD BRISBANE INTL CANBERRA CANTERBURY RACECOURSE GOLD COAST GOLD COAST SEAWAY JANDAKOT AP KENT TOWN LAVERTON RAAF
138.531E 6 153.008E 19 153.117E 5 149.195E 575 151.117E 3 153.505E 6 153.433E 3 115.883E 31 138.617E 51 144.750E 20
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 12.7 11.2 10.0 5284 72 15.7 13.5 12.2 5659 53 5.0 4.4 3.9 4146 304 8.3 7.2 6.2 4727 179 11.3 9.9 8.9 4414 235 11.5 10.0 8.9 4317 272 9.4 8.1 7.0 4687 168 12.8 11.3 10.0 5308 11 10.0 8.9 7.9 4879 137 9.3 8.0 6.8 4819 150 9.1 7.9 6.9 4913 103 9.2 8.1 7.1 4417 218 8.9 7.8 7.0 4404 231 10.8 9.4 8.3 4610 184 11.0 9.5 8.4 4477 207 5 sites, 47 more on CD-ROM 12.3 11.0 9.8 5392 116 9.3 8.3 7.5 6204 67 12.6 11.2 10.0 5774 111 11.1 9.8 8.7 5904 92 10.7 9.5 8.5 5903 101 1 site, 17 more on CD-ROM 10.3 9.3 8.3 6771 7 1 site, 0 more on CD-ROM 7.8 6.6 5.7 1532 687 3 sites, 38 more on CD-ROM 9.9 8.4 7.4 1658 856 10.4 9.0 7.8 984 897 11.6 9.9 8.6 901 913 15 sites, 42 more on CD-ROM 11.4 10.1 9.0 965 750 9.8 8.5 7.4 396 1635 8.1 6.8 5.8 1217 918 9.5 8.4 7.4 1181 668 11.1 9.8 8.7 891 761 10.9 9.8 8.8 1857 241 10.6 9.3 8.2 833 914 8.2 7.0 6.1 318 1772 8.9 7.6 6.7 465 1597 11.1 9.9 8.4 1012 806 7.5 6.4 5.5 928 574 13.7 11.6 10.0 1164 1157 8.7 7.1 5.9 562 1261 10.1 8.4 7.3 587 1498 14.9 12.3 10.6 809 1045 1 site, 5 more on CD-ROM 10.0 8.4 6.9 2740 771 25 sites, 500 more on CD-ROM 11.4 10.3 9.3 1156 478 9.2 8.2 7.4 361 1068 9.9 8.8 7.9 333 1014 10.5 9.4 8.4 2048 273 10.2 8.4 7.4 894 518 9.8 9.0 8.2 317 943 13.1 11.4 10.1 186 1113 10.3 9.2 8.3 940 697 8.3 7.3 6.6 1064 604 11.8 10.4 9.3 1656 238
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station MELBOURNE INTL MELBOURNE MOORABBIN MELBOURNE REGIONAL OFFICE MOUNT LOFTY NEWCASTLE NOBBYS SIGNAL PERTH INTL PERTH METRO SCORESBY RESEARCH INSTITUTE SWANBOURNE SYDNEY BANKSTOWN SYDNEY INTL SYDNEY OBSERVATORY HILL SYDNEY OLYMPIC PARK ARCHERY TUGGERANONG ISABELLA PLAINS WILLIAMTOWN
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 35.1 18.2 32.3 17.7 29.9 17.4 34.2 19.3 31.5 18.7 28.9 18.3 34.9 18.8 32.3 18.3 29.8 17.9 31.1 16.1 29.0 15.2 27.2 14.5 30.4 19.6 27.6 19.7 25.8 20.4 37.2 19.4 35.4 19.3 33.6 19.0 36.3 20.3 34.4 19.9 32.6 19.5 34.1 19.3 31.8 19.0 29.6 18.6 35.2 20.1 32.9 19.9 30.9 19.6 33.8 20.6 31.4 20.5 29.5 20.1 32.9 19.3 30.2 20.1 28.5 20.0 31.1 19.8 28.9 20.2 27.3 20.3 33.6 19.8 31.3 19.8 29.4 19.7 33.9 18.2 31.7 17.7 29.7 17.1 34.4 20.7 31.9 20.7 29.8 20.4
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 20.8 27.8 19.7 27.2 21.6 27.9 20.6 26.9 21.2 28.5 20.2 27.6 18.5 25.4 17.5 24.7 23.5 25.6 22.8 24.8 22.4 30.9 21.4 30.0 22.7 30.4 21.7 29.8 21.5 29.3 20.5 28.1 23.3 28.8 22.3 27.6 23.3 29.0 22.5 27.6 23.2 27.4 22.5 26.5 23.0 27.3 22.4 26.2 22.9 28.0 22.2 27.1 20.2 27.8 19.4 26.8 23.7 29.1 23.0 27.7
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 18.8 13.9 22.9 17.6 12.8 21.6 20.2 14.9 23.0 18.8 13.6 22.5 19.1 13.9 23.8 18.0 13.0 22.9 16.7 13.0 19.8 15.2 11.8 18.9 22.9 17.7 24.3 22.2 17.0 23.9 20.0 14.7 25.0 19.0 13.8 24.2 20.4 15.1 25.5 19.5 14.3 24.5 19.3 14.2 23.7 18.0 13.1 22.7 21.7 16.5 25.6 20.7 15.5 24.3 21.7 16.4 25.3 21.0 15.7 24.5 22.0 16.7 24.7 21.2 15.9 24.2 21.7 16.5 25.0 21.1 15.9 24.4 21.5 16.1 24.6 20.7 15.4 24.0 18.6 14.4 21.6 17.4 13.3 21.1 22.2 17.0 25.2 21.7 16.3 24.6
99.6% 2.9 2.7 4.7 2.5 7.7 3.8 3.8 2.3 6.4 3.3 6.4 7.3 5.6 -3.7 4.0
99% 3.9 4.0 5.8 3.1 8.6 5.1 5.1 3.4 7.5 4.4 7.2 8.0 6.6 -2.5 5.1
233 175 200 171 183
-9.6 -11.2 -9.3 -8.0 -10.3
-7.5 -8.4 -7.1 -5.9 -8.1
31.3 31.7 31.2 31.8 31.2
21.5 21.5 21.9 22.2 20.7
29.4 29.8 29.3 30.0 29.4
20.6 20.6 20.9 21.3 20.0
27.7 28.0 27.7 28.4 27.8
19.6 19.7 20.0 20.4 19.3
22.2 22.4 22.5 23.0 21.7
29.8 30.1 29.9 30.4 28.9
21.1 21.3 21.5 22.1 20.8
28.2 28.3 28.2 28.7 27.6
19.6 19.8 19.9 20.6 19.2
14.7 14.8 14.9 15.5 14.3
25.9 26.4 26.6 26.9 24.4
18.6 18.8 19.0 19.7 18.2
13.8 13.9 14.1 14.7 13.4
24.8 25.0 25.3 25.9 23.6
23.898E 31.017E 24.050E 27.540E 30.067E 30.350E
143 144 134 228 192 208
-17.9 -20.6 -19.4 -19.6 -21.8 -21.6
-14.2 -17.1 -15.9 -16.2 -18.6 -18.3
30.2 30.6 28.8 29.2 28.9 28.6
19.9 20.1 19.8 19.7 19.8 19.7
28.3 28.8 26.8 27.5 27.0 26.8
19.0 19.5 18.9 18.8 18.9 18.9
26.5 27.0 25.1 25.9 25.2 25.0
18.2 18.6 17.8 18.0 18.2 17.9
21.0 21.4 21.0 20.9 21.0 20.8
27.6 27.9 26.5 27.0 27.0 26.6
20.1 20.5 19.9 19.8 19.9 19.7
26.3 26.6 25.2 25.5 25.2 24.9
19.0 19.3 19.1 18.9 18.9 18.7
14.0 14.3 14.1 14.0 14.1 13.9
23.3 23.6 24.1 24.0 23.8 23.5
18.1 18.5 18.0 17.8 18.0 17.8
13.2 13.6 13.2 13.1 13.2 13.1
22.5 22.8 22.3 22.6 22.5 22.3
51.189N 4.460E 50.901N 4.484E 50.800N 4.350E
12 56 101
-6.5 -6.5 -6.1
-4.4 -4.4 -4.2
29.3 29.1 29.0
20.6 20.1 19.8
27.2 27.0 27.0
19.6 19.4 19.0
25.4 25.1 25.1
18.6 18.4 17.9
21.4 21.0 20.7
27.6 27.4 26.9
20.3 20.0 19.7
25.9 25.6 25.4
19.2 18.7 18.5
14.0 13.7 13.6
24.4 23.6 23.4
18.2 17.8 17.6
13.1 12.9 12.7
23.1 22.5 22.1
6
22.0
22.8
32.9
27.1
32.2
27.1
32.1
27.1
28.9
31.6
28.4
31.0
28.1
24.3
31.1
27.8
23.8
30.8
17.421S 66.177W 2548 16.513S 68.192W 4062 17.645S 63.135W 373
1.9 -5.0 9.2
3.1 -3.9 10.9
30.0 17.9 34.9
15.1 5.9 23.6
29.1 17.0 33.9
14.7 5.8 23.8
28.1 16.1 33.0
14.5 5.7 24.0
17.2 9.1 26.1
25.9 13.8 31.0
16.7 8.6 25.7
25.4 13.2 30.6
14.9 7.2 24.9
14.5 10.5 21.0
20.0 10.0 28.3
14.1 6.9 24.2
13.8 10.3 20.1
18.7 9.7 27.4
43.704N 18.257E 2070 43.825N 18.331E 521 43.867N 18.433E 638
-18.7 -12.9 -11.3
-16.3 -10.0 -8.9
19.5 33.1 33.0
12.0 19.7 19.6
17.9 31.1 31.0
11.4 19.6 19.2
16.6 29.2 29.1
11.0 19.2 18.6
13.8 21.6 21.6
17.0 29.5 29.1
12.8 20.6 20.3
16.1 28.3 28.0
12.3 19.0 19.0
11.5 14.7 14.9
15.4 25.8 26.5
11.3 18.0 17.6
10.8 13.8 13.7
14.3 24.1 24.2
25.528S 16.229S 10.984S 1.379S 19.851S 15.863S 20.469S 19.634S 15.653S 3.039S 27.670S 3.776S 16.632S 23.334S 0.051N
2.9 13.2 21.0 22.8 11.1 10.6 8.1 10.8 13.0 21.8 7.8 22.8 12.5 7.9 22.8
5.0 14.2 21.9 22.9 12.2 11.8 10.2 11.8 15.0 21.9 9.7 23.0 13.8 9.9 22.9
30.9 32.9 32.1 33.2 32.9 32.2 36.2 32.0 38.2 35.9 32.1 32.1 35.2 34.0 35.1
20.1 18.9 26.6 25.9 20.1 17.5 22.5 20.0 22.1 26.3 25.2 24.9 19.7 21.7 26.5
29.9 31.8 31.8 33.0 31.9 31.2 35.2 31.0 37.2 35.0 31.0 31.9 34.2 33.0 34.9
20.2 19.2 26.5 25.8 20.2 17.9 22.7 20.3 22.2 26.0 25.0 24.8 20.2 21.9 26.5
28.8 30.8 31.1 32.2 31.0 30.6 34.2 30.0 36.2 34.2 29.9 31.2 33.2 32.0 34.1
20.2 19.6 26.2 25.8 20.1 18.2 22.9 20.3 22.6 26.0 24.4 24.6 20.5 22.1 26.6
23.1 23.9 27.3 28.1 22.7 22.1 26.1 23.8 28.0 28.4 26.5 26.7 24.6 25.6 28.0
26.9 27.5 30.6 30.4 28.2 26.4 32.0 27.3 31.5 32.4 30.0 29.2 29.8 29.0 32.7
22.5 23.5 26.9 27.7 22.3 21.6 25.7 23.1 27.2 27.9 25.9 26.2 24.1 25.1 27.6
26.3 27.3 30.4 30.2 27.8 26.1 31.5 27.1 30.6 31.8 29.2 29.0 29.4 28.5 32.3
22.1 23.0 26.2 27.2 21.3 21.1 24.8 22.9 27.1 27.2 25.3 26.1 23.2 25.0 26.9
18.8 20.4 21.7 23.1 17.6 17.9 21.3 19.6 23.4 23.3 20.4 21.6 19.6 21.5 22.6
24.3 25.7 29.4 29.5 23.7 23.3 28.8 25.3 29.6 29.5 27.9 27.7 26.1 27.0 30.5
21.3 22.2 26.1 27.1 21.1 20.3 24.2 22.1 26.2 27.1 25.0 25.8 22.9 24.2 26.2
17.9 19.4 21.4 22.8 17.3 17.0 20.4 18.5 22.2 23.1 20.1 21.1 19.4 20.4 21.7
23.6 25.1 29.4 29.4 23.5 22.8 27.9 24.6 28.8 29.2 27.6 27.5 25.9 26.3 29.7
37.673S 37.976S 37.817S 34.967S 32.918S 31.940S 31.917S 37.867S 31.950S 33.924S 33.946S 33.850S 33.833S 35.417S 32.795S
144.843E 132 145.102E 15 144.967E 32 138.700E 730 151.799E 33 115.967E 20 115.867E 25 145.250E 80 115.767E 41 150.988E 9 151.177E 6 151.200E 40 151.067E 4 149.100E 588 151.834E 9
48.033N 48.321N 48.250N 48.200N 48.110N
16.283E 16.112E 16.367E 16.367E 16.570E
52.108N 52.527N 53.600N 53.864N 53.950N 55.127N
Austria GUMPOLDSKIRCHEN TULLN WIEN HOHE WARTE WIEN INNERE STADT WIEN SCHWECHAT
Licensed for single user. © 2017 ASHRAE, Inc.
Belarus BREST GOMEL GRODNO MINSK 1 MOGILEV VITEBSK
Belgium ANTWERP INTL BRUSSELS NATL UCCLE
Benin COTONOU CAJEHOUN
6.357N
2.384E
Bolivia COCHABAMBA AP EL ALTO INTL VIRU VIRU INTL
Bosnia and Herzegovina BJELASNICA SARAJEVO SARAJEVO-BJELAVE
Brazil AFONSO PENA AP ANAPOLIS AB ARACAJU AP BELEM VAL DE CANS INTL BELO HORIZONTE AP BRASILIA INTL CAMPO GRANDE AB CONFINS INTL CUIABA INTL EDUARDO GOMES INTL FLORIANOPOLIS INTL FORTALEZA INTL GOIANIA AP LONDRINA AP MACAPA INTL
49.176W 48.964W 37.070W 48.476W 43.951W 47.913W 54.673W 43.969W 56.117W 60.050W 48.547W 38.533W 49.221W 51.130W 51.072W
911 1136 7 17 789 1060 559 828 188 81 6 25 747 569 17
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 13.8 12.3 10.8 1663 268 11.6 10.4 9.3 1587 232 7.4 6.4 5.5 1258 351 15.2 13.5 11.9 2554 184 16.6 14.5 13.1 592 567 11.1 9.9 9.0 772 812 8.1 7.1 6.4 737 783 8.0 7.1 6.4 1636 271 12.8 11.0 9.6 626 713 9.7 8.5 7.6 911 559 13.1 11.7 10.5 669 662 N/A N/A N/A 632 616 9.1 7.6 6.7 736 649 8.3 7.3 6.5 2061 284 11.9 10.3 9.2 813 591 5 sites, 89 more on CD-ROM 8.0 6.6 5.7 2972 271 11.8 10.3 9.0 3098 227 9.8 8.4 7.3 2947 271 8.8 7.7 6.9 2683 386 12.2 10.8 9.6 3061 243 6 sites, 13 more on CD-ROM 8.2 7.1 6.1 3746 152 8.8 7.4 6.8 4092 179 10.4 9.1 7.9 4102 92 8.2 7.2 6.3 4259 116 9.8 8.7 7.8 4465 97 8.1 7.1 6.3 4434 108 3 sites, 16 more on CD-ROM 9.8 8.4 7.4 2771 109 11.1 9.6 8.4 2860 98 9.2 7.9 6.9 2829 113 1 site, 5 more on CD-ROM 8.0 7.4 7.0 0 3420 3 sites, 0 more on CD-ROM 9.1 7.7 5.7 511 281 8.5 7.5 6.6 3897 0 13.1 11.5 10.2 87 2198 3 sites, 5 more on CD-ROM 33.0 29.6 26.2 6013 1 8.0 6.2 4.9 3111 237 5.5 4.7 4.1 2977 288 30 sites, 21 more on CD-ROM 8.7 7.6 6.6 627 592 7.5 6.5 5.7 8 1632 8.1 7.4 6.9 0 3082 8.3 7.0 6.0 0 3400 6.3 5.5 5.1 22 1577 7.5 6.5 5.8 13 1426 10.3 9.5 8.5 64 2532 7.6 6.6 6.0 59 1243 7.7 6.4 5.6 12 3343 6.1 5.2 4.5 0 3402 8.4 7.4 6.6 216 1302 10.0 9.1 8.3 0 3363 7.9 6.8 5.6 3 2367 6.7 5.8 5.2 119 1627 8.2 7.3 6.5 0 3606
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station MACEIO INTL MANAUS AB NATAL AB PORTO ALEGRE INTL PORTO VELHO INTL RECIFE INTL RIO GALEAO TOM JOBIN INTL RIO SANTOS DUMONT AP SALVADOR INTL SAO LUIS INTL SAO PAULO CONGONHAS AP SAO PAULO INTL TERESINA AP VIRACOPOS INTL VITORIA AP
Lat
Long
Elev
9.511S 3.146S 5.911S 29.994S 8.709S 8.126S 22.809S 22.910S 12.911S 2.585S 23.627S 23.432S 5.060S 23.007S 20.258S
35.792W 59.986W 35.248W 51.171W 63.902W 34.924W 43.244W 43.163W 38.331W 44.234W 46.655W 46.470W 42.823W 47.135W 40.286W
118 81 52 3 90 10 9 3 20 54 802 750 67 661 3
42.583N 42.068N 42.695N 43.232N
23.267E 2292 24.851E 182 23.406E 531 27.825E 70
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 33.0 25.5 32.2 25.1 31.9 25.0 34.9 25.8 34.1 25.8 33.4 25.8 32.8 25.5 32.1 25.2 31.9 25.2 34.9 24.7 33.1 24.1 31.9 23.7 36.0 24.4 35.1 24.6 34.2 25.0 34.0 27.1 33.2 26.5 32.9 26.3 36.9 25.0 35.2 24.8 34.1 24.8 34.1 25.1 32.8 24.9 31.8 24.8 32.3 26.6 32.0 26.5 31.2 26.2 33.9 26.1 33.2 25.8 32.9 25.6 32.2 20.2 31.1 20.2 30.1 20.2 32.8 21.5 31.2 21.3 30.2 21.2 38.8 22.8 38.0 23.2 37.1 23.5 33.2 20.9 32.2 21.0 31.2 21.0 34.1 25.5 33.2 25.2 32.6 25.0
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 26.9 30.3 26.5 29.8 27.2 31.5 27.0 31.3 26.7 29.9 26.4 29.7 26.4 31.7 25.7 30.6 28.0 31.0 27.6 30.8 27.6 32.4 27.1 31.9 28.0 32.1 27.3 31.3 26.6 31.0 26.2 30.4 27.5 30.8 27.1 30.5 27.7 30.7 27.2 30.5 23.2 27.7 22.6 27.1 24.4 28.3 23.8 27.6 27.2 31.2 26.9 31.2 24.1 28.6 23.6 28.2 27.2 30.6 26.7 30.1
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 26.1 21.8 28.4 25.8 21.3 28.2 26.2 21.8 29.2 26.0 21.5 29.1 26.1 21.6 28.3 25.5 20.8 27.9 25.1 20.2 28.4 24.2 19.1 27.5 27.2 23.2 29.0 27.0 22.9 28.8 26.2 21.6 30.7 25.9 21.3 30.4 27.0 22.8 30.0 26.2 21.6 29.2 25.2 20.4 28.9 25.0 20.1 28.7 26.5 22.1 29.6 26.1 21.5 29.4 27.0 22.9 29.4 26.3 22.0 28.7 22.0 18.4 25.5 21.2 17.5 24.5 23.8 20.4 25.5 22.9 19.4 25.0 26.2 21.8 28.5 26.1 21.6 28.4 23.1 19.3 25.1 22.2 18.3 24.6 26.2 21.7 28.5 26.0 21.3 28.3
99.6% 18.8 22.3 20.8 3.9 18.8 21.3 14.9 16.2 20.8 22.8 9.0 7.8 21.9 8.9 16.8
99% 19.1 22.9 21.1 5.8 20.0 22.0 15.9 17.1 21.2 23.1 10.2 9.1 22.6 10.2 17.8
-19.1 -10.2 -12.8 -9.1
-17.2 -7.5 -10.0 -6.9
17.3 34.8 33.0 31.9
11.0 20.7 19.0 22.3
15.9 33.1 31.0 30.2
10.5 20.6 18.7 22.1
14.7 31.8 29.2 29.0
10.1 20.1 18.5 21.6
12.4 22.7 20.9 24.5
15.1 31.0 28.4 29.1
11.6 21.8 20.1 23.6
14.2 29.7 27.4 28.1
11.3 20.1 18.3 23.1
11.1 15.2 14.0 18.0
13.5 25.9 23.6 27.4
10.5 19.2 17.8 22.1
10.4 14.3 13.6 16.9
12.7 24.7 23.1 26.6
Bulgaria CHERNI VRAH PLOVDIV SOFIA VARNA
Burkina Faso BOBO DIOULASSO OUAGADOUGOU INTL
11.160N 4.331W 12.353N 1.512W
461 316
18.4 16.2
19.6 17.3
38.1 40.9
20.3 20.3
37.4 40.0
20.3 20.4
36.6 39.0
20.4 20.6
26.1 26.5
32.1 33.4
25.6 26.1
31.5 32.9
24.5 25.0
20.6 20.9
28.8 28.5
24.1 24.2
20.1 19.9
28.3 28.2
12.134N 15.034E
295
13.1
14.8
43.0
21.9
42.1
21.8
41.1
21.6
28.3
34.4
27.6
33.7
27.0
23.5
30.9
26.1
22.3
30.4
23.444S 70.445W 139 33.393S 70.786W 474
10.0 -1.1
10.9 0.1
24.2 31.9
18.9 17.2
23.5 30.8
18.4 17.1
23.0 29.8
18.0 17.0
20.0 19.0
22.8 28.9
19.2 18.4
22.3 28.4
18.9 15.0
13.9 11.3
22.1 22.2
18.0 14.1
13.2 10.6
21.0 21.5
30.617N 116.967E 36.050N 114.133E 40.851N 111.824E 38.733N 115.483E 34.350N 107.133E 40.080N 116.585E 32.850N 117.300E 41.317N 123.783E 38.333N 116.833E 43.996N 125.681E 29.050N 111.683E 28.117N 112.783E 41.550N 120.450E 40.967N 117.917E 30.579N 103.947E 42.300N 118.833E 29.719N 106.642E 38.966N 121.539E 40.050N 124.333E 40.083N 113.417E 37.433N 116.317E 25.935N 119.663E 34.850N 119.133E 23.392N 113.299E 26.539N 106.801E 20.000N 110.250E 30.228N 120.432E 45.933N 126.567E 31.780N 117.298E
-2.1 -8.7 -23.1 -10.3 -6.1 -11.2 -5.1 -22.3 -9.4 -25.5 -1.0 -1.2 -19.1 -18.2 0.8 -20.4 2.8 -12.2 -16.1 -20.7 -8.3 4.6 -7.1 5.8 -3.1 10.7 -2.1 -27.8 -4.2
-0.9 -6.8 -20.1 -8.4 -4.6 -9.5 -3.5 -20.1 -7.7 -23.1 0.1 0.0 -16.9 -16.4 1.9 -18.6 3.8 -10.4 -14.1 -18.8 -6.8 5.8 -5.5 6.9 -1.8 12.3 -1.0 -25.5 -2.9
35.7 35.4 32.1 35.3 35.1 35.0 35.6 31.3 34.2 31.0 36.5 36.4 33.7 33.1 33.9 32.9 37.2 31.2 30.0 32.0 34.2 35.6 33.3 35.9 30.3 35.1 36.6 31.4 35.6
27.2 23.2 17.4 22.6 21.7 22.0 26.5 21.8 23.2 20.6 26.7 26.1 21.1 20.4 25.0 19.5 25.0 23.4 23.6 17.5 24.3 26.7 26.4 26.2 21.2 26.8 26.3 20.4 27.7
34.7 34.0 30.5 33.7 33.7 33.5 34.4 30.1 33.0 29.6 35.3 35.3 32.1 31.5 32.8 31.2 35.9 30.0 28.6 30.4 32.9 34.4 31.8 34.9 29.3 34.2 35.5 29.9 34.4
27.1 23.9 17.1 22.9 21.4 22.2 26.1 21.7 23.5 20.9 26.7 26.1 21.2 20.4 24.5 19.1 25.1 22.9 23.0 17.1 24.3 26.5 25.8 26.1 21.0 26.7 26.3 20.9 27.3
33.6 32.7 29.1 32.4 32.2 32.1 33.0 28.9 31.9 28.3 34.1 34.2 30.9 30.2 31.2 29.9 34.5 28.9 27.5 29.0 31.9 33.2 30.5 33.9 28.3 33.5 34.2 28.6 33.2
26.7 23.9 16.8 23.0 21.4 22.3 25.4 21.5 23.4 20.6 26.4 26.0 20.9 20.2 23.9 18.7 25.0 22.6 22.6 17.0 24.0 26.3 25.5 26.1 20.8 26.6 26.2 20.6 26.8
28.5 27.8 21.0 27.0 24.8 27.0 28.3 24.6 27.4 24.2 28.5 27.8 25.5 24.3 27.0 22.7 27.2 26.1 25.7 21.2 27.8 27.7 28.2 27.7 23.0 28.0 28.1 24.1 28.9
33.1 31.7 27.3 31.3 30.9 30.5 32.8 28.5 31.0 27.8 33.3 32.9 29.8 28.9 31.0 28.8 32.5 28.7 27.9 26.8 31.5 33.2 31.6 31.7 27.7 32.5 32.8 28.1 33.5
28.0 27.0 20.1 26.2 24.0 26.1 27.7 23.9 26.6 23.3 27.9 27.3 24.6 23.5 26.2 21.8 26.7 25.4 24.9 20.4 26.9 27.2 27.6 27.4 22.5 27.6 27.5 23.3 28.2
32.6 30.8 25.8 30.2 29.9 29.5 32.2 27.7 30.0 26.9 32.7 32.4 28.5 28.0 30.1 27.4 32.0 27.8 26.8 25.9 30.5 32.2 30.7 31.3 26.9 31.9 32.3 27.1 32.7
27.3 26.8 19.2 25.9 23.1 26.1 27.1 23.5 26.4 23.1 27.3 26.6 24.3 23.0 26.0 20.9 26.0 25.2 25.1 19.7 26.8 26.2 27.3 26.9 21.7 26.9 27.0 22.9 27.8
23.3 22.9 15.9 21.3 19.3 21.6 23.0 18.7 21.9 18.3 23.2 22.5 19.6 18.7 22.7 16.7 22.5 20.5 20.2 16.4 22.4 21.7 23.1 22.6 18.8 22.6 22.8 17.9 24.0
31.3 30.4 23.3 29.7 28.3 29.0 31.1 26.8 29.8 26.4 31.6 30.5 28.1 26.9 29.6 25.8 30.0 27.4 26.9 23.5 30.3 30.7 30.6 29.6 25.2 30.0 30.1 26.3 31.9
26.8 26.0 18.2 25.0 22.3 25.1 26.5 22.7 25.6 22.2 26.6 26.0 23.4 22.1 25.1 20.1 25.4 24.8 24.3 18.7 25.9 26.0 26.7 26.2 21.2 26.6 26.2 22.1 27.1
22.6 21.8 14.9 20.1 18.3 20.3 22.1 17.8 20.9 17.3 22.3 21.7 18.6 17.7 21.5 15.8 21.7 19.9 19.2 15.4 21.3 21.4 22.3 21.7 18.2 22.2 21.6 17.0 22.9
30.9 29.6 23.4 29.0 27.5 28.1 30.7 26.1 28.9 25.6 31.0 30.0 27.1 26.0 28.7 25.0 29.5 27.0 26.0 23.1 29.6 30.2 29.9 29.2 24.5 29.7 29.7 25.7 31.2
Licensed for single user. © 2017 ASHRAE, Inc.
Chad N'DJAMENA INTL
Chile CERRO MORENO INTL SANTIAGO PUDAHUEL INTL
China ANQING ANYANG BAITA BAODING BAOJI BEIJING CAPITAL INTL BENGBU BENXI CANGZHOU CHANGCHUN LONGJIA CHANGDE CHANGSHA CHAOYANG CHENGDE CHENGDU SHUANGLIU CHIFENG CHONGQING JIANGBEI INTL DALIAN INTL DANDONG DATONG DEZHOU FUZHOU CHANGLE INTL GANYU GUANGZHOU BAIYUN INTL GUIYANG LONGDONGBAO INTL HAIKOU HANGZHOU XIAOSHAN INTL HARBIN HEFEI LUOGANG
63 196 1084 17 610 35 28 185 11 215 35 120 176 423 495 567 416 33 14 1054 22 14 10 15 1139 24 7 118 33
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 7.8 7.1 6.5 0 2687 5.7 5.1 4.4 0 3455 10.0 9.2 8.4 0 3114 9.4 8.1 7.0 488 1139 6.0 5.1 4.3 1 3343 8.7 8.0 7.3 0 3371 8.4 7.4 6.4 6 2406 8.4 7.4 6.5 4 2268 8.9 8.1 7.3 0 3018 9.3 8.4 7.7 0 3603 8.0 7.0 6.2 212 1163 7.3 6.4 5.8 220 1064 5.2 4.4 4.0 0 4030 11.2 10.2 9.4 99 1421 10.3 9.3 8.3 0 2609 4 sites, 33 more on CD-ROM 28.3 24.1 19.9 6389 0 11.2 9.7 8.3 2531 559 9.4 8.1 7.0 3042 298 12.3 10.1 8.4 2537 434 2 sites, 7 more on CD-ROM 7.4 6.6 6.1 0 3458 7.6 6.7 6.0 0 3867 1 site, 1 more on CD-ROM 9.2 8.0 7.1 1 3892 2 sites, 11 more on CD-ROM 9.0 8.2 7.5 712 175 8.5 7.8 7.0 1486 264 87 sites, 327 more on CD-ROM 7.9 6.9 6.2 1575 1328 7.5 6.4 5.5 2357 993 9.0 7.6 6.4 4428 339 6.4 5.3 4.4 2632 959 6.2 5.2 4.4 2342 816 10.0 8.3 6.8 2845 871 7.2 6.2 5.4 1902 1146 6.5 5.4 4.7 4091 496 8.7 7.3 6.2 2649 927 10.4 8.8 7.5 4834 422 6.2 5.1 4.3 1476 1335 6.6 5.6 5.0 1460 1389 8.3 7.2 6.2 3689 643 7.1 5.6 4.5 3811 520 5.7 4.5 3.8 1352 1018 8.2 7.0 6.0 4218 445 5.4 4.6 4.0 1171 1301 10.3 9.1 8.1 3083 641 8.5 7.4 6.4 3610 458 9.4 8.1 6.9 4201 356 7.4 6.4 5.5 2496 966 10.4 9.1 8.1 703 1641 7.3 6.3 5.4 2306 858 7.2 6.1 5.3 385 2106 6.6 5.7 5.1 1734 656 7.9 6.7 5.9 104 2528 7.0 6.0 5.2 1561 1293 7.6 6.6 5.8 5218 400 7.2 6.2 5.4 1822 1193
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station JINAN TSINAN JINGDEZHEN JINGZHOU JIANGLING JINZHOU JIXI KUNMING WUJIABA LANZHOU LIANGJIANG LINGXIAN LIUZHOU MENGJIN MUDANJIANG NANCHANG CHANGBEI INTL NANJING LUKOU NANNING WUXU INTL NEIJIANG QINGDAO INTL QINGJIANG QIQIHAR SANJIAZI SHANGHAI BAOSHAN SHANGHAI HONGQIAO INTL SHANTOU SHAOGUAN SHENGYANG TAOXIAN SHENYANG SHENZHEN BAO'AN INTL SHIJIAZHUANG SIPING TAI SHAN TAIYUAN WUSU INTL TANGSHAN TIANJIN TIANJIN BINHAI INTL URUMQI DIWOPU WEIFANG WENZHOU WUHAN TIANHE WUHUXIAN WULUMUQI XIAMEN GAOQI INTL XIANYANG XIHUA XINGTAI XINING XINYANG XINZHENG INTL XUZHOU YANGJIANG YANJI YICHANG YINCHUAN YINGKOU YUEYANG YUNCHENG ZHANGJIAKOU ZHANJIANG ZHAOQING GAOYAO ZUNYI
Lat
Long
36.683N 116.983E 29.300N 117.200E 30.333N 112.183E 41.133N 121.117E 45.300N 130.917E 24.992N 102.744E 36.050N 103.883E 25.218N 110.039E 37.333N 116.567E 24.350N 109.400E 34.817N 112.433E 44.500N 129.667E 28.865N 115.900E 31.742N 118.862E 22.608N 108.172E 29.583N 105.050E 36.266N 120.374E 33.600N 119.033E 47.240N 123.918E 31.400N 121.467E 31.198N 121.336E 23.400N 116.683E 24.667N 113.600E 41.640N 123.483E 41.733N 123.517E 22.639N 113.811E 38.067N 114.350E 43.183N 124.333E 36.256N 117.105E 37.747N 112.628E 39.650N 118.100E 39.100N 117.167E 39.124N 117.346E 43.907N 87.474E 36.767N 119.183E 28.017N 120.667E 30.784N 114.208E 31.150N 118.583E 43.783N 87.617E 24.544N 118.128E 34.447N 108.752E 33.783N 114.517E 37.183N 114.367E 36.617N 101.767E 32.133N 114.050E 34.520N 113.841E 34.283N 117.150E 21.867N 111.967E 42.883N 129.500E 30.733N 111.367E 38.483N 106.217E 40.667N 122.200E 29.383N 113.083E 35.117N 111.067E 40.783N 114.883E 21.217N 110.400E 23.050N 112.467E 27.700N 106.883E
Elev 58 60 33 70 274 1895 1518 174 19 97 333 307 44 15 128 357 10 19 145 4 3 3 68 60 43 4 105 167 1536 785 29 5 3 648 22 7 34 16 919 18 479 53 184 2262 115 151 42 22 178 258 1112 4 52 377 726 28 12 845
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB 99.6% -8.4 -1.3 -1.7 -16.0 -24.8 0.7 -11.3 1.0 -10.6 3.4 -6.6 -26.5 -0.7 -4.8 4.8 2.5 -8.2 -6.0 -28.1 -2.1 -2.9 7.2 2.2 -24.1 -23.1 7.0 -8.3 -23.4 -16.8 -14.9 -13.5 -10.6 -11.0 -24.1 -10.7 1.3 -2.5 -3.1 -22.1 6.8 -8.1 -5.6 -7.3 -17.3 -4.6 -6.4 -6.4 7.0 -22.3 -1.0 -16.4 -17.6 -1.2 -8.7 -16.9 7.6 6.3 -0.7
99% -6.6 -0.1 -0.6 -14.1 -22.9 1.9 -9.8 2.2 -8.6 4.7 -5.2 -24.4 0.4 -3.2 6.1 3.6 -6.8 -4.1 -26.0 -0.7 -1.2 8.7 3.5 -21.9 -20.7 8.5 -6.7 -21.3 -14.7 -12.9 -11.5 -8.8 -9.1 -21.9 -9.1 2.8 -1.1 -1.8 -20.0 7.8 -6.2 -4.1 -5.9 -15.5 -3.3 -5.0 -4.7 8.3 -20.2 0.1 -14.2 -15.6 0.0 -6.8 -15.2 9.0 7.5 0.4
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 35.1 23.2 33.8 23.4 32.6 23.1 36.2 26.4 35.2 26.3 34.2 26.0 35.0 27.7 34.1 27.2 33.1 26.6 31.8 21.8 30.5 21.7 29.3 21.5 30.5 20.6 28.9 20.3 27.5 20.1 28.0 16.4 27.0 16.4 26.0 16.6 32.6 18.0 31.1 17.4 29.7 16.8 35.1 25.4 34.1 25.3 33.1 25.3 35.1 23.6 33.6 23.9 32.2 24.0 35.3 25.6 34.5 25.6 33.7 25.5 35.0 22.0 33.5 22.3 32.1 22.4 31.3 21.1 29.8 20.5 28.3 20.2 35.9 26.6 34.9 26.6 33.9 26.4 35.7 26.7 34.4 26.5 33.1 26.1 35.0 26.2 34.1 26.1 33.2 25.9 35.3 26.0 34.1 25.5 32.8 25.1 32.8 23.8 31.1 23.7 29.9 23.2 33.8 27.1 32.7 26.5 31.5 25.7 31.9 20.7 30.2 20.3 28.8 20.1 35.3 26.6 34.0 26.5 32.8 26.2 35.9 27.1 34.2 27.1 33.1 26.7 34.2 27.1 33.2 27.0 32.2 26.7 35.3 25.8 34.4 25.7 33.5 25.5 31.9 22.9 30.8 22.9 29.8 22.3 31.5 23.1 30.3 22.5 29.3 22.2 34.0 26.3 33.1 26.3 32.4 26.2 36.1 21.9 34.4 22.7 33.0 22.9 31.0 21.4 29.8 21.4 28.7 21.1 22.7 17.2 21.7 17.4 20.9 17.7 33.5 19.8 32.1 19.9 30.9 19.6 33.5 23.0 32.1 23.0 31.0 22.7 34.4 23.2 33.0 23.3 31.8 23.0 34.3 23.2 33.1 23.1 32.0 23.0 35.2 17.9 34.0 17.7 32.8 17.4 34.4 24.1 33.0 24.0 31.6 23.6 33.9 27.3 32.9 27.0 32.0 26.7 36.1 27.6 35.0 27.3 34.0 27.0 36.2 27.4 35.1 27.1 33.8 26.7 33.4 16.3 31.9 16.1 30.4 15.8 34.8 26.3 33.8 26.3 32.8 26.1 36.2 23.0 34.9 22.9 33.2 22.7 35.2 25.6 33.9 25.8 32.7 25.2 35.9 22.3 34.4 22.8 33.0 23.1 27.6 15.1 26.0 14.3 24.5 13.6 34.8 26.4 33.6 25.9 32.4 25.3 35.3 23.5 34.0 23.9 32.8 23.9 34.8 25.8 33.6 25.4 32.4 24.8 33.1 26.5 32.2 26.3 31.5 26.2 31.1 21.5 29.4 20.9 27.9 20.3 35.9 26.6 34.6 26.0 33.3 25.5 32.7 18.8 31.4 18.6 30.1 18.0 30.5 24.0 29.4 23.5 28.5 23.1 34.5 27.2 33.7 26.8 32.9 26.6 36.5 22.4 35.1 22.5 33.7 22.2 33.1 18.7 31.5 18.6 30.0 18.4 33.9 26.6 33.1 26.6 32.4 26.6 35.1 26.1 34.3 26.0 33.4 26.0 32.7 22.7 31.6 22.5 30.6 22.3
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 27.0 31.8 26.4 30.9 27.7 33.4 27.3 32.7 28.6 33.3 28.1 32.7 25.6 28.6 24.8 27.6 23.5 27.5 22.6 26.3 19.8 24.3 19.4 23.7 20.3 28.6 19.5 27.5 27.1 31.4 26.7 30.9 27.8 31.2 27.0 30.5 27.2 32.2 26.8 31.7 26.7 31.0 25.9 29.9 23.5 28.3 22.6 27.0 28.2 32.7 27.7 32.2 28.2 32.8 27.7 32.1 27.7 31.9 27.3 31.3 27.4 32.8 26.8 31.9 26.8 29.4 26.2 28.5 28.3 32.2 27.7 31.4 23.8 27.8 22.9 27.0 27.8 32.4 27.3 31.6 28.7 32.7 28.0 31.9 28.6 31.8 28.1 31.0 27.2 32.2 26.7 31.5 25.9 29.5 25.0 28.5 25.6 29.1 24.7 28.2 28.8 31.0 28.2 30.4 27.1 31.4 26.2 30.5 24.7 28.1 24.0 27.3 20.9 21.2 20.2 20.5 24.2 29.2 23.2 28.2 26.7 30.4 25.8 29.3 27.2 30.9 26.3 29.8 27.6 30.5 26.7 29.7 20.2 30.0 19.4 29.2 27.5 31.6 26.6 30.2 28.0 32.6 27.6 31.6 29.2 33.5 28.6 32.7 28.6 33.7 28.0 33.1 18.0 28.2 17.4 27.6 27.8 31.2 27.3 30.8 26.3 32.2 25.5 31.2 28.6 32.8 27.8 31.8 27.3 31.7 26.5 30.6 17.1 24.1 16.1 22.5 27.8 32.5 27.2 31.8 27.9 31.6 27.1 30.8 28.1 32.6 27.5 31.6 27.7 30.4 27.5 30.1 23.9 28.4 22.9 26.5 28.1 33.1 27.4 32.3 22.1 28.5 21.1 27.6 25.8 28.7 25.1 28.0 28.3 32.8 27.7 32.1 25.9 32.4 25.2 31.6 22.7 28.4 21.9 27.3 28.1 31.1 27.7 30.7 27.5 32.1 27.2 31.5 24.2 29.7 23.7 29.0
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Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 25.7 21.1 29.8 25.1 20.3 29.4 26.3 22.0 30.3 25.9 21.4 29.9 27.4 23.4 31.8 26.8 22.6 31.3 24.8 19.9 27.5 23.9 18.9 26.6 22.2 17.5 25.5 21.4 16.6 24.8 18.6 17.0 21.5 18.2 16.5 21.1 17.8 15.4 24.6 16.9 14.5 23.5 26.1 22.0 28.9 25.7 21.5 28.6 26.9 22.6 30.1 26.0 21.4 29.2 26.0 21.6 29.4 25.6 21.1 29.2 25.6 21.7 29.3 24.8 20.7 28.3 22.0 17.3 26.2 21.2 16.5 25.4 27.2 23.0 30.6 26.6 22.3 30.3 27.1 22.9 30.9 26.5 22.1 30.4 26.7 22.7 29.7 26.2 22.0 29.2 26.0 22.3 30.5 25.5 21.6 29.8 26.1 21.5 28.3 25.5 20.7 27.7 27.2 23.0 30.8 26.6 22.2 30.3 22.6 17.7 26.2 21.6 16.6 25.3 26.6 22.2 30.4 26.2 21.6 29.8 27.9 24.0 31.0 27.1 22.8 30.3 28.0 24.1 30.6 27.2 23.0 29.7 25.9 21.5 29.2 25.6 21.0 28.9 24.9 20.1 28.6 23.9 19.0 27.5 24.5 19.6 27.8 23.7 18.6 26.9 28.2 24.4 30.2 27.8 23.9 30.0 26.0 21.6 30.1 25.0 20.4 29.1 23.7 18.9 26.8 22.9 18.0 26.1 20.8 18.7 21.0 20.1 17.9 20.4 22.9 19.4 27.1 21.8 18.2 25.9 25.6 20.9 29.2 24.8 19.9 28.2 26.2 21.6 29.7 25.3 20.4 28.9 27.0 22.6 29.6 26.0 21.4 28.5 17.8 13.8 22.2 16.2 12.4 22.3 26.3 21.8 30.0 25.6 20.9 29.1 26.8 22.4 30.5 26.5 22.0 29.9 28.1 24.4 31.9 27.6 23.7 31.4 27.4 23.2 31.6 26.7 22.3 30.9 15.1 12.0 20.2 14.1 11.2 20.3 27.1 22.8 29.5 26.4 21.9 29.1 24.9 21.2 29.9 24.0 20.0 29.1 27.5 23.5 31.5 26.7 22.5 30.6 26.1 22.0 30.1 25.3 20.9 29.6 15.0 14.1 19.3 14.1 13.3 18.5 26.6 22.5 30.6 26.0 21.7 30.1 27.0 23.1 30.5 26.1 21.9 29.6 26.9 22.7 31.0 26.3 21.8 30.3 27.1 22.9 29.0 26.7 22.4 28.8 22.5 17.6 26.4 21.6 16.7 25.3 26.8 23.2 31.1 26.2 22.3 30.3 20.1 17.0 25.6 19.1 15.9 24.9 24.9 20.0 28.0 24.2 19.1 27.2 27.1 23.0 31.7 26.5 22.1 31.0 24.1 19.9 29.8 23.3 18.9 29.1 21.1 17.2 25.8 20.1 16.2 25.3 27.2 23.1 29.5 27.0 22.7 29.3 26.4 22.0 29.5 26.1 21.5 29.2 22.7 19.4 26.5 22.3 18.9 26.1
Extreme Annual WS 1% 2.5% 5% 8.9 7.6 6.5 5.3 4.4 3.8 6.7 5.8 5.1 8.9 7.6 6.6 11.0 9.4 8.2 9.1 7.9 6.9 4.3 3.5 3.1 7.2 6.3 5.4 8.2 7.0 6.1 5.1 4.4 3.8 8.7 7.2 6.1 8.7 7.2 6.0 5.7 5.0 4.3 7.6 6.7 5.9 6.4 5.4 4.6 5.1 4.3 3.5 10.6 9.4 8.3 6.5 5.5 4.9 8.9 7.5 6.5 7.5 6.6 6.0 9.2 8.1 7.3 8.3 7.2 6.2 7.1 6.1 5.3 10.2 8.7 7.5 9.5 8.1 6.9 8.0 7.1 6.3 5.8 4.8 4.1 8.5 7.3 6.2 17.8 15.8 14.1 9.3 7.7 6.4 7.9 6.5 5.4 9.1 7.4 6.1 10.3 8.8 7.5 7.3 5.7 4.7 9.0 7.7 6.7 6.4 5.5 4.9 7.0 5.9 5.1 7.5 6.4 5.5 7.5 5.9 4.9 8.8 7.7 6.9 8.2 7.0 6.0 5.5 4.7 4.1 5.4 4.5 4.0 5.4 4.4 3.5 7.8 6.6 5.7 8.3 6.8 5.7 6.3 5.4 4.8 8.7 7.4 6.5 10.3 8.9 7.5 4.4 3.8 3.3 8.2 6.3 5.2 9.9 8.6 7.7 7.1 6.2 5.5 8.7 7.3 6.3 7.4 6.2 5.3 8.1 6.9 6.1 6.9 5.8 5.0 4.6 3.9 3.3
Heat./Cool. Degree-Days HDD / CDD 18.3 2266 1079 1326 1426 1577 1241 3494 618 5226 285 1151 356 3064 451 1030 1503 2579 921 701 1889 2193 931 5149 345 1361 1464 1856 1133 493 1959 1218 1160 2494 800 2093 964 5397 407 1597 1193 1600 1253 337 1885 774 1729 4149 562 4112 546 252 2219 2423 1042 4469 480 4471 46 3191 559 2971 803 2728 934 2759 921 4299 774 2693 814 1076 1306 1599 1327 1699 1227 4346 543 452 1771 2358 906 2067 1026 2327 1062 4199 49 1891 1070 2157 1001 2127 1039 279 2037 4735 296 1472 1224 3449 512 3629 615 1466 1334 2321 1045 3707 545 221 2229 381 2087 1635 856
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
99.6%
99%
17.8 3.6 22.8 10.1 23.2
18.0 5.0 23.1 11.0 23.9
32.2 21.2 34.2 23.9 33.1
22.1 13.4 27.3 15.9 27.6
31.7 20.8 33.8 23.2 32.2
22.0 13.3 27.2 15.7 27.2
31.0 20.1 33.1 23.0 32.1
21.9 13.2 27.0 15.6 27.1
23.3 15.2 28.7 17.5 28.3
29.5 18.8 31.7 21.3 31.5
22.8 14.9 28.2 17.1 28.0
29.2 18.4 31.3 21.0 31.2
21.5 14.1 28.0 16.2 27.2
18.2 13.8 24.3 15.0 23.0
26.2 16.6 30.0 18.3 30.5
21.1 13.8 27.4 16.1 27.1
17.7 13.5 23.3 14.9 22.8
25.5 16.3 29.6 18.1 30.4
319
18.0
18.9
34.2
24.6
33.5
24.6
32.9
24.5
26.0
31.1
25.7
30.6
24.9
20.8
28.1
24.2
19.9
27.5
9.994N 84.209W 921
16.8
17.3
30.8
20.8
30.0
20.5
29.2
20.4
24.2
26.8
23.7
26.5
23.3
20.3
25.5
22.9
19.8
25.2
5.261N 3.926W
6
21.2
22.0
32.9
27.5
32.2
27.2
31.8
27.0
28.9
31.2
28.5
30.7
28.2
24.5
29.7
28.0
24.2
29.6
45.817N 16.033E 45.743N 16.069E
128 108
-9.8 -11.1
-7.0 -8.2
32.6 32.8
21.5 21.8
30.9 31.1
20.9 21.2
29.3 29.5
20.3 20.7
22.5 23.0
30.0 30.2
21.7 22.2
29.0 29.0
20.1 20.9
15.0 15.7
25.6 26.4
19.1 20.0
14.1 14.8
25.0 25.5
21.420N 77.848W 126 22.989N 82.409W 64 19.970N 75.835W 76
14.9 10.1 18.5
16.2 12.0 19.5
33.8 33.1 32.1
23.6 25.0 25.4
33.1 32.7 31.5
23.8 25.0 25.4
32.4 32.1 31.1
23.8 25.0 25.4
26.2 27.7 27.3
29.9 30.4 29.8
25.7 27.0 26.9
29.6 30.2 29.7
25.2 27.0 26.9
20.6 22.9 22.7
27.5 29.5 29.0
24.9 26.2 26.1
20.2 21.7 21.7
27.3 28.9 28.7
49.151N 49.696N 50.117N 50.000N 50.101N
16.694E 18.111E 14.533E 14.450E 14.260E
237 257 286 304 380
-12.1 -14.9 -12.6 -12.0 -13.1
-9.8 -11.9 -10.0 -9.5 -10.4
30.8 30.4 29.7 30.6 29.6
20.2 20.2 19.5 19.2 19.2
29.0 28.4 27.8 28.7 27.7
19.4 19.5 18.9 18.6 18.5
27.2 26.7 26.1 26.8 25.8
18.6 18.6 18.2 17.8 17.8
21.3 21.2 20.9 20.4 20.2
28.3 28.1 27.0 27.7 26.9
20.4 20.2 19.9 19.5 19.3
27.1 26.7 25.6 26.3 25.7
19.0 18.9 18.9 18.1 18.0
14.2 14.2 14.2 13.5 13.5
24.3 23.6 22.8 22.4 23.1
18.0 18.0 18.1 17.3 17.1
13.3 13.4 13.4 12.8 12.8
23.3 22.8 21.9 21.5 21.6
55.533N 55.618N 55.586N 55.767N
12.717E 12.656E 12.131E 12.343E
6 5 45 18
-5.3 -6.9 -9.1 -9.8
-4.0 -5.2 -7.0 -7.1
22.3 25.6 25.9 26.6
18.4 18.3 18.3 18.4
21.2 24.1 24.1 24.9
17.8 17.7 17.7 17.9
20.2 22.6 22.7 23.0
17.2 17.0 17.0 17.2
19.2 19.7 19.9 19.9
21.3 23.3 23.3 23.3
18.4 18.7 18.8 18.9
20.4 22.2 22.3 22.9
18.3 18.4 18.7 18.9
13.2 13.2 13.6 13.7
20.5 21.3 21.6 21.4
17.6 17.5 17.5 17.8
12.6 12.5 12.6 12.8
19.7 20.4 20.5 20.3
18.430N 69.669W 18.433N 69.883W
18 14
18.0 19.8
18.9 20.5
33.0 32.6
26.4 27.2
32.2 32.1
26.3 27.1
32.0 31.6
26.2 27.0
28.1 28.5
31.4 31.4
27.6 28.0
31.0 31.1
27.1 27.7
22.9 23.7
30.9 30.9
26.7 27.1
22.3 22.9
30.5 30.5
18.9 6.8
19.2 7.3
33.0 21.9
23.9 11.8
32.1 21.1
24.0 11.9
31.8 20.8
24.0 11.9
26.5 14.5
29.8 18.8
25.8 14.1
29.4 18.4
25.8 13.0
21.1 13.3
28.8 16.0
24.9 12.4
20.0 12.7
27.5 15.1
Colombia ALFONSO BONILLA ARAGON INTL ELDORADO INTL ERNESTO CORTIZZOS INTL JOSE MARIA CORDOVA INTL RAFAEL NUNEZ AP
3.543N 4.702N 10.890N 6.165N 10.442N
76.382W 964 74.147W 2548 74.781W 30 75.423W 2142 75.513W 1
Congo BRAZZAVILLE MAYA MAYA INTL
4.252S 15.253E
Costa Rica JUAN SANTAMARIA INTL
Côte d'Ivoire ABIDJAN
Croatia ZAGREB MAKSIMIR ZAGREB PLESO
Cuba CAMAGUEY INTL HAVANA JOSE MARTI INTL SANTIAGO ANTONIO MACEO INTL
Licensed for single user. © 2017 ASHRAE, Inc.
Czech Republic BRNO-TURANY OSTRAVA MOSNOV PRAHA-KBELY PRAHA-LIBUS PRAHA-RUZYNE
Denmark DROGDEN FYR KOEBENHAVNS KASTRUP ROSKILDE VAERLOSE
Dominican Republic LAS AMERICAS INTL SANTO DOMINGO
Ecuador JOSE JOAQUIN DE OLMEDO INTL QUITO PARQUE BICENTENARIO
2.157S 79.884W 6 0.141S 78.488W 2813
Egypt ALEXANDRIA INTL ASSIUT CAIRO INTL LUXOR INTL PORT SAID PORT SAID EL GAMIL
31.184N 27.047N 30.122N 25.671N 31.267N 31.283N
29.949E 31.012E 31.406E 32.707E 32.300E 32.233E
-2 235 116 90 6 6
7.0 4.7 7.9 5.8 9.8 9.6
8.0 5.8 8.9 6.9 10.8 10.7
33.2 41.1 38.2 43.2 32.2 31.9
22.4 20.6 21.2 22.8 25.4 25.4
31.9 39.8 36.9 42.2 31.2 31.1
23.3 20.5 21.5 22.6 25.4 25.3
30.9 38.4 35.7 41.2 30.9 30.4
23.7 20.2 21.6 22.4 25.4 25.0
25.7 23.1 25.2 24.4 27.0 26.9
30.2 35.1 32.1 40.1 30.5 30.3
25.1 22.2 24.5 23.8 26.5 26.3
29.6 34.6 31.3 39.5 30.0 29.8
24.2 19.8 23.2 19.1 26.1 25.9
19.1 14.9 18.3 14.1 21.5 21.2
28.6 27.8 27.7 33.3 29.5 29.4
23.8 18.4 22.8 18.2 25.2 25.2
18.6 13.7 17.8 13.2 20.4 20.3
28.4 27.6 27.4 33.5 29.1 29.1
59.413N 24.833E
40
-18.5
-15.1
26.7
19.1
24.8
18.0
23.0
17.1
20.2
24.6
19.0
23.0
18.7
13.6
22.2
17.5
12.6
21.0
60.317N 24.963E 60.100N 25.067E
55 7
-21.2 -17.4
-18.0 -14.0
27.1 22.7
18.1 18.9
25.2 21.5
17.2 18.3
23.8 20.4
16.4 17.5
19.6 19.8
24.6 21.9
18.6 18.8
23.3 21.0
18.0 18.9
13.0 13.7
21.5 21.4
16.8 17.9
12.1 12.9
20.4 20.3
Estonia TALLINN
Finland HELSINKI VANTAA ISOSAARI
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 5 sites, 1 more on CD-ROM 6.7 5.9 5.3 0 2196 7.9 6.8 6.0 1666 0 10.9 9.6 8.5 0 3673 8.0 6.4 5.5 392 24 8.3 7.2 6.2 0 3631 1 site, 1 more on CD-ROM 5.9 5.0 4.4 0 2847 1 site, 0 more on CD-ROM 10.4 9.5 8.6 0 1836 1 site, 2 more on CD-ROM 7.1 6.4 5.9 0 3243 2 sites, 14 more on CD-ROM 5.6 4.7 4.1 2712 352 8.4 7.1 5.9 2826 317 3 sites, 6 more on CD-ROM 9.4 8.4 7.4 4 2678 9.4 8.3 7.4 30 2303 8.2 7.1 6.2 0 2865 5 sites, 35 more on CD-ROM 10.1 8.8 7.8 3371 192 10.1 8.9 8.0 3551 127 9.1 7.9 6.8 3407 143 7.1 6.1 5.3 3378 155 11.8 10.0 8.6 3633 104 4 sites, 45 more on CD-ROM 17.6 15.3 14.0 3406 30 12.5 11.2 10.1 3513 54 12.4 11.0 9.8 3712 33 11.8 10.3 9.1 3764 40 2 sites, 4 more on CD-ROM 7.5 6.5 5.9 0 2896 6.4 5.5 4.5 0 3057 2 sites, 2 more on CD-ROM 7.2 6.4 5.8 0 2777 7.7 6.8 6.0 1408 0 6 sites, 22 more on CD-ROM 9.8 8.6 7.7 452 1352 10.1 9.0 8.2 474 2114 9.3 8.1 7.2 344 1887 7.0 6.1 5.3 264 2843 10.4 9.4 8.5 276 1619 11.4 10.1 9.2 308 1528 1 site, 22 more on CD-ROM 9.2 8.1 7.2 4559 41 2 sites, 70 more on CD-ROM 10.3 9.1 8.1 4699 53 15.8 14.1 12.5 4485 31
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB 99.6%
99%
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
France CAP COURONNE CAP FERRAT CAP POMEGUES LYON-BRON LYON-SAINT EXUPERY MARSEILLE PROVENCE NICE COTE D'AZUR PARIS CHARLES DE GAULLE PARIS LE BOURGET PARIS ORLY PARIS-MONTSOURIS TOULOUSE BLAGNAC TRAPPES VELIZY-VILLACOUBLAY
43.333N 43.683N 43.267N 45.727N 45.726N 43.436N 43.658N 49.013N 48.969N 48.725N 48.817N 43.629N 48.767N 48.774N
5.050E 7.333E 5.300E 4.944E 5.091E 5.214E 7.216E 2.550E 2.441E 2.359E 2.333E 1.364E 2.000E 2.202E
27 144 70 201 250 23 4 120 66 89 77 152 168 178
-2.9 4.0 -1.4 -5.6 -6.0 -2.5 1.9 -5.1 -4.7 -5.0 -3.1 -3.8 -4.9 -4.9
0.4 5.2 1.8 -3.9 -4.2 -1.1 3.1 -3.4 -3.0 -3.2 -1.8 -2.2 -3.3 -3.2
30.7 29.1 28.6 33.1 32.6 32.8 29.5 30.8 31.1 31.1 31.2 33.2 30.2 30.0
22.6 22.4 22.0 19.8 19.9 21.1 22.6 19.9 19.8 20.0 20.1 20.7 19.6 19.5
29.5 28.1 27.3 31.2 30.7 31.5 28.5 28.7 28.9 29.0 29.1 31.2 28.0 28.0
22.3 22.4 21.9 19.6 19.7 20.8 22.4 19.1 19.1 19.4 19.5 20.2 18.9 18.9
28.3 27.1 26.3 29.5 28.9 30.2 27.7 26.7 26.9 27.1 27.1 29.6 26.1 26.1
21.8 22.1 21.5 19.2 19.2 20.5 22.2 18.4 18.4 18.6 18.5 19.7 18.1 18.1
24.8 24.6 24.2 21.3 21.4 23.4 24.7 21.1 21.1 21.2 21.3 22.3 20.7 20.8
28.7 27.3 26.5 29.3 28.9 28.8 27.7 28.0 28.4 28.4 28.9 29.5 27.5 27.5
24.0 23.9 23.4 20.6 20.6 22.6 23.9 20.2 20.1 20.3 20.2 21.5 19.7 19.8
27.8 26.7 25.7 28.4 27.9 28.3 27.2 26.4 26.6 26.8 26.9 28.4 25.8 25.8
23.6 23.7 23.5 18.8 19.0 21.8 23.8 18.9 18.8 19.0 18.8 20.1 18.6 18.6
18.5 18.9 18.4 14.0 14.2 16.5 18.6 13.9 13.7 14.0 13.7 15.0 13.7 13.8
27.4 26.5 25.6 24.0 24.3 26.4 27.0 23.7 23.3 23.7 24.1 25.3 23.0 23.3
22.8 23.0 22.7 18.1 18.2 20.8 22.8 18.0 17.8 18.1 17.8 19.2 17.6 17.7
17.6 18.1 17.5 13.3 13.5 15.5 17.5 13.1 12.9 13.1 12.9 14.2 12.9 13.0
26.6 25.9 25.0 23.4 23.3 25.8 26.5 22.7 22.3 22.8 23.0 24.5 21.8 22.3
0.459N
9.412E
12
21.9
22.4
31.8
27.4
31.1
27.2
30.9
27.1
28.3
30.3
27.9
29.8
27.9
24.1
29.9
27.2
23.0
29.2
13.338N 16.652W
29
16.6
17.3
37.8
20.3
36.1
20.2
35.0
21.0
27.7
31.5
27.2
31.0
26.8
22.5
30.1
26.2
21.7
29.5
41.667N 44.950E
495
-6.0
-4.1
35.0
21.7
33.5
21.4
32.0
21.2
23.9
31.4
22.8
30.5
21.7
17.4
28.5
20.3
15.9
27.1
52.467N 52.380N 52.560N 52.473N 53.048N 52.591N 51.133N 51.289N 51.400N 50.026N 48.206N 51.923N 53.630N 52.461N 49.393N 53.983N 50.866N 51.424N 51.317N 48.354N 50.831N 49.499N 52.383N 53.733N 49.218N 48.690N 48.833N 52.457N
13.300E 13.523E 13.288E 13.404E 8.787E 10.022E 13.767E 6.767E 6.967E 8.543E 11.267E 8.306E 9.988E 9.685E 8.652E 9.567E 7.143E 12.236E 12.450E 11.786E 6.658E 11.078E 13.067E 9.883E 11.100E 9.222E 9.200E 9.427E
51 48 37 51 4 39 230 45 161 111 519 72 16 56 109 26 92 142 151 453 118 319 100 17 387 389 315 57
-12.0 -13.5 -12.4 -11.8 -9.4 -10.4 -13.3 -9.9 -9.9 -8.8 -15.1 -8.8 -11.6 -12.7 -8.2 -9.6 -8.3 -13.3 -11.0 -13.0 -7.7 -14.4 -12.6 -9.7 -13.1 -12.7 -11.5 -9.5
-9.1 -10.5 -9.2 -9.2 -7.3 -7.9 -10.3 -6.8 -6.9 -6.5 -12.1 -6.1 -8.9 -9.7 -5.8 -7.3 -6.2 -10.4 -8.2 -10.0 -5.3 -10.9 -10.0 -7.4 -10.1 -10.0 -9.0 -7.1
29.3 29.9 30.0 30.0 28.9 30.2 29.8 29.6 28.2 31.4 29.1 30.2 27.8 28.9 32.1 28.4 30.3 29.8 30.3 29.9 30.7 30.2 29.7 28.3 31.0 29.3 29.6 30.2
19.0 19.1 18.7 18.9 19.5 19.0 18.8 19.6 19.3 19.1 18.9 19.2 18.9 19.4 20.4 18.7 19.6 19.2 19.3 19.1 19.5 18.5 18.9 18.9 19.3 18.9 19.6 19.2
27.3 27.9 28.0 28.0 26.9 28.2 27.7 27.8 26.6 29.3 27.1 28.0 25.9 27.0 30.2 26.1 28.2 27.7 28.3 27.9 28.4 28.3 27.7 26.2 28.9 27.6 27.8 28.1
18.2 18.2 18.0 18.2 18.7 18.3 18.2 18.7 18.4 18.7 18.1 18.6 18.1 18.5 19.7 18.3 18.8 18.5 18.4 18.6 18.9 17.9 18.3 18.5 18.5 18.4 18.6 18.6
25.6 26.1 26.2 26.2 25.0 26.3 25.9 26.1 24.8 27.6 25.2 26.0 24.0 25.1 28.2 24.1 26.3 25.9 26.5 26.1 26.4 26.5 25.9 24.3 27.0 25.8 26.1 26.1
17.4 17.5 17.1 17.5 17.8 17.5 17.5 17.9 17.6 18.1 17.2 17.9 17.2 17.6 18.7 17.7 18.0 17.7 17.9 17.9 18.0 17.0 17.7 17.7 17.8 17.7 17.9 17.7
20.2 20.3 20.0 20.1 20.6 20.2 20.1 20.5 20.1 20.8 19.7 20.6 20.0 20.4 21.6 20.3 20.8 20.2 20.5 20.1 20.7 19.7 20.3 20.4 20.2 20.0 20.6 20.4
26.5 26.9 26.9 26.9 26.5 27.7 27.2 27.3 26.2 27.8 27.2 27.3 25.8 26.5 29.0 25.7 27.7 27.1 27.3 26.9 27.9 26.8 27.0 25.6 28.2 27.4 27.3 27.7
19.3 19.3 19.1 19.2 19.5 19.3 19.1 19.7 19.2 20.0 18.8 19.7 18.9 19.4 20.7 19.3 19.7 19.3 19.6 19.4 19.7 18.9 19.4 19.4 19.4 19.2 19.7 19.4
25.2 25.8 25.7 25.8 24.8 26.2 25.5 26.0 25.1 26.3 25.8 25.7 24.2 25.2 27.9 24.2 25.9 25.7 25.7 25.8 26.4 25.5 25.6 24.1 26.8 26.0 25.9 26.1
18.1 18.2 17.9 18.0 18.6 17.8 17.6 18.3 18.0 18.9 17.0 18.5 17.9 18.3 19.2 18.5 18.6 17.8 18.3 18.0 18.2 17.6 18.2 18.7 17.8 17.4 18.2 18.0
13.1 13.1 12.9 13.0 13.4 12.9 13.0 13.2 13.2 13.9 12.9 13.5 12.9 13.3 14.1 13.4 13.6 13.0 13.4 13.7 13.3 13.1 13.3 13.5 13.4 13.0 13.6 13.0
22.3 22.3 22.1 22.1 22.9 21.9 22.0 22.9 22.4 22.3 23.3 22.4 22.1 22.2 23.9 22.4 23.3 22.7 22.2 22.0 23.0 21.3 21.8 21.9 21.9 23.1 23.2 22.4
17.1 17.1 16.9 16.9 17.6 16.9 16.8 17.4 17.1 17.9 16.1 17.7 16.9 17.3 18.2 17.5 17.6 17.0 17.5 17.1 17.4 16.8 17.1 17.7 17.0 16.6 17.4 17.0
12.3 12.3 12.1 12.1 12.6 12.1 12.3 12.5 12.4 13.0 12.2 12.8 12.1 12.5 13.2 12.5 12.8 12.3 12.7 12.9 12.6 12.4 12.4 12.7 12.7 12.4 12.9 12.2
21.1 21.3 21.0 21.2 21.5 21.8 21.5 21.9 21.2 22.0 21.8 21.6 21.2 21.6 23.3 21.3 22.2 21.8 21.6 21.4 22.1 20.6 21.2 20.9 21.5 22.2 22.4 22.0
37.882N 23.735E 38.064N 23.556E 40.520N 22.971E
21 44 7
1.8 0.8 -3.0
3.3 2.1 -1.3
35.8 36.9 34.8
21.1 20.6 21.7
34.2 35.2 33.1
21.0 20.4 21.7
33.0 34.1 31.9
21.1 20.1 21.2
24.6 22.6 23.8
31.5 31.7 31.0
23.9 22.0 23.0
30.8 31.4 30.2
22.3 20.0 21.2
17.0 14.7 15.9
29.0 26.6 28.3
21.6 19.0 20.3
16.3 13.8 15.0
28.5 26.1 27.4
10.9
11.9
28.1
17.8
27.1
17.9
26.2
17.8
20.2
24.8
19.7
24.1
19.0
16.7
21.3
18.2
15.8
20.6
Gabon LIBREVILLE INTL
Gambi BANJUL INTL
Georgia TBILISI
Licensed for single user. © 2017 ASHRAE, Inc.
Germany BERLIN DAHLEM BERLIN SCHONEFELD BERLIN TEGEL BERLIN TEMPELHOF BREMEN CELLE DRESDEN DUSSELDORF ESSEN MULHEIM FRANKFURT AM MAIN FURSTENFELDBRUCK GUTERSLOH HAMBURG FUHLSBUTTEL HANNOVER HEIDELBERG ITZEHOE KOLN BONN LEIPZIG HALLE LEIPZIG-HOLZHAUSEN MUNICH NORVENICH NURNBERG POTSDAM QUICKBORN ROTH STUTTGART FILDERSTADT STUTTGART SCHNARREN WUNSTORF
Greece ATHINAI HELLINIKON ELEFSIS THESSALONIKI MAKEDONIA
Guatemala LA AURORA INTL
14.583N 90.527W 1509
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 14 sites, 163 more on CD-ROM 17.1 15.0 13.3 1578 567 12.2 9.9 8.0 1270 560 23.5 20.7 17.9 1523 457 11.2 9.6 8.4 2372 368 10.6 9.2 7.9 2464 332 16.0 13.9 12.1 1616 644 12.0 10.2 8.4 1383 547 11.3 9.8 8.6 2555 172 9.6 8.5 7.5 2491 168 10.6 9.2 8.1 2556 183 7.3 6.4 5.7 2293 235 10.7 9.4 8.3 2021 399 6.8 6.0 5.3 2668 138 9.3 8.2 7.3 2700 155 1 site, 0 more on CD-ROM 6.8 6.1 5.5 0 3014 1 site, 0 more on CD-ROM 8.6 7.8 6.9 0 3141 1 site, 7 more on CD-ROM 19.1 15.8 13.6 2286 713 28 sites, 117 more on CD-ROM 7.4 6.5 5.9 3390 118 11.1 9.6 8.4 3485 112 10.4 9.2 8.2 3317 147 10.4 9.1 8.1 3283 147 11.5 10.1 8.9 3360 87 9.1 7.9 6.9 3226 127 9.6 8.3 7.3 3386 135 10.4 9.2 8.1 2928 139 9.7 8.4 7.4 3178 103 10.1 8.8 7.6 3026 194 11.1 9.3 7.7 3706 82 9.9 8.5 7.4 3068 122 10.2 9.0 8.1 3513 61 10.2 8.9 8.0 3368 80 8.3 7.1 6.2 2721 276 9.4 8.1 7.1 3486 62 8.9 7.8 6.9 3053 117 12.5 10.8 9.4 3393 120 6.5 5.7 5.0 3169 153 11.2 9.4 8.1 3492 108 9.8 8.2 7.1 2893 131 9.2 7.9 6.8 3506 128 10.7 9.4 8.3 3437 119 8.8 7.5 6.6 3461 56 7.8 6.5 5.5 3541 115 9.4 7.9 6.8 3489 106 9.1 7.7 6.6 3152 160 10.3 9.1 8.1 3119 127 3 sites, 26 more on CD-ROM 9.9 8.8 7.9 1098 1152 10.0 8.9 8.1 1217 1185 11.5 9.5 8.1 1763 854 1 site, 3 more on CD-ROM 11.7 10.2 9.3 62 706
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
99.6%
99%
15.453N 87.924W 28 14.061N 87.217W 1004
17.8 11.8
18.8 13.0
36.9 32.0
25.7 19.3
35.8 31.0
26.0 19.6
34.9 30.1
26.0 19.6
28.4 22.6
32.7 27.7
28.0 22.2
31.9 27.3
27.2 21.2
23.1 17.9
30.2 24.2
27.0 20.9
22.8 17.6
30.0 23.9
22.309N 113.915E 22.300N 114.167E
9 62
8.8 9.6
10.1 10.9
34.0 32.2
26.5 26.5
33.2 31.7
26.3 26.4
32.8 31.2
26.2 26.3
27.7 27.4
31.1 30.5
27.3 27.1
30.8 30.1
26.9 26.6
22.5 22.3
30.2 29.3
26.2 26.2
21.6 21.8
29.8 29.1
47.450N 18.967E 47.437N 19.256E 47.433N 19.183E
132 151 139
-11.2 -11.9 -9.9
-9.0 -9.1 -7.8
31.0 33.0 33.1
20.1 22.0 20.4
29.3 31.1 31.3
19.7 20.8 19.8
27.8 29.2 29.6
19.2 19.9 19.2
21.3 23.0 21.7
28.9 30.6 29.6
20.5 21.9 20.8
27.5 28.8 28.9
18.6 20.2 19.2
13.7 15.2 14.2
24.3 26.5 24.4
17.9 19.8 18.2
13.1 14.8 13.4
23.6 25.6 23.5
23.077N 20.700N 19.863N 12.967N 15.859N 23.287N 20.244N 28.000N 12.994N 11.031N 17.700N 22.722N 26.106N 26.293N 17.452N 23.178N 26.824N 22.813N 26.251N 22.655N 11.250N 26.761N 12.961N 19.089N 21.092N 14.450N 28.567N 28.585N 30.333N 25.591N 18.583N 22.309N 17.628N 21.200N 8.483N 10.765N
72.635E 77.033E 75.398E 77.583E 74.618E 77.337E 85.818E 73.300E 80.181E 77.044E 83.300E 75.801E 91.586E 78.228E 78.461E 80.052E 75.812E 86.169E 73.049E 88.447E 75.783E 80.889E 74.890E 72.868E 79.047E 79.983E 77.103E 77.206E 76.467E 85.088E 73.917E 70.780E 75.935E 72.833E 76.950E 78.710E
58 282 583 921 758 524 42 224 16 404 66 564 49 188 531 495 385 154 219 5 5 125 103 11 315 20 237 215 251 52 592 134 483 12 64 88
10.8 12.2 10.8 15.3 13.0 9.5 13.9 6.1 19.9 18.3 20.0 9.4 10.8 5.9 13.6 8.3 7.1 10.1 8.8 11.3 22.6 6.7 20.8 16.8 11.4 20.5 5.8 6.0 5.0 7.8 9.8 11.8 15.4 14.1 22.1 20.0
12.1 13.6 12.1 16.0 14.2 10.8 15.0 7.5 20.8 19.2 20.8 10.8 11.8 7.0 14.9 9.6 8.5 11.3 9.9 12.4 23.1 7.9 21.5 18.1 12.8 21.2 6.9 7.0 6.1 9.0 10.9 13.2 16.9 15.4 22.8 20.8
42.5 43.2 40.1 34.2 36.3 41.8 39.1 44.2 38.9 36.3 33.8 40.7 34.8 43.6 40.6 42.5 42.4 42.2 42.6 37.8 34.4 42.2 34.4 35.9 44.2 40.7 43.2 42.3 41.8 41.2 38.2 41.1 41.1 37.9 34.0 39.1
23.0 21.9 22.8 19.9 19.2 21.9 27.1 21.5 25.9 22.0 27.2 20.0 27.1 23.2 21.7 20.7 21.3 22.5 21.3 27.2 28.0 23.1 24.9 22.7 22.6 27.0 22.4 23.0 25.1 23.1 19.9 22.4 22.6 22.5 25.9 26.0
41.2 42.1 39.2 33.5 35.3 40.6 37.7 42.9 37.4 35.5 33.1 39.6 34.0 42.4 39.2 41.3 41.2 40.7 41.3 36.7 33.9 40.9 33.8 34.9 43.0 39.3 42.0 40.8 40.3 39.8 37.2 40.0 40.0 36.5 33.4 38.2
22.9 21.6 22.7 19.8 19.3 21.6 26.9 21.9 25.8 22.3 27.6 20.0 27.0 23.2 21.7 20.7 21.4 22.8 21.7 27.2 27.7 23.0 24.9 23.0 22.6 27.1 22.3 23.2 24.9 23.2 19.8 22.2 22.8 22.9 25.8 25.8
40.0 41.0 38.2 32.6 34.4 39.4 36.5 41.6 36.2 34.7 32.6 38.4 33.2 41.1 38.1 40.1 40.0 39.1 40.1 35.7 33.4 39.2 33.3 34.0 41.9 38.1 40.8 39.5 38.7 38.1 36.2 38.9 39.0 35.2 32.9 37.7
22.9 21.6 22.4 19.8 19.4 21.5 26.7 22.4 25.8 22.6 27.5 19.9 26.9 23.1 21.7 20.9 21.4 23.2 21.9 27.0 27.4 23.7 24.8 23.3 22.3 26.9 22.4 23.4 24.9 23.9 19.8 22.7 22.5 23.1 25.7 25.7
28.7 27.0 26.7 23.6 24.2 26.3 29.6 28.1 28.4 25.7 29.2 25.6 29.1 28.6 25.7 26.6 27.7 28.2 27.5 29.6 28.9 29.4 27.2 27.8 27.7 29.0 29.4 28.7 29.6 29.0 24.7 27.9 26.7 28.2 27.7 27.7
33.9 33.0 35.2 29.0 29.4 31.5 34.5 34.4 33.0 31.3 32.2 30.4 32.8 33.3 31.6 31.1 30.8 33.4 32.3 34.6 33.2 34.3 31.3 31.1 32.5 35.9 33.2 34.2 33.9 33.9 29.9 33.2 33.5 31.8 31.7 35.0
28.1 26.5 25.7 23.1 23.7 25.9 29.1 27.6 28.0 25.2 28.8 25.1 28.7 28.1 25.1 26.2 27.2 27.7 27.1 29.2 28.5 28.9 26.9 27.5 27.1 28.5 28.7 28.4 29.2 28.6 24.2 27.4 25.9 27.9 27.3 27.2
32.7 31.9 33.1 28.4 28.5 30.7 33.8 33.9 32.3 30.6 31.8 29.6 32.2 32.7 30.9 30.3 30.6 32.5 32.0 34.0 32.8 33.7 30.9 30.7 31.9 35.0 32.5 33.7 33.8 33.2 29.1 32.1 32.4 31.5 31.3 34.2
27.6 25.7 24.7 22.2 22.9 25.2 28.5 26.7 27.2 24.5 28.5 24.5 28.2 27.5 24.2 25.6 27.1 27.1 26.5 28.5 27.7 28.2 26.2 27.1 26.9 27.5 28.9 27.6 28.6 27.9 23.5 26.8 25.1 27.4 26.6 26.2
23.7 21.7 21.2 19.0 19.3 21.7 25.0 22.9 23.1 20.5 25.0 20.8 24.5 23.9 20.4 22.2 24.0 23.2 22.6 24.8 23.6 24.8 21.8 22.8 23.4 23.5 26.2 24.2 25.8 24.1 19.6 22.9 21.5 23.3 22.3 21.8
30.8 28.9 29.8 25.4 26.0 28.1 31.6 30.5 30.4 27.2 31.7 27.2 31.1 30.8 27.7 28.3 28.8 30.3 29.6 32.4 32.2 31.7 29.0 29.8 29.8 31.5 30.9 31.3 31.8 31.1 26.2 29.3 29.4 30.2 30.0 30.0
27.1 25.2 24.1 21.7 22.5 24.8 28.1 26.2 27.0 24.1 28.0 24.1 27.7 27.1 23.8 25.2 26.5 26.6 26.0 28.1 27.2 28.0 25.8 26.6 26.1 27.1 28.0 27.2 28.2 27.5 23.1 26.5 24.5 27.0 26.2 25.8
23.0 21.1 20.4 18.4 18.9 21.1 24.4 22.2 22.7 20.0 24.3 20.4 23.9 23.3 20.0 21.6 23.1 22.6 21.9 24.3 23.0 24.4 21.4 22.2 22.3 22.8 24.8 23.6 25.2 23.5 19.2 22.3 20.7 22.8 21.8 21.3
30.2 28.3 28.5 25.0 25.4 27.6 31.3 30.3 30.2 26.9 31.4 26.8 30.7 30.3 27.2 27.9 28.6 29.7 29.2 32.0 31.6 31.4 28.7 29.5 29.2 31.2 30.6 31.0 31.6 30.6 25.8 29.0 28.6 29.9 29.6 29.7
8.748S 5.062S 7.380S 3.558N 0.875S 1.549N 6.126S 0.461N
115.167E 119.554E 112.787E 98.672E 100.352E 124.926E 106.656E 101.445E
4 14 3 35 3 81 10 31
21.9 20.2 21.0 22.5 21.4 20.6 22.2 22.0
22.8 21.1 21.8 22.8 22.0 21.4 22.8 22.4
32.5 34.3 34.1 34.2 32.2 33.1 34.0 34.6
26.6 24.1 24.4 26.0 25.9 24.4 25.5 26.4
32.1 33.8 33.5 33.8 31.9 32.6 33.2 34.1
26.5 24.3 24.5 26.0 25.9 24.5 25.6 26.4
31.7 33.1 33.0 33.1 31.5 32.1 33.0 33.7
26.4 24.7 24.7 26.0 25.8 24.7 25.7 26.3
27.7 27.7 27.0 27.5 27.2 26.6 27.7 28.0
30.7 30.8 31.0 31.8 30.9 30.6 31.3 32.8
27.2 27.3 26.7 27.1 26.8 26.2 27.4 27.6
30.4 30.6 30.6 31.4 30.5 30.2 31.0 32.3
26.9 27.1 26.0 26.2 26.1 25.3 26.9 26.6
22.6 22.9 21.4 21.7 21.5 20.6 22.6 22.3
29.9 28.9 28.8 29.6 29.5 28.5 30.1 31.2
26.2 26.6 25.6 26.0 25.7 25.1 26.2 26.2
21.7 22.2 20.9 21.5 21.0 20.4 21.7 21.7
29.2 28.6 28.6 29.4 29.2 28.3 29.3 30.7
Honduras RAMON VILLEDA MORALES INTL TONCONTIN INTL
Hong Kong HONG KONG INTL HONG KONG OBSERVATORY
Hungary BUDAORS BUDAPEST FERIHEGY BUDAPEST PESTSZENTLORINC
Licensed for single user. © 2017 ASHRAE, Inc.
India AHMEDABAD AKOLA AURANGABAD BANGALURU BELGAUM BHOPAL BHUBANESHWAR BIKANER CHENNAI INTL COIMBATORE INTL CWC VISHAKHAPATNAM WALTAIR DEVI AHILYA BAI HOLKAR INTL GUWAHATI INTL GWALIOR HYDERABAD BEGUMPET AP JABALPUR JAIPUR JAMSHEDPUR JODHPUR KOLKATA BOSE INTL KOZHIKODE LUCKNOW MANGALURU INTL MUMBAI SHIVAJI INTL NAGPUR AMBEDKAR INTL NELLORE NEW DELHI INDIRA GANDHI INTL NEW DELHI SAFDARJUNG PATIALA PATNA PUNE INTL RAJKOT SOLAPUR SURAT THIRUVANANTHAPURAM TIRUCHIRAPPALLI
Indonesia BALI NGURAH RAI INTL HASANUDDIN INTL JUANDA INTL MEDAN POLONIA INTL MINANGKABAU INTL SAM RATULANGI INTL SOEKARNO HATTA INTL SYARIF KASIM II INTL
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 2 sites, 1 more on CD-ROM 8.3 7.3 6.4 0 3206 9.1 8.0 7.0 11 1510 2 sites, 4 more on CD-ROM 10.5 9.3 8.3 180 2346 8.6 7.4 6.5 237 1976 3 sites, 33 more on CD-ROM 13.9 11.6 9.3 3072 246 12.5 10.3 8.6 3105 286 7.4 6.4 5.5 2896 372 36 sites, 46 more on CD-ROM 6.7 6.0 5.3 11 3497 4.3 3.4 2.9 1 3572 7.4 6.2 5.4 6 2802 5.4 4.5 3.9 0 2160 8.1 7.2 6.2 0 2224 8.8 7.9 6.7 61 2740 9.8 8.5 7.7 1 3380 6.6 5.2 4.1 177 3464 8.1 7.2 6.3 0 3830 8.9 8.1 7.3 0 3139 7.8 6.7 5.7 0 3448 10.1 9.1 8.4 44 2639 4.9 4.1 3.4 57 2428 4.3 3.4 2.9 192 3025 7.4 6.3 5.7 1 3103 4.1 3.4 3.0 84 2848 7.6 6.3 5.4 166 2983 4.3 3.4 3.0 24 3094 5.0 3.8 3.3 76 3379 6.3 5.4 4.8 20 3118 5.6 4.6 3.6 0 3583 6.7 5.8 5.0 193 2808 7.9 6.5 5.9 0 3337 7.2 6.4 5.9 0 3476 7.6 6.2 5.3 6 3318 5.0 3.9 3.4 0 4103 7.9 6.7 5.9 292 2971 7.0 5.9 5.0 267 2823 3.9 3.2 2.5 401 2452 6.4 5.7 5.1 145 2892 5.1 4.0 3.3 7 2375 10.1 8.8 8.0 6 3451 3.1 2.5 2.3 0 3514 5.4 4.6 3.8 1 3425 5.2 4.3 3.5 0 3422 11.1 9.7 8.5 0 4049 8 sites, 50 more on CD-ROM 8.7 7.7 6.7 0 3382 6.6 5.8 5.0 0 3304 8.4 7.4 6.3 0 3493 6.2 5.4 4.9 0 3454 6.1 5.2 4.4 0 3167 7.5 6.0 4.9 0 3089 9.1 7.9 6.9 0 3469 5.3 4.5 4.0 0 3520
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
99.6%
99%
3 20 1662 7 -26 1749 1546 1056 1748 995 1208 1306 1500 1359 1324 1378 1640
4.0 4.8 -15.8 9.1 1.2 -18.0 -8.1 -5.0 -7.1 -9.2 -3.6 -7.9 -2.2 -11.8 -12.0 -5.1 -14.3
5.5 6.0 -11.6 10.8 2.6 -14.3 -6.1 -2.5 -5.2 -6.1 -1.9 -5.6 -1.0 -9.2 -9.5 -3.2 -11.4
47.9 47.9 36.5 41.9 30.8 35.6 39.1 42.0 38.1 37.2 38.8 39.9 39.2 35.9 33.2 39.1 34.3
22.5 23.1 16.1 23.9 25.4 16.8 17.1 19.6 16.1 17.8 18.1 17.7 17.8 16.8 17.7 16.4 16.0
46.9 46.9 35.3 40.1 30.1 34.5 38.1 40.8 37.0 36.1 37.4 38.8 38.2 34.2 32.0 38.0 33.0
22.3 22.9 15.8 25.3 25.1 16.2 16.6 19.2 15.7 17.6 17.9 17.4 17.4 16.4 17.7 16.1 16.0
45.9 45.8 34.1 38.9 29.4 33.3 36.9 39.6 36.0 34.9 36.2 37.5 37.2 33.0 30.8 37.0 31.6
22.0 22.6 15.4 25.8 24.8 15.8 16.4 18.8 15.4 17.2 17.6 16.7 17.0 16.1 17.3 15.4 15.6
28.7 28.2 18.6 31.1 26.8 18.9 18.7 21.7 17.7 21.4 21.9 19.9 20.3 18.4 19.7 18.7 18.5
35.1 35.9 32.2 35.3 29.7 32.5 36.1 38.4 34.5 33.2 32.6 37.0 34.9 31.5 29.7 33.9 30.4
27.4 26.7 17.4 30.7 26.2 17.8 17.9 20.7 16.9 20.2 20.4 18.8 19.3 17.8 19.0 17.4 17.6
35.2 37.5 31.7 34.9 29.1 31.4 35.4 37.6 34.2 32.4 33.3 35.9 34.6 30.5 29.1 34.9 29.3
27.2 26.2 13.5 30.1 26.0 13.7 11.9 15.7 11.0 17.2 18.0 13.1 15.1 14.1 16.2 12.8 14.0
23.0 21.7 11.9 27.4 21.3 12.1 10.5 12.7 10.1 13.8 15.0 11.0 12.9 11.8 13.6 10.9 12.2
32.5 32.9 26.4 33.8 29.1 26.1 27.2 32.0 23.1 28.8 30.3 25.8 30.2 23.0 24.9 25.2 23.6
25.2 24.0 11.7 29.8 25.3 12.2 10.2 14.1 9.7 15.7 15.8 11.8 13.2 13.1 15.2 10.5 13.2
20.4 18.9 10.5 26.8 20.4 11.0 9.4 11.4 9.2 12.5 13.0 10.1 11.4 11.1 12.7 9.3 11.6
32.1 33.1 24.2 33.7 28.6 25.5 24.5 31.3 22.2 27.1 28.3 24.7 28.4 23.2 24.6 20.7 22.9
53.302N 6.451W 53.421N 6.270W
97 74
-3.1 -2.7
-1.6 -1.2
22.7 21.9
17.4 17.0
21.1 20.5
16.6 16.3
19.9 19.3
15.9 15.6
18.2 17.8
21.2 20.6
17.4 17.0
20.1 19.5
17.1 16.6
12.3 12.0
19.6 19.1
16.2 15.9
11.6 11.4
18.7 18.3
32.011N 34.887E 32.117N 34.783E
41 13
5.8 7.3
7.0 8.7
35.1 31.4
20.6 23.4
33.2 30.4
22.1 24.2
32.2 29.9
22.8 24.2
25.9 26.7
30.9 29.5
25.2 26.0
30.3 29.1
24.2 25.8
19.2 21.2
29.2 29.0
23.7 25.0
18.6 20.1
28.8 28.6
41.133N 44.535N 37.467N 37.400N 43.810N 44.412N 41.061N 45.445N 40.884N 38.176N 41.659N 41.799N 41.804N 45.033N 45.201N 45.650N
54 38 12 31 44 4 9 108 72 20 13 130 5 710 301 20
0.9 -4.2 1.5 1.0 -3.0 1.2 -0.9 -4.8 1.1 6.8 0.8 -1.1 -0.2 -5.0 -5.8 -1.8
2.0 -2.9 2.9 2.1 -1.2 2.9 0.2 -3.2 2.3 7.8 2.0 0.1 0.9 -3.2 -4.1 0.0
34.0 34.4 34.4 37.1 35.2 29.9 32.9 33.2 33.0 33.1 31.0 33.9 31.2 28.2 30.9 31.7
22.4 22.4 23.2 22.2 21.8 23.0 23.7 23.7 23.1 22.4 23.1 21.7 21.9 20.6 22.3 23.3
32.1 33.0 32.8 35.4 33.9 28.9 31.7 32.0 31.8 31.1 30.0 32.6 30.1 27.0 29.8 30.3
22.2 22.2 23.1 22.1 21.7 23.4 23.5 23.0 23.0 23.0 23.3 21.6 22.3 20.2 21.7 23.1
30.9 31.8 31.2 33.9 32.2 28.0 30.2 30.8 30.8 29.9 29.0 31.1 29.1 25.9 28.4 29.3
21.8 21.8 22.8 22.0 21.1 23.2 23.3 22.4 23.0 23.6 23.5 21.2 22.2 19.6 20.9 22.5
25.2 24.5 26.2 24.9 24.1 26.0 26.9 25.1 26.1 26.4 26.2 24.7 25.3 23.0 23.7 25.3
29.8 30.8 30.0 31.2 31.1 27.8 30.0 30.9 29.9 29.0 28.6 28.6 28.3 25.7 28.6 29.9
24.2 23.6 25.3 24.2 23.2 25.2 25.9 24.2 25.2 25.7 25.4 23.9 24.5 22.1 22.9 24.3
29.0 29.9 29.5 30.7 30.1 27.3 29.1 29.7 29.2 28.5 28.1 28.2 27.7 24.9 27.7 29.1
24.0 22.7 25.2 23.2 22.1 25.3 26.1 23.3 25.0 25.8 25.2 23.8 24.2 22.2 22.2 23.8
19.0 17.5 20.3 18.0 16.9 20.5 21.5 18.3 20.3 21.2 20.4 19.0 19.1 18.4 17.5 18.6
27.7 28.1 28.1 27.4 27.0 27.3 29.1 28.1 28.5 28.3 28.2 26.7 27.0 24.3 25.9 29.3
22.8 21.8 24.1 22.2 21.1 24.3 25.0 22.4 24.0 24.9 24.2 22.8 23.5 21.1 21.2 22.6
17.7 16.5 19.1 16.9 15.8 19.3 20.0 17.3 19.0 20.0 19.2 17.9 18.3 17.2 16.5 17.3
27.0 27.1 27.5 27.1 26.5 26.7 28.3 27.2 27.7 27.8 27.4 26.0 26.5 23.8 25.3 28.4
3
22.2
22.9
33.2
26.0
33.0
26.0
32.3
25.8
28.1
30.5
27.6
30.4
27.6
23.5
29.3
27.0
22.7
29.2
-4.9 -17.4 -0.9 -0.9 0.4 -28.6 -0.1 -2.6 -2.4 11.1 -3.2 -0.8 -2.1 -0.7
-3.9 -15.2 0.1 0.1 1.2 -26.7 1.0 -1.6 -1.5 11.9 -2.2 0.1 -1.3 0.2
31.6 29.6 32.4 33.1 32.5 11.9 34.0 34.0 33.5 32.2 34.8 33.0 33.5 33.8
24.3 22.3 25.8 25.5 25.7 5.2 25.6 25.3 24.9 26.4 25.6 25.6 25.5 25.4
30.1 28.1 31.2 32.0 31.5 10.4 33.0 33.0 32.0 32.0 33.1 31.2 32.5 32.8
23.7 21.2 25.8 25.0 25.4 4.7 25.5 25.2 24.9 26.4 25.1 25.3 25.2 25.2
28.6 26.7 30.2 30.9 30.5 9.0 32.0 32.0 30.5 31.2 32.0 30.2 31.5 31.9
23.1 20.4 25.5 24.8 25.1 4.2 25.0 25.0 24.5 26.3 24.7 25.0 24.9 24.8
25.4 23.7 26.7 26.2 26.6 8.1 26.7 26.2 26.3 27.9 26.6 26.7 26.4 26.3
29.4 28.0 30.2 30.5 30.7 9.4 31.3 32.0 30.6 30.2 31.2 30.3 31.4 31.5
24.7 22.7 26.3 25.8 26.1 7.3 26.2 25.7 25.7 27.3 26.1 26.2 25.9 25.9
28.6 26.5 29.7 30.0 30.0 8.4 30.7 31.3 29.9 29.8 30.6 29.5 30.6 30.8
24.2 22.3 25.9 25.1 25.4 7.7 25.2 24.6 25.1 27.2 25.2 26.0 25.1 24.9
19.2 17.3 21.4 20.4 20.6 10.4 20.4 19.6 20.2 23.2 20.5 21.4 20.3 20.1
27.5 26.2 29.1 28.1 29.1 8.3 29.0 29.0 28.8 29.5 28.0 28.5 28.8 28.8
23.6 21.4 25.2 24.8 24.9 6.9 25.0 24.2 24.5 26.7 25.0 25.2 24.6 24.5
18.4 16.3 20.4 19.9 20.0 9.9 20.1 19.1 19.5 22.4 20.2 20.4 19.7 19.6
27.0 25.2 28.3 28.1 28.6 7.6 28.8 28.7 28.2 29.0 27.9 27.9 28.5 28.6
Iran, Islamic Republic of ABADAN AHVAZ ARAK BANDAR ABBASS INTL BANDAR ANZALI HAMEDAN ISFAHAN SHAHID BEHESHTI INTL KASHAN KERMAN MASHHAD INTL MEHRABAD INTL SHAHID ASHRAFI ESFAHANI SHIRAZ SHAHID DASTGHAIB INTL TABRIZ INTL URMIA ZAHEDAN INTL ZANJAN
30.371N 31.337N 34.138N 27.218N 37.467N 34.850N 32.751N 33.895N 30.274N 36.235N 35.689N 34.346N 29.539N 38.134N 37.668N 29.476N 36.774N
48.228E 48.762E 49.847E 56.378E 49.467E 48.533E 51.862E 51.577E 56.951E 59.641E 51.313E 47.158E 52.589E 46.235E 45.069E 60.907E 48.359E
Ireland CASEMENT DUBLIN
Israel
Licensed for single user. © 2017 ASHRAE, Inc.
TEL AVIV BEN GURION TEL AVIV SDE DOV
Italy BARI KAROL WOJTYLA BOLOGNA CATANIA FONTANAROSSA CATANIA SIGONELLA FIRENZE PERETOLA GENOVA SESTRI GRAZZANISE MILANO LINATE NAPOLI CAPODICHINO PALERMO FALCONE-BORSELLINO PRATICA DI MARE ROMA CIAMPINO ROMA LEONARDO DA VINCI TORINO BRIC DELLA CROCE TORINO CASELLE TRIESTE
16.767E 11.289E 15.066E 14.917E 11.205E 8.842E 14.082E 9.277E 14.291E 13.091E 12.445E 12.595E 12.251E 7.733E 7.650E 13.750E
Jamaica NORMAN MANLEY INTL
17.936N 76.788W
Japan AKITA ASAHIKAWA ASHIYA ATSUGI NAS CHIBA FUJISAN FUKUOKA FUKUYAMA FUSHIKI FUTENMA GIFU HAMAMATSU HIMEJI HIROSHIMA
39.717N 140.100E 22 43.750N 142.367E 140 33.883N 130.653E 30 35.450N 139.450E 51 35.600N 140.100E 6 35.367N 138.733E 3778 33.586N 130.451E 10 34.450N 133.250E 3 36.800N 137.050E 13 26.267N 127.750E 75 35.394N 136.870E 39 34.750N 137.703E 46 34.833N 134.667E 40 34.400N 132.467E 54
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Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 17 sites, 50 more on CD-ROM 10.4 9.2 8.1 416 3313 8.7 7.3 6.3 429 3326 8.5 7.6 6.5 2424 895 8.8 7.6 6.9 74 3252 11.9 9.7 7.7 1503 889 10.3 8.6 7.3 2832 556 10.2 8.4 7.1 2015 1056 7.6 6.0 5.0 1481 1848 10.8 9.2 7.7 1617 1031 9.0 7.8 6.8 2049 1047 11.0 9.6 7.7 1592 1548 9.5 8.2 7.2 2054 1018 9.1 7.6 6.4 1371 1425 10.5 9.4 8.1 2646 823 8.9 7.0 5.6 2880 457 11.5 9.8 8.4 1181 1469 11.5 9.7 7.9 2938 454 2 sites, 15 more on CD-ROM 14.4 12.7 11.3 3148 7 13.4 12.0 10.7 3164 4 2 sites, 5 more on CD-ROM 9.9 8.7 7.8 522 1438 11.6 9.6 8.1 488 1339 16 sites, 77 more on CD-ROM 9.3 8.1 7.1 1482 689 7.3 6.2 5.4 2130 682 9.8 8.4 7.4 1099 829 9.7 8.4 7.5 1076 990 8.3 7.0 6.0 1667 726 11.3 10.1 9.1 1356 647 9.8 8.2 6.9 1534 634 7.0 5.4 4.4 2142 634 8.7 7.3 6.3 1276 791 13.3 11.6 10.2 799 974 10.2 8.7 7.6 1349 610 11.0 9.0 7.5 1573 681 10.8 9.3 8.1 1482 572 8.7 7.0 5.4 2600 283 6.4 5.0 4.2 2437 399 12.1 10.2 8.6 1752 665 1 site, 1 more on CD-ROM 13.5 12.2 10.9 0 3586 65 sites, 133 more on CD-ROM 12.2 10.6 9.3 2796 519 8.7 7.3 6.1 4271 245 10.3 9.0 8.0 1711 843 10.6 9.3 8.2 1671 893 12.1 10.3 8.8 1623 886 29.2 25.8 22.9 8881 0 9.3 8.3 7.4 1532 1060 6.0 5.1 4.4 1851 966 7.5 6.4 5.5 2198 739 11.3 9.7 8.5 186 1870 8.2 7.1 6.1 1972 923 9.6 8.8 8.0 1602 894 8.3 7.1 6.1 1876 928 9.2 8.1 7.2 1643 1045
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station IIZUKA IRUMA KADENA AB KAGOSHIMA KANAZAWA KANSAI INTL KOBE KOCHI KOMATSU KUMAGAYA KUMAMOTO KURE KYOTO MATSUYAMA MINAMITORISHIMA MIYAZAKI NAGANO NAGASAKI NAGOYA NAHA NARA NAZE NIIGATA NYUTABARU OITA OKAYAMA OMAEZAKI ONAHAMA OSAKA OSAKA INTL OTARU OZUKI SAPPORO SENDAI SHIMOFUSA SHIMONOSEKI SHIZUHAMA SHIZUOKA SUMOTO TADOTSU TAKAMATSU TOKUSHIMA TOKYO TOKYO INTL TOYAMA TSUIKI UTSUNOMIYA WAKAYAMA YOKOHAMA YOKOSUKA YOKOTA AB
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
33.650N 130.700E 35.842N 139.411E 26.350N 127.767E 31.550N 130.550E 36.583N 136.633E 34.433N 135.233E 34.700N 135.217E 33.567N 133.550E 36.395N 136.407E 36.150N 139.383E 32.817N 130.700E 34.233N 132.550E 35.017N 135.733E 33.850N 132.783E 24.290N 153.979E 31.933N 131.417E 36.667N 138.200E 32.733N 129.867E 35.255N 136.924E 26.196N 127.646E 34.700N 135.833E 28.383N 129.500E 37.900N 139.017E 32.084N 131.451E 33.233N 131.617E 34.667N 133.917E 34.600N 138.217E 36.950N 140.900E 34.683N 135.517E 34.786N 135.438E 43.183N 141.017E 34.045N 131.052E 43.067N 141.333E 38.267N 140.900E 35.799N 140.011E 33.950N 130.933E 34.813N 138.298E 34.983N 138.400E 34.333N 134.900E 34.283N 133.750E 34.317N 134.050E 34.067N 134.567E 35.683N 139.767E 35.552N 139.780E 36.717N 137.200E 33.685N 131.040E 36.550N 139.867E 34.233N 135.167E 35.433N 139.650E 35.283N 139.667E 35.750N 139.350E
38 90 48 32 34 8 31 5 11 32 39 5 47 34 7 15 420 36 16 4 109 8 6 79 14 19 47 5 83 15 26 4 26 44 30 20 7 16 112 5 14 6 36 11 17 17 140 18 43 49 142
99.6% -1.8 -3.1 9.1 1.4 -1.2 1.8 0.1 -0.7 -2.0 -2.1 -1.6 0.0 -0.8 0.0 18.0 -0.1 -6.7 0.8 -2.0 12.8 -2.1 9.5 -1.7 -1.1 -0.4 -1.1 0.2 -2.3 0.7 -1.2 -9.5 -0.2 -9.9 -3.6 -2.0 1.6 -0.2 -0.2 0.0 -0.1 -0.3 0.6 0.9 0.9 -2.4 -2.0 -4.1 0.5 0.7 1.9 -3.8
99% -0.7 -2.1 10.8 2.6 -0.5 2.8 1.2 0.3 -1.0 -1.2 -0.5 0.9 0.1 0.9 18.7 1.1 -5.4 1.9 -1.0 13.1 -1.4 10.4 -0.9 0.1 0.7 -0.2 1.2 -1.3 1.6 -0.2 -8.3 0.8 -8.6 -2.6 -0.9 2.7 0.8 0.9 0.9 0.9 0.6 1.5 1.7 1.9 -1.5 -1.0 -3.1 1.4 1.5 2.9 -2.7
31.973N 35.992E 32.550N 35.850E 31.723N 35.993E
779 619 730
1.1 1.7 -0.8
2.7 3.3 0.7
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 33.6 25.6 32.6 25.4 31.5 25.1 34.2 25.3 33.0 25.0 31.2 24.5 33.1 26.8 32.2 26.7 32.0 26.6 33.5 25.8 32.6 25.6 31.8 25.4 33.2 24.8 32.1 24.7 31.0 24.4 33.0 25.4 32.0 25.2 31.1 25.1 33.2 25.1 32.1 24.9 31.1 24.7 32.9 25.3 32.0 25.1 31.2 24.9 33.2 24.6 32.0 24.5 30.9 24.4 35.5 25.5 34.0 25.1 32.5 24.5 34.5 25.3 33.5 25.1 32.5 24.8 32.5 25.2 31.6 24.9 30.7 24.6 35.0 24.7 33.9 24.4 32.7 24.1 33.4 24.8 32.6 24.6 31.7 24.4 31.9 26.3 31.4 26.2 31.0 26.0 33.8 25.6 32.6 25.6 31.5 25.4 32.9 23.6 31.5 23.2 30.1 22.6 32.8 25.5 31.8 25.4 30.9 25.2 35.1 25.0 33.9 24.8 32.8 24.5 32.2 26.4 32.0 26.4 31.2 26.5 34.1 24.9 33.0 24.7 31.9 24.4 32.9 26.0 32.2 26.0 31.6 25.9 32.9 25.1 31.6 24.7 30.2 24.3 32.9 25.6 31.2 25.7 30.2 25.6 33.5 25.3 32.5 25.2 31.4 24.9 34.4 25.2 33.5 24.9 32.4 24.7 30.3 25.9 29.6 25.6 28.9 25.2 29.1 24.2 28.0 24.0 27.1 23.5 34.4 24.8 33.4 24.6 32.4 24.4 34.9 25.5 33.8 25.2 32.8 24.9 28.2 22.2 26.6 21.2 25.2 20.6 32.2 25.8 31.2 25.8 30.8 25.8 29.2 22.6 27.7 21.6 26.3 20.7 31.4 24.5 29.9 24.0 28.5 23.3 33.6 25.8 32.2 25.3 31.0 25.1 32.2 25.7 31.3 25.4 30.5 25.2 32.2 25.8 31.0 25.8 30.0 25.4 32.9 25.3 31.6 25.1 30.6 24.8 31.8 25.4 30.8 25.2 29.8 24.9 33.6 25.0 32.7 24.8 31.7 24.6 34.4 25.2 33.4 25.0 32.3 24.9 33.3 25.5 32.3 25.3 31.3 25.0 33.5 25.2 32.4 24.9 31.4 24.5 32.9 25.8 31.9 25.4 30.8 25.2 33.8 25.4 32.5 25.2 31.1 24.7 32.8 25.9 31.2 25.8 30.2 25.5 33.6 25.4 32.2 25.0 30.8 24.4 33.4 24.7 32.4 24.7 31.4 24.6 32.6 25.5 31.6 25.1 30.5 24.7 34.1 25.9 32.1 25.2 30.8 24.8 34.0 25.8 32.8 25.4 31.2 24.6
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 26.6 31.5 26.1 30.7 26.4 31.6 25.8 30.8 28.6 30.6 28.1 30.1 26.8 31.1 26.5 30.6 25.9 30.7 25.4 30.2 26.7 30.2 26.3 29.7 26.4 30.4 26.0 29.9 26.6 30.4 26.2 29.8 26.1 30.6 25.5 30.0 26.4 32.7 25.9 31.8 26.5 31.5 26.1 30.9 25.9 30.7 25.5 30.1 25.7 32.2 25.2 31.4 25.7 31.1 25.3 30.6 27.2 30.5 26.9 30.2 26.8 30.9 26.4 30.4 24.4 30.7 23.8 29.6 26.7 30.2 26.3 29.8 26.2 31.4 25.8 30.8 27.7 30.2 27.3 30.0 25.9 31.7 25.4 31.0 27.1 30.9 26.7 30.5 26.0 30.8 25.4 30.1 26.9 30.0 26.4 29.5 26.3 31.2 25.8 30.5 26.1 31.9 25.7 31.2 26.7 29.0 26.3 28.4 25.1 27.7 24.7 27.0 26.1 31.5 25.6 31.0 26.6 31.9 26.1 31.3 23.2 27.0 22.3 25.6 26.9 30.5 26.5 30.0 23.7 27.8 22.8 26.5 25.3 29.3 24.8 28.3 26.6 31.2 26.1 30.3 26.5 30.4 26.1 29.9 27.0 30.1 26.5 29.4 26.5 30.6 26.0 29.9 26.4 30.0 25.9 29.3 26.2 31.0 25.7 30.6 26.3 31.7 25.9 31.2 26.5 31.1 26.1 30.5 26.1 31.3 25.7 30.6 26.6 30.7 26.2 30.0 26.6 31.2 25.9 30.4 27.1 30.3 26.6 29.9 26.2 31.7 25.6 30.6 26.3 30.9 25.9 30.4 26.3 30.6 25.7 29.8 26.7 31.5 26.0 30.2 26.6 31.8 26.0 30.8
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 25.2 20.4 28.8 24.8 19.9 28.4 25.1 20.4 28.5 24.2 19.3 27.9 28.2 24.5 29.6 27.8 24.0 29.4 25.7 21.0 29.1 25.3 20.5 29.0 24.6 19.7 28.7 24.0 19.0 28.3 25.8 21.2 28.9 25.1 20.3 28.5 25.3 20.6 28.6 24.9 20.0 28.5 25.6 20.9 28.6 25.2 20.3 28.3 24.9 20.0 28.5 24.1 19.0 28.0 25.0 20.1 28.3 24.4 19.5 28.3 25.3 20.6 28.6 24.9 20.0 28.4 24.5 19.5 28.6 24.1 19.0 28.3 24.0 19.0 28.5 23.6 18.5 28.3 24.3 19.3 28.2 23.8 18.7 28.1 26.2 21.6 29.3 25.9 21.2 29.2 25.8 21.2 28.8 25.4 20.6 28.5 22.7 18.3 27.2 22.2 17.7 26.7 25.8 21.2 28.8 25.4 20.7 28.5 25.1 20.2 28.1 24.2 19.2 27.8 27.0 22.8 29.9 26.2 21.7 29.3 24.4 19.7 28.1 23.9 19.0 27.7 26.0 21.4 29.4 25.6 20.9 29.2 24.6 19.6 28.9 24.0 18.9 28.5 26.1 21.7 28.6 25.8 21.3 28.4 25.0 20.2 28.6 24.5 19.5 28.3 24.7 19.7 28.4 24.2 19.2 28.4 26.1 21.7 28.1 25.6 21.0 27.7 24.3 19.3 26.7 23.9 18.8 26.3 24.6 19.8 28.8 24.1 19.2 28.7 25.1 20.3 28.8 24.8 19.9 28.7 21.8 16.5 25.3 21.1 15.8 24.8 26.0 21.4 29.6 25.2 20.4 28.7 22.3 17.1 26.3 21.5 16.2 25.5 24.2 19.2 27.3 23.7 18.6 26.8 25.2 20.4 28.8 25.0 20.1 28.6 25.4 20.7 29.0 25.0 20.1 28.7 26.2 21.6 28.6 25.9 21.2 28.4 25.4 20.6 28.7 24.9 20.0 28.4 25.4 20.8 28.4 24.9 20.2 27.9 24.9 20.0 28.7 24.3 19.2 28.6 24.9 20.0 28.7 24.5 19.5 28.6 25.2 20.4 29.0 24.8 19.8 28.6 24.7 19.8 29.0 24.2 19.2 28.9 25.2 20.4 28.9 25.0 20.1 28.7 25.3 20.5 29.1 24.7 19.7 28.5 26.2 21.6 29.1 25.8 21.1 28.8 24.7 20.0 28.6 24.2 19.4 28.2 25.0 20.2 29.3 24.5 19.5 28.9 25.1 20.3 28.6 24.6 19.7 28.3 25.2 20.5 28.9 24.9 20.1 28.7 25.2 20.6 28.7 24.8 20.2 28.5
Jordan AMMAN IRBID MET QUEEN ALIA INTL
36.0 34.4 37.0
18.6 19.6 19.8
34.2 32.9 35.2
18.3 19.4 19.0
33.1 31.7 34.1
18.1 19.3 18.9
22.3 23.4 22.8
30.4 29.2 32.2
21.4 22.6 21.9
29.4 28.0 31.4
(Click here to access additional station tables on CD-ROM.)
20.0 21.9 20.0
16.2 17.8 16.1
25.6 25.9 26.6
19.0 21.1 18.9
15.1 17.0 15.0
24.8 24.7 26.3
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 7.1 6.1 5.3 1733 947 10.0 8.5 7.4 2005 786 11.9 10.0 8.9 216 1881 8.9 7.6 6.7 1085 1293 11.8 10.1 8.7 2013 824 12.4 10.7 9.4 1521 1073 9.7 8.4 7.3 1580 1079 5.4 4.5 4.0 1383 1059 11.4 9.8 8.5 2098 757 8.0 6.8 5.8 1858 900 7.2 6.1 5.3 1498 1158 7.5 6.3 5.4 1606 994 5.4 4.8 4.3 1743 1064 6.0 5.2 4.5 1567 1026 12.1 10.9 9.9 0 2749 9.2 7.9 6.8 1244 1103 7.9 6.9 6.1 2718 653 7.7 6.5 5.6 1356 1072 9.9 8.5 7.4 1782 1048 13.4 11.7 10.4 117 2042 4.5 3.9 3.3 1943 883 7.4 6.4 5.7 348 1617 10.2 8.8 7.7 2233 737 9.9 8.3 6.9 1361 962 7.1 6.2 5.4 1533 978 9.8 8.3 7.1 1729 1073 12.3 11.2 10.2 1454 844 8.2 7.1 6.3 2139 504 8.6 7.4 6.4 1534 1161 8.4 7.4 6.5 1725 1067 8.0 6.9 6.1 3697 225 11.3 9.7 8.4 1704 887 9.8 8.4 7.3 3576 298 10.1 8.6 7.4 2486 501 9.4 8.1 7.0 1812 829 10.0 8.6 7.5 1431 998 10.7 9.6 8.6 1522 869 6.3 5.5 5.0 1446 934 6.9 5.9 5.2 1742 866 7.3 6.3 5.4 1641 1032 8.0 6.8 5.9 1641 1065 9.0 7.8 6.9 1531 1032 8.4 7.3 6.4 1556 968 12.6 11.2 10.0 1576 901 9.1 7.7 6.6 2142 788 10.4 9.1 8.0 1852 835 9.3 7.8 6.5 2134 746 10.8 9.2 8.0 1547 1065 9.4 8.3 7.2 1593 883 13.0 11.3 9.9 1397 938 9.5 8.2 7.0 1982 774 3 sites, 6 more on CD-ROM 10.0 8.4 7.3 1192 1147 8.5 7.4 6.6 1097 1070 11.7 10.0 8.9 1354 829
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
99.6%
99%
851 355 538 125 422 652
-20.2 -32.8 -32.7 -34.9 -16.2 -20.1
-17.2 -30.0 -29.1 -31.5 -13.0 -16.9
34.2 32.3 32.0 33.0 37.8 35.8
18.4 17.7 16.7 18.4 19.0 17.9
32.7 30.2 30.1 31.0 36.2 34.2
18.0 17.2 16.3 18.2 18.6 17.6
31.0 28.4 28.2 29.1 35.0 32.8
17.6 16.8 15.6 17.6 18.3 17.3
20.4 19.6 18.4 20.6 20.9 19.6
29.8 27.3 27.0 27.9 33.3 31.0
19.5 18.8 17.7 19.7 20.0 18.9
28.9 26.4 26.0 27.1 32.4 30.6
17.2 17.2 16.0 18.3 16.9 16.1
13.7 12.8 12.1 13.4 12.7 12.4
23.8 21.5 20.3 23.1 25.7 22.2
16.2 16.2 15.0 17.2 15.8 15.1
12.8 12.0 11.4 12.5 11.8 11.6
23.3 21.0 19.7 22.6 24.9 22.2
4.035S 39.594E 61 1.319S 36.928E 1624
20.1 10.0
20.8 11.2
33.1 29.0
25.1 15.8
32.6 28.2
25.1 15.8
32.1 27.4
25.0 16.0
26.5 18.9
30.2 23.5
26.2 18.5
29.8 23.2
25.5 17.7
20.9 15.5
27.9 19.6
25.2 17.2
20.5 15.0
27.7 19.2
Kazakhstan ALMATY ASTANA INTL KARAGANDY INTL PAVLODAR SHYMKENT TARAZ
43.233N 51.022N 49.671N 52.195N 42.364N 42.850N
76.933E 71.467E 73.334E 77.074E 69.479E 71.300E
Kenya MOMBASA INTL NAIROBI JOMO KENYATTA INTL
Korea, Democratic People's Republic of CHONGJIN HAMHUNG KAESONG NAMPO PYONGYANG SUNAN INTL SINUIJU WONSAN
41.783N 129.817E 39.933N 127.550E 37.967N 126.567E 38.717N 125.383E 39.224N 125.670E 40.100N 124.383E 39.183N 127.433E
43 22 70 47 36 7 36
-12.4 -12.9 -12.8 -12.8 -14.7 -15.5 -10.2
-10.9 -11.0 -10.9 -10.9 -12.7 -13.6 -8.4
27.4 31.5 30.9 30.2 31.2 30.9 31.7
22.6 23.9 25.2 25.5 24.4 24.1 23.5
26.2 29.9 29.6 29.1 30.1 29.5 30.1
22.0 23.4 24.3 24.7 23.9 23.4 23.0
25.1 28.4 28.4 28.0 29.0 28.2 28.5
21.5 22.6 23.6 24.0 23.3 22.9 22.6
24.1 25.6 26.5 26.5 26.1 25.9 25.5
26.2 29.5 29.4 29.2 29.6 28.8 29.2
23.2 24.7 25.8 25.8 25.4 25.1 24.7
25.1 28.1 28.1 28.0 28.4 27.5 28.1
23.4 24.4 25.7 25.6 25.1 25.1 24.5
18.2 19.5 21.2 21.0 20.3 20.2 19.5
25.4 27.7 28.0 28.2 28.1 27.6 27.5
22.6 23.7 25.1 25.1 24.5 24.4 23.7
17.4 18.6 20.4 20.3 19.5 19.4 18.7
24.6 26.8 27.3 27.4 27.3 26.6 26.6
35.100N 129.033E 35.180N 128.938E 35.170N 128.573E 36.633N 127.450E 36.717N 127.499E 35.883N 128.617E 35.894N 128.659E 36.300N 127.400E 37.558N 126.791E 35.167N 126.900E 35.126N 126.809E 37.467N 126.633E 33.517N 126.533E 33.511N 126.493E 35.833N 127.117E 35.167N 128.033E 37.091N 127.030E 35.988N 129.420E 36.967N 127.033E 33.250N 126.567E 37.446N 127.114E 37.567N 126.967E 37.533N 126.967E 37.267N 126.983E 35.593N 129.352E 34.400N 126.700E 34.733N 127.750E
71 2 37 59 58 61 35 78 18 74 12 70 23 36 63 30 12 21 16 51 28 87 29 35 14 35 67
-5.0 -6.1 -5.0 -10.7 -13.2 -6.9 -8.1 -10.7 -13.2 -6.5 -7.1 -10.2 0.4 -0.1 -8.7 -8.5 -13.1 -7.1 -12.6 0.5 -14.8 -11.4 -11.8 -11.3 -5.8 -3.9 -4.7
-3.3 -4.9 -3.4 -8.8 -11.1 -5.3 -6.8 -8.8 -11.2 -5.0 -5.8 -8.3 1.3 1.0 -6.9 -7.0 -11.1 -5.8 -10.8 1.7 -12.0 -9.3 -10.1 -9.4 -4.3 -2.5 -3.3
31.2 32.9 32.4 32.8 33.1 34.3 34.8 32.5 32.1 32.6 34.1 31.0 32.0 32.0 33.2 33.0 33.0 34.0 32.9 31.5 33.2 32.1 33.2 32.2 33.2 31.3 30.4
25.4 25.9 25.4 24.6 26.2 24.1 25.4 25.2 24.8 25.0 26.3 24.7 25.3 26.2 24.9 24.8 26.3 25.9 25.8 26.4 25.3 24.1 25.2 25.0 24.8 25.9 25.1
30.1 31.2 31.1 31.5 32.0 32.9 33.1 31.2 30.9 31.4 32.9 29.7 31.0 30.9 32.1 31.7 31.8 32.7 31.2 30.6 31.9 30.9 32.0 30.9 31.9 30.1 29.4
25.1 25.4 25.1 23.8 25.3 23.7 24.8 24.4 24.4 24.5 25.6 24.1 25.2 26.4 24.4 24.4 25.5 25.4 24.9 26.2 24.6 23.3 24.5 24.2 24.5 25.4 24.7
29.0 30.1 29.9 30.3 30.8 31.6 31.9 30.0 29.8 30.3 31.2 28.6 30.0 29.9 30.8 30.4 30.2 31.0 30.1 29.7 30.2 29.7 30.8 29.7 30.5 29.0 28.4
24.7 24.8 24.5 23.1 24.6 23.2 24.2 23.6 23.5 23.9 24.8 23.4 25.0 26.0 23.6 23.9 24.5 24.8 24.2 25.8 23.5 22.7 24.1 23.5 24.0 24.8 24.4
26.5 26.7 26.5 26.0 27.2 25.7 26.7 26.5 26.7 26.1 27.1 26.3 26.8 28.1 26.4 26.3 27.8 26.8 27.1 27.3 26.3 25.9 26.3 26.2 26.1 26.6 26.3
29.5 30.8 30.4 30.0 31.2 30.7 31.7 30.4 29.6 30.2 31.6 29.0 30.0 30.1 30.1 30.4 30.0 32.0 30.4 30.1 31.1 29.7 31.0 29.7 30.6 29.9 28.6
25.9 26.2 25.9 25.3 26.4 25.2 26.1 25.7 26.0 25.6 26.5 25.5 26.2 27.3 25.8 25.7 26.9 26.2 26.3 26.9 25.7 25.1 25.8 25.6 25.5 26.1 25.7
28.8 30.1 29.6 29.2 30.1 30.1 30.8 29.4 28.8 29.5 30.9 28.1 29.5 29.3 29.5 29.5 29.6 30.9 29.4 29.6 30.0 28.7 30.1 28.9 29.8 29.2 28.0
25.7 25.8 25.4 24.9 26.1 24.5 25.2 25.4 26.0 25.0 26.1 25.6 25.9 27.8 25.5 25.2 27.2 25.2 26.2 26.6 25.1 24.8 25.1 25.3 24.9 25.7 25.7
21.1 21.1 20.6 20.2 21.7 19.6 20.4 20.8 21.4 20.3 21.5 21.0 21.3 23.9 20.9 20.4 23.0 20.4 21.7 22.3 20.2 20.1 20.3 20.5 20.0 21.0 21.1
28.3 29.6 28.8 27.9 29.4 27.9 28.7 28.4 28.0 28.0 29.1 27.8 29.2 29.8 28.2 28.2 28.5 29.2 28.3 29.3 28.7 27.8 28.3 27.8 28.2 28.8 27.7
25.1 25.0 24.8 24.2 25.2 23.9 24.9 24.8 25.1 24.5 25.2 24.7 25.2 26.9 24.8 24.7 26.2 24.9 25.7 26.1 24.2 24.1 24.7 24.7 24.3 25.1 25.1
20.3 20.1 19.9 19.3 20.5 18.9 20.1 20.0 20.3 19.6 20.3 19.9 20.4 22.7 20.0 19.8 21.6 20.1 21.0 21.6 19.2 19.1 19.8 19.8 19.2 20.3 20.3
28.0 28.8 28.3 27.5 28.4 27.6 28.5 27.7 27.3 27.8 28.4 27.1 28.6 29.0 27.7 27.8 27.9 29.0 28.0 28.9 27.6 27.2 28.0 27.4 28.0 28.1 27.2
42.850N 74.533E
760
-17.9
-14.9
35.2
19.2
33.7
18.4
32.3
17.9
21.7
32.3
20.2
30.5
18.0
14.2
27.3
16.5
12.9
25.2
56.967N 24.050E
26
-19.2
-15.1
28.9
20.1
27.0
19.7
25.1
18.3
21.6
26.6
20.4
25.0
19.9
14.7
24.3
18.9
13.7
23.0
33.821N 35.488E
27
8.1
9.5
32.6
23.4
31.3
24.2
30.8
24.3
26.8
30.5
26.2
29.9
25.9
21.3
29.9
25.1
20.3
29.4
32.097N 20.269E 32.417N 15.050E 32.664N 13.159E
132 32 80
6.8 8.3 4.6
7.8 9.2 5.8
37.1 36.9 42.0
20.7 21.5 22.9
35.2 34.5 39.9
20.5 21.4 22.3
33.7 32.5 38.0
20.2 21.6 21.9
24.8 26.4 26.7
29.7 29.5 35.2
24.2 25.8 25.5
29.0 29.0 32.9
23.7 25.6 24.7
18.8 20.9 19.9
26.9 28.4 30.2
22.9 24.9 23.5
17.9 20.0 18.5
26.6 28.0 29.1
Licensed for single user. © 2017 ASHRAE, Inc.
Korea, Republic of BUSAN BUSAN GIMHAE INTL CHANGWON CHEONGJU CHEONGJU INTL DAEGU DAEGU AB DAEJEON GIMPO INTL GWANGJU GWANGJU INTL INCHEON JEJU JEJU INTL JEONJU JINJU OSAN AB POHANG PYEONGTAEK A511 SEOGWIPO SEOUL AB SEOUL CITY SEOUL SINYONGSAN SUWON ULSAN WANDO YEOSU
Kyrgyzstan BISHKEK
Latvia RIGA
Lebanon BEIRUT RAFIC HARIRI INTL
Libyan Arab Jamahiriya BENINA INTL MISRATA TRIPOLI INTL
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 6 sites, 75 more on CD-ROM 6.3 5.1 4.2 3586 477 12.0 10.3 9.0 5735 191 12.3 10.4 9.0 5624 166 10.3 8.9 7.7 5726 246 8.0 6.8 5.9 2581 864 11.1 8.4 6.4 3178 608 2 sites, 15 more on CD-ROM 9.4 8.4 7.8 0 3036 9.4 8.4 7.6 84 590 7 sites, 20 more on CD-ROM 7.5 5.9 4.6 3776 232 8.0 6.6 5.3 3187 423 9.2 7.8 6.5 3061 570 9.8 8.1 6.8 3163 583 7.0 5.8 4.9 3243 619 7.5 6.3 5.4 3481 525 7.7 6.4 5.4 2918 457 27 sites, 32 more on CD-ROM 9.9 8.4 7.4 1862 717 8.8 7.6 6.9 2098 786 6.4 5.6 5.1 1968 797 6.2 5.2 4.4 2657 755 7.2 6.0 5.2 2850 724 7.5 6.4 5.6 2185 862 8.9 7.7 6.6 2329 823 7.0 5.7 4.8 2682 696 8.2 7.1 6.2 3002 654 7.3 6.2 5.3 2261 812 7.6 6.5 5.7 2386 838 8.9 7.5 6.4 2701 638 10.2 8.8 7.5 1647 823 12.3 11.0 9.8 1760 760 5.7 5.1 4.4 2437 817 6.7 5.6 4.8 2392 724 8.4 7.2 6.2 2874 709 9.6 8.3 7.4 2237 686 8.0 6.8 5.9 2873 701 7.9 6.8 5.9 1384 905 6.2 5.3 4.5 2911 671 6.8 5.9 5.3 2684 736 6.3 5.2 4.5 2620 796 6.2 5.3 4.6 2773 712 6.9 6.0 5.3 2083 724 12.0 10.3 9.1 2096 706 12.1 10.5 9.2 2018 696 1 site, 8 more on CD-ROM 7.6 6.3 5.1 3019 640 1 site, 19 more on CD-ROM 9.1 8.1 7.3 4123 93 1 site, 2 more on CD-ROM 10.4 8.5 7.2 399 1493 3 sites, 5 more on CD-ROM 15.5 14.0 12.8 606 1363 13.1 10.7 9.5 448 1408 10.5 9.5 8.5 631 1687
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB 99.6%
99%
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
Lithuania KAUNAS VILNIUS INTL
54.883N 23.833E 54.634N 25.286E
77 197
-18.7 -19.3
-15.3 -16.0
28.4 28.7
19.9 19.3
26.6 26.8
18.9 18.5
24.8 25.0
17.9 17.6
21.1 20.8
26.4 26.1
19.9 19.7
24.6 24.4
19.3 19.1
14.2 14.2
23.4 22.9
18.2 18.1
13.2 13.3
22.2 21.7
22.150N 113.592E
6
7.4
9.0
32.9
27.0
32.1
26.9
31.2
26.8
28.2
30.7
27.8
30.2
27.2
23.1
29.6
27.1
22.9
29.5
238
-12.0
-8.2
36.0
20.3
34.2
20.1
32.8
19.6
21.9
32.0
21.2
31.0
18.9
14.1
24.6
18.0
13.3
23.8
18.797S 47.479E 1280
8.0
9.0
29.9
19.7
29.0
19.7
28.1
19.7
23.1
26.7
22.4
26.2
22.0
19.6
25.0
21.1
18.5
24.1
5.937N 3.131N 3.775N 1.485N 5.901N 4.313N
3 27 18 27 14 17
22.8 22.7 21.7 22.0 22.9 22.1
23.0 23.0 22.1 22.4 23.3 22.7
33.8 34.8 34.1 34.0 33.8 32.7
28.0 26.3 26.8 25.9 26.5 26.0
33.2 34.1 33.4 33.2 33.1 32.2
27.7 26.1 26.7 25.9 26.4 26.1
32.9 33.8 33.0 32.9 32.5 31.9
27.5 26.1 26.6 25.9 26.4 26.1
28.8 28.1 28.2 27.5 27.7 27.6
32.9 32.3 32.1 31.5 31.3 30.9
28.4 27.6 27.8 27.0 27.4 27.2
32.5 31.7 31.7 31.1 31.0 30.7
27.8 27.0 27.1 26.2 26.8 26.7
23.8 22.8 22.9 21.7 22.4 22.3
32.5 30.7 31.0 29.7 30.2 30.2
27.1 26.4 26.8 26.0 26.2 26.2
22.9 22.0 22.5 21.4 21.7 21.6
32.0 29.9 30.7 29.4 29.5 29.8
12.534N 7.950W
380
14.9
16.2
40.2
19.7
39.6
19.9
38.8
20.0
27.1
31.6
26.3
31.0
26.1
22.5
28.7
25.2
21.3
27.8
18.098N 15.948W
2
12.9
14.1
41.2
20.2
39.6
20.1
37.8
20.0
28.8
30.5
28.0
30.0
28.5
24.9
29.4
27.6
23.6
28.9
16.757N 99.754W 21.037N 86.877W 18.505N 88.327W 20.522N 103.311W 20.983N 101.483W 29.096N 111.048W 23.161N 106.266W 20.937N 89.658W 19.436N 99.072W 25.778N 100.107W 20.680N 105.254W 22.254N 100.931W 22.296N 97.866W 14.917N 92.250W 32.541N 116.970W 19.337N 99.566W 25.568N 103.411W 19.146N 96.187W
4 6 12 1529 1815 191 12 12 2230 390 7 1840 24 118 149 2580 1124 27
19.2 13.2 15.6 1.9 4.0 4.8 8.2 13.8 3.1 3.1 14.1 0.4 10.2 20.0 5.8 -2.0 3.8 14.8
20.8 14.9 17.4 3.1 5.9 6.2 9.9 15.8 4.9 5.0 15.2 2.2 12.0 20.9 6.8 -0.9 5.7 16.0
33.2 34.0 34.1 33.2 34.1 42.8 34.0 38.8 29.1 38.8 33.2 32.2 34.1 35.4 32.2 26.1 38.0 35.1
26.6 27.0 26.8 15.8 14.7 22.8 25.5 24.3 12.6 23.0 26.8 14.7 26.7 26.0 20.3 12.1 20.6 26.8
33.1 33.2 33.4 32.2 33.0 41.4 33.2 37.2 28.1 37.8 33.1 31.0 33.2 34.8 30.2 25.0 36.9 34.1
26.6 26.8 26.8 15.4 14.7 23.0 25.3 24.4 12.4 22.9 26.8 14.8 26.6 26.0 19.8 11.9 20.4 26.8
32.8 32.9 32.8 31.2 31.9 40.2 32.9 36.2 27.0 36.8 32.8 29.8 33.0 34.2 28.9 24.0 35.9 33.2
26.4 26.7 26.7 15.0 14.7 22.8 25.2 24.4 12.3 23.1 26.7 14.8 26.6 25.9 19.4 11.7 20.2 26.6
27.9 28.2 28.1 20.1 19.7 27.0 27.9 28.2 16.1 26.4 28.0 18.7 28.6 28.2 23.0 15.9 24.5 28.1
32.4 32.3 32.2 26.8 26.4 35.4 31.7 31.6 22.5 34.3 32.1 25.0 32.1 33.0 29.4 21.1 33.5 33.2
27.5 27.8 27.7 19.7 19.1 26.5 27.3 27.6 15.7 25.9 27.5 18.2 27.7 27.7 22.0 15.2 23.6 27.5
31.8 31.9 31.7 26.1 25.5 34.8 31.0 31.3 21.9 33.4 31.6 24.5 31.4 32.6 27.9 20.4 32.8 32.3
26.8 27.1 27.1 18.1 18.1 25.2 27.0 27.8 14.2 24.2 27.0 17.2 27.7 26.9 21.0 14.2 22.0 26.9
22.5 22.9 22.8 15.8 16.3 20.8 22.7 23.9 13.3 20.1 22.8 15.4 23.8 22.9 15.9 13.9 19.1 22.7
31.3 30.9 31.2 22.2 19.7 29.6 30.7 29.1 17.1 29.2 30.8 19.6 31.7 31.8 26.9 17.1 30.3 30.9
26.2 26.9 26.5 17.9 17.8 24.8 26.2 27.0 14.0 24.0 26.2 16.8 26.9 26.4 20.0 13.9 21.0 26.2
21.6 22.5 22.1 15.5 16.0 20.3 21.6 22.7 13.1 19.8 21.7 15.1 22.6 22.1 14.9 13.6 17.9 21.7
30.6 30.7 30.6 22.0 19.6 29.4 29.9 28.7 17.0 29.0 30.1 19.4 30.7 31.3 25.3 16.9 28.6 30.0
47.017N 28.983E
173
-14.3
-11.5
32.5
19.8
30.6
19.4
29.0
18.8
21.5
28.6
20.7
27.5
19.2
14.3
24.3
18.4
13.6
23.5
47.843N 106.767E 1330
-36.1
-33.9
31.0
15.7
28.8
15.0
26.7
14.3
17.7
25.6
16.7
24.3
15.2
12.7
19.8
14.1
11.8
19.3
Macao MACAU INTL
Macedonia, the former Yugoslav Republic of SKOPJE ALEXANDER THE GREAT
41.962N 21.621E
Madagascar IVATO
Malaysia KOTA KINABALU INTL KUALA LUMPUR SUBANG KUANTAN KUCHING INTL SANDAKAN TAWAU
116.051E 101.549E 103.209E 110.347E 118.059E 118.122E
Mali BAMAKO-SENOU INTL
Mauritania NOUAKCHOTT
Licensed for single user. © 2017 ASHRAE, Inc.
Mexico ACAPULCO INTL CANCUN INTL CHETUMAL INTL GUADALAJARA INTL GUANAJUATO INTL HERMOSILLO INTL MAZATLAN INTL MERIDA INTL MEXICO CITY INTL MONTERREY INTL PUERTO VALLARTA INTL SAN LUIS POTOSI INTL TAMPICO INTL TAPACHULA TIJUANA INTL TOLUCA INTL TORREON INTL VERACRUZ INTL
Moldova, Republic of CHISINAU
Mongolia CHINGGIS KHAAN INTL
Morocco AGADIR AL MASSIRA INTL CASABLANCA ANFA CASABLANCA MOHAMMED V INTL FES-SAISS INEZGANE MARRAKECH MENARA MEKNES BASSATINE OUJDA ANGADS RABAT SALE INTL TANGIER IBN BATOUTA TETOUAN SANIA RAMEL
30.325N 33.557N 33.367N 33.927N 30.381N 31.607N 33.879N 34.787N 34.051N 35.727N 35.594N
9.413W 7.660W 7.590W 4.978W 9.546W 8.036W 5.515W 1.924W 6.752W 5.917W 5.320W
76 62 200 579 27 468 576 468 84 19 3
5.0 6.4 3.0 0.8 5.0 4.0 2.3 0.9 4.8 4.1 6.1
6.1 7.6 4.2 2.0 6.5 5.2 3.8 2.2 5.9 5.8 7.6
38.8 29.7 35.9 39.6 35.2 42.0 39.0 37.9 32.9 33.2 33.0
19.9 21.5 21.4 19.9 19.4 20.6 20.9 20.6 21.5 21.4 20.4
35.0 27.6 33.1 37.6 31.8 39.9 36.8 35.9 30.0 31.9 31.1
19.1 21.8 21.3 19.9 18.8 20.5 20.8 20.7 21.2 21.3 20.4
32.1 26.2 31.1 35.8 29.0 37.9 34.8 34.0 28.0 30.2 29.7
19.0 21.8 21.0 19.7 18.4 20.3 20.5 20.4 21.4 21.1 20.4
22.5 24.0 23.5 22.4 22.5 23.0 23.7 23.7 24.2 23.1 24.2
30.6 26.7 31.3 33.6 28.7 35.5 33.8 32.5 28.5 29.7 27.4
21.9 23.3 22.7 21.5 21.8 22.2 22.6 22.9 23.3 22.6 23.6
29.2 25.9 29.9 32.8 26.6 34.4 32.4 31.0 27.0 28.7 26.8
20.2 23.1 21.2 19.0 21.0 19.4 20.5 21.1 23.0 21.1 23.2
15.0 18.0 16.2 14.8 15.7 14.9 16.3 16.7 17.9 15.8 17.9
24.2 25.6 25.5 26.8 24.1 26.5 29.9 26.9 26.6 25.8 25.7
19.8 22.3 20.7 18.0 20.2 18.8 19.4 20.4 22.1 20.5 22.7
14.7 17.2 15.7 13.9 15.0 14.4 15.2 16.0 16.9 15.2 17.4
24.0 24.7 24.9 25.6 23.2 25.7 27.9 26.3 25.4 25.5 25.4
25.921S 32.573E
44
11.9
13.0
35.4
23.7
33.7
23.7
32.1
23.7
26.6
31.3
26.1
30.5
25.3
20.6
28.7
24.9
20.1
28.3
Mozambique MAPUTO INTL
Netherlands (Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 2 sites, 8 more on CD-ROM 9.4 8.3 7.4 4100 84 10.0 8.7 7.7 4273 88 1 site, 0 more on CD-ROM 11.2 9.9 8.9 282 2014 1 site, 6 more on CD-ROM 8.7 7.5 6.2 2568 547 1 site, 3 more on CD-ROM 7.4 6.8 6.1 294 721 6 sites, 18 more on CD-ROM 6.3 5.2 4.4 0 3491 6.0 5.2 4.6 0 3683 6.0 5.2 4.6 0 3358 5.4 4.7 4.2 0 3277 6.5 5.6 5.2 0 3466 5.5 5.0 4.4 0 3245 1 site, 0 more on CD-ROM 8.1 6.9 6.1 0 3516 1 site, 1 more on CD-ROM 9.4 8.5 7.7 2 2986 18 sites, 22 more on CD-ROM 8.2 7.2 6.1 0 3280 9.5 8.2 7.4 3 2858 10.5 10.3 10.0 0 3259 9.8 8.0 6.3 352 721 10.3 9.3 7.8 267 794 8.1 6.9 5.9 211 2682 9.7 7.7 6.6 31 2131 10.1 8.8 7.9 1 3276 9.9 8.4 7.5 581 204 11.2 9.8 8.4 370 2142 7.7 6.6 5.8 1 2572 9.7 8.3 7.2 667 431 14.9 11.9 10.0 77 2561 10.4 6.6 4.8 0 3482 8.2 7.0 6.2 704 505 8.8 7.6 6.4 1776 2 8.9 6.8 5.5 315 2048 19.6 14.6 10.4 3 2743 1 site, 1 more on CD-ROM 7.4 6.2 5.3 3199 399 1 site, 59 more on CD-ROM 10.3 8.8 7.6 7049 98 11 sites, 9 more on CD-ROM 9.4 8.1 7.1 393 938 6.5 5.5 4.9 636 648 9.7 8.3 7.4 816 791 10.1 8.3 7.0 1201 876 10.5 8.7 7.2 522 649 7.4 6.3 5.3 620 1430 8.5 7.3 6.3 1097 869 11.7 10.0 8.9 1111 898 8.5 7.3 6.4 802 545 14.7 13.0 11.5 793 731 11.5 10.3 9.3 626 836 1 site, 0 more on CD-ROM 15.2 12.8 10.5 17 1978 6 sites, 37 more on CD-ROM
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station AMSTERDAM SCHIPHOL HOEK VAN HOLLAND IJMUIDEN ROTTERDAM THE HAGUE VALKENBURG WOENSDRECHT AB
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 27.8 19.8 25.6 19.0 23.7 18.0 27.2 19.3 24.7 18.4 22.8 18.0 25.6 18.7 23.6 17.8 21.8 17.6 28.1 19.9 25.8 19.1 24.0 18.1 27.1 19.6 24.8 18.7 22.9 17.8 29.1 19.8 26.7 19.1 24.8 18.2
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 20.7 25.9 19.7 24.2 20.6 25.0 19.6 23.2 20.0 23.3 19.2 21.5 20.9 26.2 19.8 24.4 20.6 25.2 19.5 23.4 21.0 26.7 20.0 25.0
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 19.0 13.8 22.8 18.1 13.0 21.6 19.1 13.9 22.2 18.2 13.2 21.1 19.0 13.8 21.1 18.4 13.3 20.2 19.1 13.9 23.0 18.1 13.0 21.8 19.0 13.7 22.5 18.0 13.0 21.2 19.1 13.9 22.9 18.1 13.1 22.0
Lat
Long
Elev
52.309N 51.983N 52.467N 51.957N 52.167N 51.449N
4.764E 4.100E 4.567E 4.437E 4.433E 4.342E
-3 14 4 -5 1 19
99.6% -6.2 -5.2 -6.4 -6.2 -6.3 -7.6
99% -4.2 -3.3 -4.1 -4.1 -4.2 -5.2
37.000S 37.008S 43.483S 43.489S
174.800E 174.792E 172.517E 172.532E
7 7 37 38
4.4 4.0 -2.4 -2.8
5.6 5.2 -1.4 -1.9
25.2 25.7 27.6 28.0
19.7 19.9 16.6 16.9
24.3 24.8 25.5 25.9
19.2 19.4 15.8 16.0
23.5 23.8 23.5 23.9
18.8 18.9 15.2 15.4
21.2 21.4 18.1 18.3
23.4 23.8 24.1 24.5
20.5 20.6 17.2 17.5
22.8 23.0 22.5 22.9
20.4 20.8 16.2 16.2
15.1 15.4 11.6 11.5
22.4 22.8 19.1 19.4
19.6 19.9 15.5 15.4
14.4 14.6 11.0 11.0
21.9 22.0 18.4 18.5
12.141N 86.168W
59
20.0
20.9
36.0
24.3
35.1
24.2
34.8
24.1
26.6
31.4
26.2
31.1
25.2
20.5
28.4
25.1
20.3
28.3
13.482N 2.184E
223
15.9
17.0
42.3
20.5
41.8
20.4
40.9
20.5
27.1
33.0
26.5
32.6
25.9
21.8
29.3
25.1
20.8
29.1
60.117N 10.833E 59.942N 10.720E
170 97
-19.1 -14.3
-16.1 -12.1
26.5 26.7
17.6 17.3
24.8 24.8
16.9 16.6
23.0 23.1
15.9 15.7
19.2 18.6
23.5 23.7
18.1 17.7
22.8 22.7
17.8 16.8
13.0 12.1
20.6 20.2
16.3 15.9
11.9 11.4
19.3 19.4
24.233N 55.783E
299
9.8
11.2
45.1
21.5
44.1
21.3
43.1
21.2
27.4
33.3
26.6
33.9
26.1
22.3
30.7
24.8
20.5
31.1
33.617N 73.099E 24.907N 67.161E 31.522N 74.404E
508 31 217
2.1 10.1 3.0
3.2 11.8 4.8
41.1 38.9 43.2
22.7 22.7 23.4
39.4 37.1 41.9
22.8 23.1 23.3
38.0 36.0 40.1
22.7 23.5 23.3
28.0 28.2 29.2
34.2 33.3 34.5
27.5 27.8 28.6
33.4 32.7 33.7
26.5 27.1 28.1
23.4 22.9 24.9
31.3 30.9 32.2
26.0 26.6 27.2
22.8 22.3 23.6
31.0 30.7 31.5
31.867N 35.217E
759
0.8
2.2
33.1
18.3
31.8
18.3
30.3
18.1
21.7
28.7
20.8
27.3
19.8
15.9
24.2
19.0
15.1
22.9
8.917N 79.600W 9.071N 79.383W
13 41
22.2 20.9
22.9 21.6
35.0 34.1
25.4 25.3
34.1 33.4
25.3 25.1
33.8 33.0
25.2 24.9
27.7 27.6
31.7 31.1
27.2 27.1
31.2 30.9
26.6 26.8
22.2 22.5
30.5 29.6
26.1 26.1
21.6 21.6
29.9 29.1
25.240S 57.519W
89
5.1
7.1
37.1
23.9
36.1
24.0
35.1
24.1
26.7
32.5
26.4
32.1
25.2
20.6
29.4
24.9
20.1
29.2
71.583W 2562 79.828W 30 71.939W 3310 79.109W 32 73.309W 93 77.114W 34 80.616W 35 74.574W 156
6.0 14.9 0.1 14.3 19.0 14.0 15.9 17.8
6.8 15.2 1.1 14.8 20.2 14.4 16.4 19.0
24.1 32.2 23.0 28.0 34.1 28.8 34.0 34.9
11.4 24.2 10.1 23.4 26.4 22.7 25.4 26.2
23.2 31.8 22.2 27.1 33.7 27.8 33.2 34.1
11.0 24.0 9.9 23.0 26.4 22.2 25.1 26.1
23.0 30.9 21.7 26.2 33.0 26.9 32.7 33.5
10.8 23.5 9.8 22.3 26.3 21.9 24.8 26.0
14.7 25.6 12.5 23.6 27.3 23.6 26.4 27.0
21.0 30.3 20.2 27.5 32.4 27.2 32.3 32.9
14.1 24.8 12.0 23.1 27.1 22.8 25.9 26.7
20.3 29.6 19.6 26.8 32.1 26.6 31.9 32.5
12.8 24.1 9.6 22.1 26.0 22.2 24.9 25.2
12.7 19.1 11.2 16.8 21.6 16.9 20.1 20.8
16.0 28.8 14.7 26.8 31.0 26.3 29.5 29.7
12.1 23.2 9.1 21.8 25.6 21.2 24.2 25.1
12.1 18.0 10.9 16.5 21.0 16.0 19.2 20.6
15.6 28.4 14.0 26.5 30.5 25.2 29.6 29.6
8.483N 124.650E 6 7.126N 125.646E 29 6.050N 125.100E 133 10.713N 122.541E 8 10.308N 123.979E 9 14.583N 120.983E 13 14.509N 121.020E 23 14.495N 120.904E 2 14.650N 121.050E 46 6.922N 122.060E 10
22.1 22.8 22.7 22.8 23.3 23.1 21.8 22.3 20.3 22.7
22.8 23.1 23.0 23.3 23.8 23.8 22.4 23.4 21.1 23.2
34.6 34.0 35.0 34.8 33.1 34.6 35.0 35.1 35.1 34.1
27.5 26.9 27.3 27.6 27.3 26.4 26.0 28.5 26.3 27.5
34.1 33.2 34.3 34.1 32.6 33.9 34.2 34.4 34.5 33.7
27.4 26.8 27.1 27.6 27.1 26.4 25.9 28.3 26.2 27.4
33.6 33.0 33.8 33.4 32.2 33.2 33.7 33.8 33.7 33.2
27.3 26.8 27.0 27.4 27.0 26.3 25.9 28.1 26.2 27.2
28.8 28.2 28.2 28.5 28.6 28.2 28.4 29.1 27.7 28.2
33.1 32.3 33.1 33.1 31.3 32.1 31.3 33.9 32.3 32.8
28.4 27.9 27.9 28.2 28.2 27.7 28.0 28.8 27.5 27.9
32.8 31.9 32.8 32.7 31.0 31.6 30.8 33.4 31.9 32.4
27.6 27.1 26.8 27.2 27.9 27.1 27.8 27.7 26.6 27.0
23.6 22.9 22.9 23.0 24.0 22.9 23.9 23.7 22.3 22.8
32.3 31.4 31.6 31.6 30.4 30.8 30.0 32.3 30.1 31.6
27.2 26.8 26.6 26.9 27.5 26.7 27.2 27.5 26.2 26.7
23.0 22.5 22.5 22.6 23.5 22.3 23.0 23.5 21.8 22.3
32.1 31.1 31.4 31.5 30.2 30.4 29.4 32.2 29.7 31.2
54.378N 54.333N 54.600N 50.233N 50.078N 51.733N
-15.0 -17.0 -9.1 -14.9 -15.8 -14.7
-12.0 -12.7 -7.1 -11.9 -12.9 -11.9
27.2 25.8 25.4 29.7 30.1 29.9
19.1 19.5 20.3 19.9 20.4 19.6
25.2 23.6 23.8 27.8 28.1 28.0
18.2 18.3 19.3 18.9 19.6 18.8
23.8 21.9 22.4 26.0 26.3 26.2
17.5 17.6 18.6 18.1 18.7 18.0
20.4 20.3 21.1 20.8 21.4 21.0
25.2 24.4 24.2 27.0 27.9 27.0
19.4 19.1 20.1 19.9 20.5 20.0
23.9 22.5 22.9 26.0 26.6 25.8
18.9 18.8 19.9 18.8 19.2 19.0
13.9 13.6 14.6 14.1 14.3 14.1
22.7 22.1 23.0 23.4 24.1 23.1
17.8 17.7 19.0 17.9 18.3 18.1
13.0 12.7 13.8 13.3 13.5 13.3
21.6 21.1 21.7 22.2 22.8 22.1
New Zealand AUCKLAND AERO AWS AUCKLAND INTL CHRISTCHURCH AP AWS CHRISTCHURCH INTL
Nicaragua MANAGUA INTL
Niger NIAMEY DIORI HAMANI INTL
Norway HAKADAL OSLO-BLINDERN
Oman AL BURAIMI
Pakistan
Licensed for single user. © 2017 ASHRAE, Inc.
BENAZIR BHUTTO INTL JINNAH INTL LAHORE ALLAMA IQBAL INTL
Palestinian Territory, Occupied JERUSALEM ATAROT
Panama PANAMA PACIFICO INTL TOCUMEN INTL
Paraguay SILVIO PETTIROSSI INTL
Peru AREQUIPA INTL CHICLAYO INTL CUSCO INTL HUANCHACO INTL IQUITOS INTL LIMA CALLAO INTL PIURA INTL PUCALLPA INTL
16.341S 6.787S 13.536S 8.081S 3.785S 12.022S 5.206S 8.378S
Philippines CAGAYAN DE ORO FRANCISCO BANGOY INTL GENERAL SANTOS INTL ILOILO MACTAN CEBU INTL MANILA NINOY AQUINO INTL SANGLEY POINT AB SCIENCE GARDEN ZAMBOANGA INTL
Poland GDANSK LECHA WALESY GDANSK-SWIBNO HEL KATOWICE KRAKOW BALICE LODZ
18.466E 18.933E 18.817E 19.033E 19.785E 19.400E
149 9 3 284 241 190
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 13.5 11.8 10.3 2929 71 16.0 14.4 13.1 2757 70 18.6 16.4 15.0 2921 51 12.1 10.5 9.3 2902 71 13.1 11.6 10.3 2941 56 9.7 8.4 7.4 2963 78 4 sites, 35 more on CD-ROM 12.6 11.1 9.9 1222 157 12.6 11.3 10.1 1218 172 11.4 10.1 9.0 2597 49 11.3 10.0 9.0 2619 56 1 site, 0 more on CD-ROM 8.3 7.4 6.5 0 3479 1 site, 12 more on CD-ROM 9.3 8.0 7.0 0 4249 2 sites, 117 more on CD-ROM 8.0 6.7 5.7 4513 44 8.0 6.9 6.0 4182 54 1 site, 20 more on CD-ROM 8.2 7.1 6.3 71 3756 3 sites, 25 more on CD-ROM 13.0 10.4 9.3 639 2044 9.7 8.5 7.7 24 3248 8.0 6.4 5.4 444 2608 1 site, 0 more on CD-ROM 9.4 8.5 7.7 1293 789 2 sites, 0 more on CD-ROM 8.1 7.2 6.4 0 3566 7.7 6.5 5.6 0 3363 1 site, 3 more on CD-ROM 10.4 9.4 8.5 257 2109 8 sites, 5 more on CD-ROM 9.0 7.6 6.8 1138 2 10.4 9.6 8.9 1 1626 8.1 6.8 5.8 2001 0 6.9 6.3 6.0 131 671 5.6 4.5 3.9 0 3057 7.9 7.0 6.3 182 779 8.3 7.5 6.9 0 2386 6.4 5.4 4.5 1 3142 10 sites, 44 more on CD-ROM 4.8 3.7 3.0 0 3601 7.8 6.0 5.1 0 3552 6.1 5.3 4.8 0 3553 7.3 6.4 5.6 0 3560 8.3 7.2 6.2 0 3533 8.4 6.9 5.8 0 3727 13.3 9.8 7.8 0 3548 8.4 7.2 6.3 0 3781 5.5 4.7 3.8 0 3392 5.4 4.8 4.2 0 3629 13 sites, 63 more on CD-ROM 11.4 9.5 8.3 3946 52 10.2 8.7 7.5 3891 34 9.6 8.3 7.3 3590 58 8.4 7.3 6.4 3669 107 9.6 8.3 7.3 3626 129 9.0 7.9 7.0 3694 127
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station LUBLIN RADAWIEC POZNAN LAWICA RACIBORZ SZCZECIN TERESPOL WARSZAWA OKECIE WROCLAW STRACHOWICE
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 29.1 20.2 27.2 19.6 25.4 18.6 30.2 19.4 28.3 18.6 26.7 18.0 30.0 20.2 28.0 19.4 26.2 18.6 29.1 20.0 27.1 19.1 25.3 18.4 29.8 20.4 27.8 19.7 26.1 18.6 30.0 20.3 28.0 19.3 26.2 18.3 30.3 20.0 28.4 19.2 26.7 18.5
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 21.4 27.1 20.4 25.6 20.9 27.6 20.0 26.1 21.2 27.3 20.3 26.1 21.4 26.9 20.3 25.1 21.7 27.6 20.6 25.9 21.4 27.6 20.5 26.0 21.2 27.8 20.2 26.2
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 19.4 14.6 24.4 18.5 13.7 22.9 18.9 13.8 22.6 18.0 13.0 22.0 19.2 14.3 23.9 18.3 13.5 22.8 19.5 14.3 23.7 18.6 13.4 22.5 19.7 14.7 24.1 18.8 13.8 23.1 19.4 14.3 23.6 18.6 13.6 22.9 19.0 14.0 23.6 18.1 13.2 22.5
Lat
Long
Elev
51.217N 52.421N 50.050N 53.400N 52.067N 52.166N 51.103N
22.400E 16.826E 18.200E 14.617E 23.617E 20.967E 16.886E
240 94 206 7 137 110 123
99.6% -16.6 -13.6 -15.1 -12.0 -18.3 -15.8 -13.9
99% -13.3 -10.6 -12.0 -9.0 -14.6 -12.4 -10.7
38.767N 9.133W
105
4.6
5.8
33.8
20.4
31.7
19.7
29.7
19.2
21.5
30.6
20.8
28.5
19.2
14.2
22.6
18.6
13.6
22.3
18.255N 65.641W 18.433N 66.011W
10 2
20.2 20.9
21.1 21.4
32.2 33.1
25.8 25.2
31.9 32.1
25.6 25.4
31.2 31.6
25.4 25.4
27.4 27.1
30.5 30.3
27.0 26.7
30.1 30.1
26.2 26.2
21.7 21.6
29.1 28.8
26.1 25.7
21.4 21.0
29.0 28.6
25.261N 51.565E
11
11.8
13.0
44.0
22.2
42.9
22.4
41.8
22.8
31.1
35.2
30.6
34.9
30.2
27.5
34.0
29.8
26.8
33.9
44.503N 44.571N 46.785N 44.217N 44.318N 47.178N 44.362N 45.810N
26.102E 26.085E 23.686E 28.650E 23.889E 27.621E 28.488E 21.338E
91 96 316 14 191 121 108 106
-13.1 -13.9 -14.6 -9.0 -12.4 -15.8 -11.0 -11.6
-10.2 -10.9 -11.8 -6.9 -9.7 -12.6 -8.9 -9.0
34.2 34.0 31.0 30.2 34.1 33.1 33.0 34.1
21.0 21.4 20.3 23.5 21.3 21.0 21.6 21.0
32.7 32.2 29.2 28.9 32.3 31.2 31.1 32.3
20.6 21.4 19.7 22.8 20.9 20.4 21.0 20.8
31.0 30.9 27.7 27.8 30.7 29.7 29.8 30.8
20.1 20.6 19.0 22.3 20.5 19.9 20.7 20.1
22.8 23.2 21.7 25.5 23.5 22.8 24.0 22.8
30.1 30.2 28.4 28.2 30.3 29.3 28.2 29.8
21.9 22.3 20.7 24.2 22.5 21.8 23.0 21.9
29.2 29.5 27.0 27.3 29.2 28.3 27.6 28.7
21.0 21.2 19.7 24.7 21.5 20.8 23.0 21.0
15.8 16.0 15.0 19.8 16.6 15.7 17.9 15.9
24.7 25.0 24.3 27.3 26.1 25.6 25.5 24.7
20.0 20.1 18.7 23.3 20.4 19.8 21.9 20.0
14.8 15.0 14.1 18.1 15.5 14.7 16.8 14.9
23.6 24.3 23.0 26.3 25.1 24.4 24.8 23.9
64.600N 40.717E 46.283N 47.983E 55.306N 61.503E 53.364N 83.539E 57.915N 56.021E 70.450N 59.083E 53.214N 34.176E 59.267N 38.017E 51.814N 39.230E 52.026N 113.306E 48.783N 44.346E 52.268N 104.389E 56.833N 53.450E 54.550N 36.367E 55.606N 49.279E 55.270N 86.107E 48.528N 135.188E 54.890N 20.593E 58.567N 49.567E 56.743N 60.803E 45.035N 39.171E 56.000N 92.883E 56.067N 92.733E 55.467N 65.400E 51.767N 36.167E 53.505N 50.164E 53.350N 59.083E 43.000N 47.500E 55.833N 37.617E 55.973N 37.415E 55.592N 37.261E 68.782N 32.751E 56.217N 43.817E 57.883N 60.067E
19 -21 234 255 123 12 202 114 157 693 147 511 159 199 126 263 75 13 158 233 36 277 235 74 247 145 384 -19 157 190 209 81 82 260
-32.8 -18.9 -29.1 -33.8 -30.8 -32.6 -22.1 -29.6 -24.2 -37.3 -22.4 -35.6 -29.3 -25.3 -27.8 -33.9 -30.1 -16.4 -29.0 -30.7 -14.7 -33.7 -36.6 -32.5 -22.1 -26.8 -29.6 -12.6 -22.2 -24.0 -23.2 -32.2 -26.8 -31.2
-29.0 -15.6 -26.2 -30.6 -27.2 -30.6 -18.8 -26.3 -21.0 -35.2 -19.7 -32.1 -26.1 -21.8 -24.5 -30.9 -28.2 -13.0 -26.0 -27.8 -11.1 -31.1 -33.8 -29.4 -19.1 -23.5 -27.0 -9.4 -19.0 -20.9 -20.1 -29.0 -23.3 -28.6
27.6 36.2 30.8 30.1 29.9 14.0 29.2 28.3 32.9 31.0 35.3 28.8 29.6 28.4 30.9 28.9 30.3 28.1 29.5 29.7 34.4 28.4 28.9 31.4 30.8 32.2 30.7 31.6 29.7 30.0 29.8 24.2 30.1 28.3
19.4 21.2 19.3 18.8 20.3 11.6 19.8 20.0 19.2 19.1 18.7 17.8 19.9 19.3 19.9 19.0 22.4 19.9 20.4 19.5 22.4 18.3 19.2 19.2 19.6 19.3 18.5 23.1 20.9 19.1 19.4 16.1 20.1 19.4
25.1 34.6 28.8 28.4 27.9 11.8 27.3 26.2 30.6 28.9 33.3 26.9 27.7 26.7 28.8 27.1 28.8 26.1 27.6 27.8 32.8 26.6 27.1 29.4 28.8 30.2 28.8 30.3 27.8 28.0 27.9 21.9 28.2 26.5
18.0 20.8 18.7 18.3 19.2 10.3 18.6 19.0 18.7 18.1 18.5 17.5 18.9 18.6 19.3 18.2 21.7 18.8 19.2 18.6 22.0 17.6 18.5 18.8 18.8 18.8 17.8 23.2 20.1 18.5 18.7 14.9 19.4 18.4
23.0 32.9 26.9 26.9 26.0 9.8 25.7 24.2 28.2 27.0 31.3 25.1 26.0 25.0 26.9 25.2 27.1 24.2 25.8 25.9 31.0 24.8 25.3 27.6 27.0 28.4 27.0 29.1 25.9 26.0 25.9 19.6 26.3 24.9
16.9 20.4 18.0 17.6 18.1 8.7 17.9 17.9 18.0 17.1 18.1 16.7 18.1 18.1 18.4 17.4 20.8 17.8 18.2 17.7 21.3 16.8 17.6 18.2 18.2 18.2 17.2 22.9 19.1 18.0 18.1 13.9 18.6 17.5
20.6 23.6 21.0 20.9 21.3 12.2 20.7 21.2 20.8 20.9 20.7 19.9 21.0 20.9 21.3 20.7 24.1 21.2 21.2 21.1 24.2 20.0 20.7 21.1 20.9 21.4 20.0 25.4 22.0 20.6 20.8 17.1 21.3 20.7
25.7 30.5 27.4 26.9 27.5 13.7 27.0 26.5 27.9 28.1 29.6 25.6 27.4 25.9 27.8 25.9 27.8 25.9 27.6 27.1 31.0 25.7 26.6 28.3 27.6 27.9 27.3 29.3 28.1 26.1 26.7 22.0 27.7 26.5
19.3 22.6 20.0 19.9 20.3 10.3 19.7 20.0 19.9 19.6 19.9 18.9 20.1 19.8 20.4 19.6 23.1 20.0 20.3 20.0 23.2 19.0 19.6 20.2 20.0 20.5 19.1 24.7 20.9 19.6 19.7 15.7 20.4 19.6
23.9 29.8 26.1 25.6 26.4 11.6 25.4 24.6 27.1 26.1 28.9 24.2 26.2 24.6 26.5 24.8 26.7 24.4 26.2 25.7 29.8 24.3 25.1 26.9 26.5 27.1 26.3 28.6 26.2 25.3 25.5 20.3 26.4 25.0
18.8 21.6 18.9 19.0 19.1 11.2 18.6 19.3 18.3 18.6 18.1 18.1 18.8 19.1 19.1 18.9 23.1 19.7 19.1 19.1 22.1 18.1 18.6 18.7 18.7 19.2 17.5 24.2 19.9 18.9 18.9 15.0 19.2 18.6
13.7 16.3 14.1 14.2 14.1 8.3 13.7 14.2 13.4 14.6 13.2 13.8 13.9 14.2 14.1 14.2 18.0 14.4 14.1 14.3 16.8 13.4 13.9 13.7 14.0 14.2 13.1 19.1 14.9 14.0 14.0 10.7 14.1 13.9
23.0 26.4 23.3 23.6 24.5 12.9 23.2 24.0 23.0 23.7 23.0 22.0 23.9 23.7 24.3 23.1 25.6 23.6 23.9 23.5 27.7 22.4 23.5 23.5 23.5 23.9 23.0 28.3 25.0 23.1 23.5 19.2 23.9 23.3
17.4 20.5 17.9 17.9 18.1 9.4 17.6 18.2 17.8 17.3 17.1 17.1 17.9 18.1 18.1 17.8 22.0 18.2 18.2 18.0 21.1 17.0 17.6 17.9 17.9 18.2 16.6 23.3 18.9 17.9 17.9 13.5 18.2 17.6
12.5 15.1 13.2 13.2 13.2 7.4 12.9 13.3 13.0 13.4 12.4 13.0 13.1 13.3 13.2 13.2 16.8 13.2 13.3 13.3 15.8 12.6 13.0 12.9 13.2 13.3 12.3 18.1 14.0 13.1 13.2 9.8 13.2 13.0
21.6 25.5 22.5 22.9 23.1 11.1 22.2 22.4 22.4 22.1 22.7 21.0 22.8 22.4 23.2 22.1 24.9 21.9 22.9 22.5 26.5 21.3 22.2 22.7 22.7 22.8 22.2 27.7 23.9 22.1 22.5 17.7 22.8 22.1
Portugal LISBOA
Puerto Rico JOSE APONTE DE LA TORRE AP LUIS MUNOZ MARIN INTL
Qatar DOHA INTL
Romania BUCURESTI BANEASA BUCURESTI HENRI COANDA INTL CLUJ NAPOCA CONSTANTA CRAIOVA IASI MIHAIL KOGALNICEANU TIMISOARA TRAIAN VUIA
Licensed for single user. © 2017 ASHRAE, Inc.
Russian Federation ARKHANGELSK TALAGI ASTRAKHAN BALANDINO BARNAUL BOLSHOYE SAVINO BOLVANSKIY BRYANSK CHEREPOVETS CHERTOVITSKOYE CHITA KADALA GUMRAK IRKUTSK IZHEVSK KALUGA KAZAN KEMEROVO KHABAROVSK KHRABROVO KIROV KOLTSOVO KRASNODAR KRASNOYARSK KRASNOYARSK MININO KURGAN KURSK KURUMOCH MAGNITOGORSK MAKHACHKALA MOSCOW BIBIREVO MOSCOW SHEREMETYEVO MOSCOW VNUKOVO MURMANSK NIZHNY NOVGOROD STRIGINO NIZHNY TAGIL
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 8.5 7.4 6.5 3871 106 9.6 8.4 7.4 3521 137 10.0 8.6 7.5 3501 122 9.4 8.3 7.4 3463 99 7.4 6.5 5.8 3852 122 10.0 8.8 7.8 3678 139 9.1 7.9 6.9 3422 136 1 site, 33 more on CD-ROM 8.4 7.4 6.5 1032 568 2 sites, 3 more on CD-ROM 8.2 7.4 6.9 0 3200 9.2 8.4 7.9 0 3162 1 site, 1 more on CD-ROM 10.3 9.2 8.2 57 3726 8 sites, 46 more on CD-ROM 8.3 7.0 5.9 2993 410 10.2 8.2 7.1 2952 444 7.6 6.1 5.0 3458 188 11.7 10.0 7.9 2573 476 11.2 8.9 7.7 2856 472 8.9 7.5 6.6 3206 362 11.2 9.5 8.5 2821 439 8.4 7.1 6.0 2839 381 62 sites, 540 more on CD-ROM 8.1 7.0 6.2 6216 47 10.4 9.0 7.9 3357 703 10.3 9.1 8.0 5564 147 10.5 8.9 7.9 5841 164 10.0 8.9 7.9 5740 110 17.1 15.3 13.9 8887 0 8.8 7.7 6.9 4503 128 8.8 7.4 6.3 5514 56 9.9 8.5 7.3 4376 214 9.9 8.5 7.4 7008 102 12.1 10.7 9.6 4071 471 10.0 8.7 7.5 6607 53 9.6 8.2 7.1 5642 124 9.2 7.5 6.6 4847 78 10.9 9.8 9.0 5230 182 10.6 9.4 8.4 6213 114 10.9 9.4 8.4 6059 227 9.4 8.1 7.3 3825 73 6.1 5.4 4.8 5592 126 9.1 8.0 7.1 5873 96 10.3 9.2 8.3 2832 541 10.1 8.4 7.0 6254 68 7.1 6.1 5.3 6166 102 10.5 9.2 8.2 5842 168 8.3 7.4 6.6 4352 205 10.4 9.2 8.2 4948 244 9.9 8.4 7.3 5748 145 10.8 9.3 8.1 2746 582 4.1 3.4 3.1 4629 138 9.2 8.2 7.4 4856 120 9.3 8.3 7.4 4813 128 10.5 9.0 7.8 6660 10 7.4 6.5 5.8 5024 134 7.4 6.4 5.7 6095 62
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Licensed for single user. © 2017 ASHRAE, Inc.
Station NOVOKUZNETSK NOVOSIBIRSK TOLMACHEVO OMSK ORENBURG ORYOL PENZA PSKOV ROSTOV-NA-DONU RYAZAN SARATOV TSENTRALNY SMOLENSK SOCHI INTL ST PETERSBURG PULKOVO STAVROPOL INTL SURGUT TOMSK TRUBCHEVSK TULA TVER TYUMEN UFA ULAN-UDE VELIKIYE LUKI VLADIKAVKAZ VLADIMIR VLADIVOSTOK VORONEZH YELABUGA
Lat
Long
53.817N 86.883E 55.013N 82.651E 54.967N 73.311E 51.796N 55.457E 52.933N 36.000E 53.133N 45.017E 57.817N 28.333E 47.250N 39.817E 54.633N 39.700E 51.565N 46.047E 54.750N 32.067E 43.450N 39.957E 59.800N 30.263E 45.109N 42.113E 61.344N 73.402E 56.500N 84.917E 52.583N 33.767E 54.233N 37.617E 56.883N 35.867E 57.117N 65.433E 54.558N 55.874E 51.808N 107.438E 56.350N 30.620E 43.033N 44.683E 56.117N 40.350E 43.117N 131.933E 51.700N 39.217E 55.767N 52.067E
Elev 308 111 95 118 196 174 45 78 158 152 239 27 24 453 61 139 178 203 137 102 137 515 106 703 172 183 149 91
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 29.3 19.4 27.6 18.6 25.9 18.0 29.8 18.9 28.0 18.2 26.2 17.4 31.0 18.8 29.1 18.2 27.2 17.6 34.7 19.5 32.7 19.0 30.6 18.4 30.4 20.0 28.4 19.4 26.6 18.5 32.0 20.0 29.8 19.5 27.9 18.7 28.7 20.3 26.8 19.2 25.0 18.1 34.9 21.3 32.9 21.0 31.0 20.2 30.3 19.9 28.2 19.1 26.4 18.3 33.2 19.9 31.1 19.3 29.1 18.8 27.9 19.9 26.2 18.9 24.5 18.1 30.8 23.9 29.2 23.4 28.2 23.0 28.2 19.6 26.2 18.6 24.3 17.7 33.6 19.6 31.8 19.2 29.9 18.8 28.9 18.6 27.0 17.7 24.8 17.2 28.6 19.7 26.9 18.8 25.3 18.0 30.0 20.5 28.0 19.6 26.2 19.0 30.0 20.0 28.0 19.1 26.4 18.5 29.5 19.5 27.4 18.9 25.5 18.0 29.5 19.6 27.8 18.8 26.1 18.1 31.3 20.5 29.6 19.8 27.8 19.0 31.5 18.2 29.2 17.6 27.3 16.9 28.0 19.2 26.4 18.7 24.7 18.0 30.9 20.3 29.1 19.9 27.4 19.2 29.5 20.8 27.4 20.0 25.6 19.1 28.2 21.3 26.7 20.5 24.9 19.9 32.4 20.1 30.4 19.5 28.4 18.7 31.0 20.4 28.8 19.7 27.0 18.6
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 20.9 26.7 19.8 25.5 20.6 26.6 19.7 25.3 20.7 27.7 19.7 26.5 21.3 30.0 20.3 29.0 21.2 27.4 20.4 26.5 21.4 28.4 20.6 27.4 21.3 27.0 20.2 25.2 23.3 31.1 22.3 29.5 21.2 27.6 20.2 26.2 21.3 29.2 20.5 28.3 20.8 26.1 19.8 24.8 25.5 28.9 24.6 28.0 20.8 26.3 19.6 24.6 21.5 29.3 20.6 28.0 20.0 25.9 18.9 24.4 21.1 26.2 20.1 25.1 21.9 27.3 20.8 26.1 21.2 27.5 20.2 26.2 20.9 26.9 19.9 25.4 21.1 27.2 20.2 25.7 21.9 28.8 20.9 27.6 20.1 27.7 19.1 26.2 20.8 25.7 19.6 24.4 22.1 28.1 21.2 26.9 22.2 27.5 20.9 26.1 23.2 26.3 22.2 24.6 21.5 29.1 20.6 27.7 21.6 28.5 20.6 27.2
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 18.9 14.3 23.7 17.8 13.3 22.4 18.8 13.8 22.8 17.8 12.9 22.1 18.2 13.3 23.1 17.2 12.5 22.2 18.5 13.6 24.4 17.4 12.6 23.4 19.2 14.3 24.4 18.3 13.5 23.1 19.1 14.2 24.4 18.2 13.4 23.3 19.3 14.1 24.1 18.2 13.1 22.7 21.0 15.8 26.9 19.9 14.8 25.7 19.2 14.2 24.0 18.2 13.3 22.6 18.9 14.0 23.8 18.0 13.2 23.2 19.0 14.2 23.5 18.0 13.3 22.2 24.2 19.2 27.9 23.4 18.3 27.2 18.9 13.7 23.6 17.8 12.8 22.3 19.0 14.6 24.9 18.1 13.7 24.0 17.9 13.0 22.2 16.9 12.1 21.7 19.4 14.3 23.4 18.3 13.4 22.5 20.0 15.0 25.1 19.0 14.1 23.8 19.1 14.2 24.1 18.2 13.4 23.1 18.9 13.9 23.2 17.9 13.1 22.2 19.0 14.0 23.9 18.1 13.1 23.1 19.5 14.5 25.3 18.5 13.6 24.2 17.7 13.5 22.8 16.7 12.7 21.9 19.0 14.0 23.3 17.9 13.0 21.8 20.2 16.2 25.3 19.2 15.2 24.3 20.3 15.3 25.5 19.0 14.0 24.0 22.2 17.2 24.3 21.3 16.3 23.3 19.1 14.1 23.9 18.3 13.4 23.3 19.2 14.2 24.4 18.2 13.3 23.4
99.6% -33.1 -36.1 -33.0 -29.8 -23.6 -27.2 -23.1 -17.6 -24.2 -23.6 -22.1 -1.9 -22.8 -17.9 -40.2 -36.3 -22.8 -24.6 -25.0 -32.1 -31.2 -36.7 -22.5 -14.6 -25.4 -25.2 -23.5 -28.7
99% -30.2 -32.9 -30.2 -26.3 -20.2 -23.9 -19.2 -14.9 -21.1 -20.8 -19.0 -0.4 -19.7 -14.2 -37.5 -33.2 -19.4 -21.3 -21.7 -29.3 -28.0 -34.0 -19.0 -11.6 -22.4 -22.6 -20.4 -25.5
6.5 7.9 3.2 15.8 7.3 16.3 9.1 5.9 1.9
7.8 9.1 5.2 16.8 8.8 17.8 10.9 7.8 3.1
31.2 45.2 45.0 41.0 32.0 45.2 45.2 44.7 41.1
13.2 23.4 20.7 23.7 15.4 24.4 19.1 19.3 18.9
30.8 44.1 44.1 39.9 31.2 44.1 44.2 43.9 39.9
13.2 23.4 19.9 24.2 15.2 24.2 18.8 18.9 18.4
29.9 43.0 43.1 38.8 30.8 43.1 43.2 43.0 38.8
13.3 23.3 19.4 24.6 15.1 24.2 18.5 18.6 18.0
19.9 31.3 22.4 29.9 19.5 28.8 22.3 21.3 21.0
24.2 35.9 40.5 35.0 24.8 38.4 37.2 37.5 36.4
19.3 30.4 21.7 29.1 19.0 28.0 21.2 20.4 20.1
23.7 35.6 40.2 34.4 24.1 37.7 37.7 37.7 35.4
18.8 30.2 18.8 28.8 18.1 26.8 17.8 17.9 15.2
17.7 27.7 14.8 25.4 16.8 23.0 13.8 13.9 11.9
22.1 34.2 24.2 33.6 22.2 35.3 28.3 22.9 27.3
18.0 29.2 17.1 28.0 17.2 25.8 16.1 16.2 14.1
16.8 25.9 13.2 24.1 15.9 21.7 12.4 12.5 11.1
21.9 33.8 23.5 32.7 21.8 34.8 26.4 21.9 27.2
Saudi Arabia ABHA DHAHARAN KING ABDULAZIZ GASSIM PRINCE ABDULAZIZ JEDDAH KING ABDULAZIZ INTL KHAMIS MUSHAIT KING KHALED AB MAKKAH MEDIAN PRINCE ABDULAZIZ INTL RIYADH KING SALMAN AB TABUK
18.240N 26.265N 26.300N 21.680N 18.297N 21.433N 24.553N 24.710N 28.365N
42.657E 2090 50.152E 26 43.767E 648 39.157E 15 42.804E 2066 39.767E 240 39.705E 656 46.725E 635 36.619E 778
Senegal LEOPOLD SEDAR SENGHOR INTL
14.740N 17.490W
26
16.8
17.2
32.2
23.1
31.2
25.0
30.9
25.3
28.0
29.8
27.4
29.4
27.2
23.1
28.9
27.0
22.8
28.7
44.800N 20.467E 44.818N 20.309E
132 102
-8.8 -10.2
-6.7 -8.0
34.3 34.4
21.3 21.2
32.6 32.8
21.0 21.3
30.9 31.0
20.3 20.5
22.7 23.0
31.1 30.9
21.8 22.2
30.1 29.7
20.0 20.7
15.0 15.6
26.7 26.5
19.1 19.8
14.1 14.7
25.1 25.5
1.350N 103.994E
7
23.1
23.8
33.2
26.4
32.9
26.4
32.2
26.3
27.7
30.7
27.5
30.4
27.1
22.8
29.4
26.8
22.5
29.2
48.170N 17.213E
133
-10.8
-8.1
32.3
20.6
30.7
19.9
28.9
19.2
21.6
29.7
20.8
28.5
19.0
14.0
24.6
18.1
13.2
23.8
29.093S 33.965S 30.667S 29.970S 33.036S 26.139S 33.985S 25.733S
1359 46 1287 10 133 1694 69 1322
-5.0 3.9 -0.7 9.1 7.9 0.1 5.2 2.9
-3.7 5.2 0.6 10.3 9.0 2.0 6.8 4.0
33.8 31.6 34.7 30.4 30.6 29.0 29.2 32.3
15.5 19.7 16.0 23.9 20.0 14.9 18.9 17.3
32.4 29.8 33.6 29.4 28.8 27.9 27.4 31.1
15.4 19.2 15.9 23.5 20.5 15.0 19.5 17.2
31.2 28.0 32.5 28.7 27.2 26.9 26.1 30.1
15.4 18.7 15.7 23.3 20.7 15.3 19.8 17.3
19.7 21.3 20.0 25.4 23.8 19.4 22.7 21.0
26.2 27.7 26.8 28.6 27.5 23.8 25.8 27.1
19.1 20.6 19.3 24.8 23.1 18.7 22.0 20.4
25.8 26.7 26.1 27.9 26.3 23.2 25.0 26.4
18.0 19.2 18.3 24.2 22.7 18.0 21.8 19.4
15.3 14.1 15.5 19.1 17.7 16.0 16.6 16.6
21.6 22.9 21.6 27.2 25.8 21.2 24.2 22.7
17.1 18.8 17.6 23.9 22.0 17.2 21.1 18.8
14.4 13.7 14.8 18.8 17.0 15.1 15.8 16.1
21.0 22.6 21.3 26.9 25.0 20.4 23.7 22.5
67
4.5
5.7
25.8
19.1
24.1
18.5
22.8
18.0
20.0
24.0
19.3
22.7
18.6
13.6
21.4
18.1
13.1
20.8
Serbia BEOGRAD BEOGRAD SURCIN
Singapore SINGAPORE CHANGI INTL
Slovakia BRATISLAVA-STEFANIK
South Africa BLOEMFONTEIN INTL CAPE TOWN INTL DE AAR DURBAN INTL EAST LONDON JOHANNESBURG INTL PORT ELIZABETH INTL PRETORIA
26.302E 18.602E 24.017E 30.951E 27.826E 28.246E 25.617E 28.183E
Spain A CORUNA
43.367N 8.417W
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 11.6 9.7 8.3 5942 105 10.1 8.7 7.6 6168 122 10.3 8.8 7.7 6041 163 10.4 9.2 8.2 5088 319 10.0 8.6 7.7 4480 166 10.0 9.2 8.4 4912 188 7.4 6.3 5.5 4536 86 12.4 10.9 9.5 3401 512 7.4 6.2 5.4 4747 157 10.2 8.9 7.8 4488 361 7.1 6.2 5.4 4685 83 7.5 6.5 5.9 1951 509 9.0 7.7 6.8 4753 71 12.5 10.7 9.3 3334 408 10.0 8.8 7.8 7392 83 7.9 6.3 5.0 6388 98 8.6 7.2 5.8 4377 135 7.1 6.1 5.3 4717 139 8.4 7.3 6.5 4870 107 6.4 5.6 5.0 5999 118 10.0 8.5 7.4 5448 161 10.3 8.8 7.5 6882 136 8.3 7.3 6.4 4592 74 5.3 4.3 3.4 3370 267 8.7 7.8 6.9 4982 127 12.9 11.1 9.7 4974 166 7.9 6.9 6.1 4251 277 12.5 9.9 8.6 5319 176 9 sites, 19 more on CD-ROM 9.3 8.3 7.5 480 819 11.0 9.8 8.9 173 3437 8.9 7.9 6.9 412 2993 9.7 8.7 7.8 1 3813 9.6 8.6 7.6 313 1042 5.9 5.0 4.3 1 4830 9.3 8.1 7.2 76 3803 9.3 8.2 7.2 276 3366 10.2 8.5 7.0 657 2133 1 site, 7 more on CD-ROM 9.6 8.7 8.0 0 2420 2 sites, 26 more on CD-ROM 7.4 6.3 5.3 2447 554 10.0 8.5 7.3 2633 456 1 site, 2 more on CD-ROM 7.0 6.2 5.4 0 3605 1 site, 17 more on CD-ROM 10.4 9.1 7.9 2992 297 8 sites, 55 more on CD-ROM 9.0 7.8 6.9 1392 503 13.5 12.2 11.0 874 412 12.3 10.8 9.5 1138 762 11.0 9.9 8.9 141 1102 12.2 10.7 9.5 415 579 9.4 8.4 7.5 1094 268 14.5 12.9 11.4 663 401 5.3 4.5 4.0 589 864 14 sites, 25 more on CD-ROM 10.1 8.9 7.9 1397 122
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station ALICANTE BARCELONA EL PRAT BILBAO GRAN CANARIA MADRID BARAJAS MALAGA MURCIA PALMA DE MALLORCA SEVILLA TORREJON AB VALENCIA VALLADOLID ZARAGOZA AB
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 32.7 21.2 31.2 21.6 30.2 21.8 30.8 23.6 29.6 23.4 28.7 23.1 32.2 20.8 29.8 20.2 27.8 19.6 30.2 19.8 28.5 20.1 27.3 20.6 36.5 18.7 35.2 18.2 33.9 17.9 35.1 20.1 33.0 20.2 31.0 19.9 36.0 21.8 34.7 21.6 33.4 21.6 33.1 22.5 31.8 22.6 30.6 22.5 39.2 22.3 37.9 21.8 36.2 21.1 36.2 19.7 35.0 19.0 33.8 18.4 33.2 20.9 31.9 21.4 30.8 21.6 34.3 18.2 32.7 17.8 30.9 17.4 36.2 21.2 34.7 20.7 32.9 20.1
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 25.3 28.5 24.6 28.2 25.5 29.0 24.7 28.2 22.7 28.6 21.7 26.8 23.8 26.3 23.1 25.8 21.1 33.6 19.9 31.4 24.0 28.3 23.4 27.6 24.7 31.2 24.0 30.2 25.6 29.4 24.8 28.8 24.6 35.5 23.5 33.7 21.8 34.4 20.6 32.3 24.8 29.1 24.1 28.5 19.5 30.8 18.7 29.7 22.7 32.7 21.7 31.1
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 24.2 19.2 27.2 23.4 18.3 26.8 24.2 19.1 27.9 23.4 18.2 27.3 21.1 15.8 24.4 20.1 14.9 23.4 22.9 17.7 25.6 22.1 16.8 25.0 16.6 12.7 25.7 15.7 12.0 24.7 22.7 17.5 26.4 22.0 16.7 26.2 23.1 18.0 26.9 22.3 17.1 26.5 24.3 19.3 27.8 23.8 18.6 27.5 22.0 16.7 26.9 21.0 15.7 26.1 17.0 13.1 28.1 16.1 12.3 26.3 23.3 18.2 27.3 22.8 17.7 27.1 16.0 12.4 22.0 15.2 11.8 21.2 19.9 15.1 25.7 18.9 14.1 25.1
Lat
Long
Elev
38.282N 41.297N 43.301N 27.932N 40.494N 36.675N 38.000N 39.550N 37.418N 40.497N 39.489N 41.633N 41.666N
0.558W 2.078E 2.911W 15.387W 3.567W 4.499W 1.167W 2.733E 5.893W 3.446W 0.482W 4.750W 1.042W
43 4 42 24 610 16 62 8 34 618 69 735 263
99.6% 3.3 1.6 -0.2 13.8 -3.9 4.1 2.5 0.2 2.0 -4.8 0.8 -3.8 -2.2
99% 4.7 2.9 1.1 14.3 -2.4 5.5 3.9 1.8 3.5 -3.1 2.1 -2.6 -0.9
7.181N 79.866E
9
21.0
22.0
33.1
24.6
32.3
25.2
32.0
25.3
27.6
30.8
27.2
30.6
26.7
22.4
30.1
26.2
21.6
29.6
57.717N 55.567N 59.354N 59.897N
12.000E 13.067E 17.942E 17.589E
2 21 14 21
-12.0 -9.7 -15.0 -18.8
-9.5 -7.3 -11.9 -15.4
26.8 26.5 27.0 26.7
18.1 18.9 17.9 18.5
25.2 24.9 25.2 24.9
17.7 18.4 17.1 17.5
23.5 23.2 23.7 23.1
16.8 17.6 16.5 16.7
19.7 20.3 19.5 19.6
24.3 24.3 24.1 24.5
18.8 19.3 18.6 18.5
23.0 22.9 22.8 23.0
18.1 18.8 18.0 17.8
13.0 13.7 13.0 12.8
21.9 21.9 21.1 21.2
17.2 17.9 17.0 16.8
12.2 12.9 12.1 12.0
20.7 21.0 20.3 20.2
46.983N 7.467E 47.483N 8.400E 47.383N 8.567E
567 843 558
-9.8 -10.7 -8.3
-7.7 -8.6 -6.3
29.7 26.3 29.1
19.4 17.4 19.3
27.9 24.5 27.3
19.0 16.7 18.6
26.1 23.0 25.6
18.2 16.2 17.9
20.2 18.4 20.1
27.5 24.0 27.0
19.5 17.7 19.4
26.3 22.9 25.6
17.7 16.4 17.9
13.6 12.9 13.7
22.9 20.6 22.4
17.1 15.7 17.2
13.1 12.4 13.1
22.3 19.8 21.5
36.181N 33.412N 32.600N 35.117N 35.533N
389 616 543 303 7
-2.1 -3.4 0.8 -1.2 4.0
-0.8 -1.8 2.3 0.3 5.4
39.3 39.7 36.2 39.3 32.8
19.9 18.7 19.3 20.9 22.5
37.9 38.1 34.8 37.7 31.8
19.7 18.3 19.5 20.6 23.8
36.3 36.9 33.4 36.4 30.9
19.7 18.1 19.6 20.2 24.2
23.0 21.3 22.8 23.1 26.5
32.7 30.7 31.1 34.1 30.4
22.2 20.6 22.1 22.3 26.0
31.9 30.1 29.9 33.5 29.9
20.0 19.1 20.6 19.6 25.2
15.4 15.0 16.4 14.9 20.4
27.5 23.3 25.1 28.6 29.6
19.2 18.2 20.0 18.7 24.7
14.6 14.1 15.7 14.0 19.7
26.9 22.9 24.8 27.6 29.2
22.783N 120.262E 10 22.000N 120.750E 24 24.818N 120.939E 8 24.833N 121.000E 27 22.633N 120.283E 29 22.577N 120.350E 8 25.150N 121.800E 3 24.428N 118.359E 28 26.160N 119.958E 68 22.700N 120.482E 30 22.672N 120.462E 24 24.186N 120.654E 115 24.265N 120.621E 202 22.950N 120.206E 19 25.033N 121.517E 9 25.069N 121.552E 6 25.078N 121.233E 33 25.056N 121.243E 46 24.250N 120.517E 5
9.9 16.0 9.1 9.1 12.9 12.1 10.5 7.1 4.5 11.1 11.8 8.0 7.9 10.2 9.8 9.8 9.1 8.7 10.3
11.2 16.9 10.2 10.4 14.1 13.6 11.5 8.1 5.6 12.7 13.0 9.2 9.0 11.8 11.0 10.9 10.2 9.8 11.4
33.2 32.6 33.1 33.2 32.7 33.2 33.9 32.9 31.6 34.4 35.0 34.2 32.2 33.5 34.9 35.1 34.8 34.0 32.9
27.2 26.8 27.9 26.7 27.1 26.7 25.9 28.5 27.7 27.3 27.3 27.8 27.2 27.7 26.9 26.5 26.7 28.2 27.4
32.9 32.1 32.6 32.5 32.2 33.0 33.0 32.1 30.9 34.0 34.2 33.9 31.8 33.0 34.1 34.2 33.9 33.2 32.3
27.1 26.7 27.7 26.5 27.1 26.6 25.8 28.1 27.5 27.2 27.1 27.8 27.0 27.6 26.7 26.5 26.9 27.9 27.2
32.1 31.7 32.1 31.9 31.7 32.2 32.2 31.6 30.0 33.2 33.7 33.1 31.1 32.3 33.4 33.8 33.0 32.7 31.8
26.9 26.6 27.5 26.4 26.9 26.4 25.8 27.9 27.1 26.9 26.9 27.5 26.9 27.3 26.4 26.4 26.7 27.7 27.0
28.2 27.9 28.9 27.6 28.1 27.9 27.2 29.5 28.2 28.2 28.3 29.0 28.3 28.7 27.7 28.0 28.5 29.2 28.0
31.4 31.0 32.0 31.6 31.5 31.1 30.9 31.8 30.5 32.8 33.0 33.2 30.9 32.2 33.2 32.6 32.1 32.8 31.9
27.9 27.6 28.3 27.2 27.7 27.6 26.8 28.9 27.7 27.8 27.9 28.4 27.7 28.2 27.3 27.5 27.9 28.5 27.7
31.2 30.7 31.5 31.3 31.2 30.8 30.7 31.4 29.9 32.3 32.5 32.7 30.5 31.7 32.6 31.9 31.5 32.3 31.5
27.2 27.1 28.0 26.5 27.2 27.1 26.2 29.0 27.6 27.1 27.1 27.9 27.6 27.8 26.2 27.0 27.3 28.2 26.9
23.0 22.9 24.2 22.1 23.0 22.9 21.7 25.6 23.8 22.9 23.0 24.3 24.2 23.9 21.7 22.6 23.1 24.5 22.6
30.2 29.7 31.5 30.1 30.4 30.0 29.2 31.3 30.3 30.6 30.9 32.4 30.6 31.4 30.8 30.3 30.4 32.2 31.1
27.0 26.7 27.2 26.0 26.7 26.8 25.8 28.1 27.1 26.7 26.8 27.2 27.0 27.2 25.8 26.2 27.0 27.3 26.5
22.7 22.3 23.0 21.5 22.4 22.5 21.0 24.4 23.0 22.3 22.5 23.2 23.2 23.0 21.1 21.7 22.8 23.3 22.1
30.1 29.6 30.9 29.8 30.2 29.8 29.1 30.7 29.8 30.3 30.5 31.7 30.2 30.8 30.4 29.8 30.2 31.6 30.9
38.543N 68.825E
-9.8
-6.1
37.9
18.9
36.8
18.7
35.3
18.4
22.6
33.5
21.3
32.4
18.9
15.1
29.2
17.5
13.8
27.7
Sri Lanka BANDARANAIKE INTL
Sweden GOTEBORG MALMO STOCKHOLMN BROMMA UPPSALA
Switzerland
Licensed for single user. © 2017 ASHRAE, Inc.
BERN-ZOLLIKOFEN LAEGEREN ZUERICH-FLUNTERN
Syrian Arab Republic ALEPPO INTL DAMASCUS INTL DARAA HAMA LATAKIA
37.224E 36.516E 36.100E 36.750E 35.767E
Taiwan GANGSHAN HENGCHUN HSINCHU HSINCHU CITY KAOHSIUNG KAOHSIUNG INTL KEELUNG KINMEN MATSU NANGAN PINGTUNG NORTH PINGTUNG SOUTH TAICHUNG TAICHUNG AP TAINAN TAIPEI TAIPEI SONGSHAN TAIWAN TAOYUAN INTL TAOYUAN WU-CHI OBSERVATORY
Tajikistan DUSHANBE
785
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 10.3 9.0 7.8 898 885 9.9 8.5 7.6 1302 646 9.9 8.3 7.2 1522 357 14.8 14.0 13.2 63 1092 9.6 8.4 7.3 1971 657 10.2 9.1 8.0 809 887 7.9 6.8 5.9 884 1120 10.2 8.9 7.8 1263 711 8.8 7.7 6.8 848 1232 9.8 8.4 7.3 2099 594 10.7 9.2 7.8 1088 834 7.9 6.6 5.6 2395 362 13.3 11.9 10.7 1703 709 1 site, 0 more on CD-ROM 8.4 7.8 7.1 0 3405 4 sites, 211 more on CD-ROM 8.2 7.1 6.2 3619 63 8.9 7.8 6.9 3506 47 8.6 7.6 6.8 4149 55 9.7 8.6 7.5 4477 31 3 sites, 62 more on CD-ROM 7.4 5.8 4.9 3349 128 11.4 9.9 8.7 3889 71 8.6 6.9 5.5 3198 149 5 sites, 7 more on CD-ROM 10.5 9.5 8.5 1496 1395 12.4 10.8 9.7 1438 1169 9.0 7.6 6.4 1139 1087 7.2 5.8 4.7 1281 1415 9.9 8.1 6.7 712 1202 19 sites, 19 more on CD-ROM 8.7 7.3 6.4 79 2262 11.1 9.7 8.5 6 2634 13.5 12.0 10.8 276 1846 10.3 9.2 8.2 267 1812 6.4 5.5 5.0 31 2542 8.4 7.1 6.2 32 2610 8.8 7.6 6.7 237 1842 9.3 8.3 7.4 497 1592 15.2 13.6 12.0 997 1197 7.4 6.1 5.2 40 2511 7.4 6.2 5.3 32 2623 9.2 8.0 7.1 181 2080 12.4 10.5 9.4 312 1652 9.4 8.2 7.3 73 2375 7.5 6.7 6.1 219 2088 9.0 7.9 7.2 218 2114 13.0 11.9 11.0 264 1941 12.1 10.7 9.8 330 1807 14.1 12.7 11.5 193 1986 1 site, 7 more on CD-ROM 6.5 5.4 4.5 1943 927
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station
Lat
Long
Elev
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB 99.6%
99%
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Cooling DB/MCWB Evaporation WB/MCDB 0.4% 1% 2% 0.4% 1% DB / MCWB DB / MCWB DB / MCWB WB / MCDB WB / MCDB
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB
Tanzania, United Republic of DAR ES SALAAM INTL
6.878S 39.203E
56
18.1
18.9
33.2
25.6
32.8
25.6
32.2
25.4
26.8
30.7
26.6
30.3
26.1
21.6
28.4
25.7
21.1
28.1
13.913N 100.607E 7.883N 98.400E
3 4
19.6 23.8
21.0 24.2
37.2 34.4
26.3 25.9
36.4 33.8
26.4 25.8
35.8 33.3
26.4 25.8
29.6 27.5
33.8 31.9
29.0 27.2
33.2 31.5
28.7 26.4
25.1 21.9
32.0 29.4
28.1 26.1
24.2 21.5
31.4 29.2
6.166N
1.255E
22
21.6
22.2
33.2
26.1
32.9
26.4
32.2
26.4
28.2
30.9
28.0
30.5
27.8
23.9
29.5
27.2
23.0
29.3
36.851N 10.227E
7
5.1
6.3
37.8
22.5
35.8
22.5
34.0
22.3
25.7
31.4
25.0
30.6
24.2
19.1
28.2
23.3
18.1
27.7
35.417E 35.280E 32.689E 30.800E 29.067E 40.201E 41.170E 32.995E 30.582E 37.479E 28.821E 27.157E 27.010E 35.495E 32.562E 38.091E 36.300E 43.332E
66 20 809 54 100 686 1757 953 787 706 50 126 5 1056 1031 862 4 1670
0.7 1.1 -11.0 1.9 -3.2 -9.2 -29.1 -14.9 -10.7 -4.8 -1.8 -2.5 -1.8 -16.0 -12.0 -11.4 -0.9 -13.2
2.0 2.9 -8.6 3.1 -1.9 -6.1 -26.1 -11.2 -8.2 -3.1 -0.2 -1.0 -0.2 -12.4 -9.2 -8.8 0.3 -11.3
36.5 36.8 34.9 38.3 34.4 40.2 30.4 33.9 33.2 39.1 31.9 37.1 36.8 34.4 34.2 38.0 28.6 29.0
22.2 22.2 17.8 20.1 22.0 19.2 14.9 17.2 19.2 21.1 21.4 20.7 21.4 17.3 16.5 19.4 22.3 18.6
35.1 35.1 33.0 36.8 33.0 39.1 29.0 32.0 31.8 37.8 30.6 35.8 35.2 32.8 32.8 36.8 27.7 27.8
22.5 23.0 17.7 20.1 21.8 19.1 14.7 16.9 18.8 20.6 21.2 20.4 21.2 16.9 16.4 18.6 22.1 18.6
34.0 34.1 31.2 35.0 31.7 38.0 27.2 30.2 30.1 36.3 29.2 34.5 34.1 31.0 31.1 35.2 27.0 26.9
22.8 23.3 17.1 20.3 21.3 19.0 14.5 16.5 18.2 20.1 20.9 20.1 20.9 16.4 16.0 18.3 21.8 18.4
26.5 26.5 20.2 26.4 23.8 22.9 17.2 19.2 21.6 23.5 24.5 22.6 23.6 19.1 18.5 22.3 23.9 21.8
31.7 32.0 30.2 30.7 31.5 36.2 26.5 29.3 30.1 36.2 27.8 32.9 32.6 29.7 29.6 34.6 27.2 26.7
25.9 26.0 19.2 25.8 23.0 21.7 16.3 18.2 20.6 22.5 23.6 21.8 22.7 18.2 17.5 20.9 23.2 20.6
30.9 31.3 28.9 30.2 30.6 35.4 25.6 28.5 28.5 35.3 27.3 32.0 31.9 29.0 29.1 33.0 26.6 26.1
25.1 25.1 16.8 25.1 21.2 18.1 13.8 15.2 18.9 19.1 23.3 19.3 20.9 15.5 14.6 17.8 22.8 20.1
20.4 20.2 13.2 20.3 16.1 14.2 12.2 12.2 15.1 15.1 18.2 14.3 15.5 12.5 11.8 14.2 17.5 18.2
28.6 28.8 24.6 29.5 28.2 32.1 22.3 23.2 25.7 32.5 26.4 26.3 28.1 22.8 22.5 32.5 26.4 26.0
24.5 24.2 15.7 24.2 20.4 16.7 12.8 14.2 17.8 17.9 22.2 18.2 19.8 14.2 13.2 16.0 22.1 18.7
19.6 19.2 12.3 19.3 15.3 12.9 11.4 11.4 14.1 14.0 17.0 13.3 14.5 11.5 10.7 12.6 16.7 16.6
28.3 28.2 23.7 29.0 27.4 30.2 20.8 22.5 25.0 31.5 25.8 26.0 27.6 22.2 21.2 29.5 26.0 25.4
37.987N 58.361E
211
-7.8
-4.9
40.2
19.5
39.1
19.4
37.9
19.3
23.1
34.6
22.2
33.7
19.0
14.2
29.7
18.0
13.3
29.3
51.467N 48.600N 48.067N 49.925N 46.633N 50.402N 48.033N 48.567N 49.813N 47.033N 46.427N 49.600N 45.052N 49.233N 47.867N
31.250E 34.967E 37.767E 36.290E 32.567E 30.451E 33.217E 39.250E 23.956E 37.500E 30.676E 34.550E 33.975E 28.600E 35.316E
141 143 225 155 54 179 124 62 326 70 52 160 195 298 114
-19.4 -17.7 -18.8 -19.3 -15.5 -17.0 -17.8 -20.6 -17.0 -15.8 -13.2 -18.7 -12.4 -18.8 -17.5
-16.5 -14.9 -15.9 -16.6 -12.6 -14.2 -14.9 -17.2 -13.9 -13.1 -10.2 -15.8 -9.8 -15.3 -14.8
30.7 33.6 33.2 32.8 34.3 31.0 33.2 34.7 29.2 32.0 32.3 31.8 33.5 29.7 34.1
20.2 21.0 19.4 19.7 21.2 20.3 20.5 20.6 20.1 21.3 20.6 20.2 20.3 19.6 20.2
28.9 31.5 31.1 30.6 32.3 29.1 31.2 32.6 27.5 30.3 30.8 29.9 31.8 27.9 32.1
19.6 20.3 19.1 19.0 20.5 19.8 19.8 20.1 19.2 21.1 20.2 19.5 19.9 18.9 19.8
27.1 29.8 29.2 28.9 30.5 27.4 29.6 30.6 25.9 28.9 29.0 28.2 29.9 26.4 30.2
18.8 19.7 18.6 18.6 19.8 18.9 19.2 19.5 18.4 20.7 19.8 18.9 19.3 18.3 19.2
21.5 22.4 21.4 21.2 22.5 21.7 22.0 22.1 21.2 23.7 23.0 21.7 22.3 21.0 22.1
28.3 30.0 28.6 28.4 30.3 28.0 29.3 30.4 27.3 28.7 27.9 28.6 28.4 27.4 29.4
20.6 21.5 20.5 20.4 21.7 20.8 21.0 21.3 20.1 22.7 22.1 20.8 21.3 20.1 21.2
26.9 29.0 27.5 27.3 29.1 26.9 28.3 29.6 25.6 27.8 27.1 27.5 27.7 26.1 28.6
19.3 20.0 19.1 18.9 20.3 19.8 19.6 19.6 19.0 22.1 21.6 19.4 20.6 18.8 20.0
14.3 15.0 14.2 14.0 15.1 14.8 14.6 14.4 14.4 16.9 16.3 14.4 15.6 14.1 14.9
24.4 25.7 23.8 23.7 24.7 24.4 24.8 25.0 24.0 26.4 25.3 24.8 24.6 23.8 24.7
18.4 19.1 18.2 18.1 19.4 18.9 18.7 18.7 18.1 21.0 20.5 18.5 19.3 18.0 19.0
13.5 14.2 13.4 13.3 14.2 14.0 13.7 13.6 13.5 15.8 15.2 13.6 14.4 13.4 14.0
23.4 24.5 22.9 23.3 24.2 23.4 23.9 24.2 22.8 25.8 24.5 23.9 23.4 22.7 23.7
24.433N 24.262N 24.428N 25.255N 25.329N
54.651E 55.609E 54.458E 55.364E 55.517E
27 265 5 10 34
11.8 10.9 13.6 13.1 10.2
13.0 12.0 14.8 14.2 11.7
44.9 46.0 44.1 43.1 44.2
23.0 22.8 23.3 23.5 23.7
43.6 45.1 42.8 41.8 43.1
23.3 22.8 23.6 23.8 24.0
42.1 44.1 41.2 40.5 42.0
23.4 22.7 24.0 24.2 24.3
30.5 29.0 30.9 30.4 30.1
35.3 35.9 34.8 35.1 36.6
30.0 28.1 30.3 29.8 29.4
34.8 35.8 34.5 34.7 36.0
29.2 27.6 30.0 29.2 28.8
26.1 24.4 27.2 25.9 25.4
33.5 32.5 33.6 33.4 33.5
28.9 26.2 29.2 28.8 28.0
25.6 22.3 25.9 25.3 24.2
33.3 32.3 33.4 33.4 33.1
53.550N 53.817N 52.454N 51.383N 51.467N
2.917W 1.867W 1.748W 2.719W 2.600W
56 267 100 190 11
-2.9 -3.9 -4.9 -3.2 -1.9
-1.5 -2.7 -3.1 -2.1 -0.5
24.5 23.7 26.2 24.9 26.7
17.4 17.3 17.8 17.6 18.3
22.4 21.7 24.2 22.9 24.8
16.7 16.2 17.0 16.8 17.2
20.5 19.9 22.7 21.1 23.0
15.9 15.3 16.3 16.2 16.6
18.3 18.1 18.8 18.6 19.3
23.0 21.8 24.2 22.9 24.3
17.4 17.0 18.0 17.6 18.2
21.2 20.4 22.6 21.2 22.6
16.7 16.7 17.0 17.1 17.5
12.0 12.3 12.3 12.5 12.5
19.4 19.3 20.5 19.6 20.8
15.8 15.7 16.1 16.2 16.6
11.3 11.5 11.6 11.8 11.8
18.7 18.2 19.6 18.5 19.9
Thailand BANGKOK INTL PHUKET
Togo LOME-TOKOIN
Tunisia TUNIS-CARTHAGE INTL
Licensed for single user. © 2017 ASHRAE, Inc.
Turkey ADANA INCIRLIK ADANA SAKIRPASA ANKARA ETIMESGUT ANTALYA BURSA DIYARBAKIR ERZURUM ESENBOGA ESKISEHIR GAZIANTEP OGUZELI ISTANBUL ATATURK IZMIR ADNAN MENDERES IZMIR CIGLI KAYSERI KONYA MALATYA ERHAC SAMSUN VAN
37.000N 36.982N 39.950N 36.899N 40.183N 37.894N 39.957N 40.128N 39.784N 36.947N 40.977N 38.292N 38.513N 38.770N 37.979N 38.435N 41.283N 38.468N
Turkmenistan ASHGABAT
Ukraine CHERNIHIV DNEPROPETROVSK DONETSK KHARKIV INTL KHERSON KIEV ZHULIANY INTL KRYVYI RIH LUHANSK LVIV INTL MARIUPOL' ODESA INTL POLTAVA SIMFEROPOL INTL VINNYTSIA ZAPORIZHIA INTL
United Arab Emirates ABU DHABI INTL AL AIN INTL BATEEN DUBAI INTL SHARJAH INTL
United Kingdom AUGHTON BINGLEY NO2 BIRMINGHAM BRISTOL BRISTOL WEATHER CENTRE
(Click here to access additional station tables on CD-ROM.)
Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 1 site, 3 more on CD-ROM 9.2 8.4 7.8 0 2955 2 sites, 72 more on CD-ROM 7.7 6.7 5.9 0 3997 4.5 4.2 3.8 0 3807 1 site, 1 more on CD-ROM 8.3 7.5 7.0 0 3421 1 site, 14 more on CD-ROM 11.3 10.1 8.9 751 1264 18 sites, 47 more on CD-ROM 8.7 7.6 6.6 1033 1346 7.9 6.8 6.0 919 1481 8.4 7.3 6.2 2801 444 10.6 9.1 7.8 995 1289 7.3 6.2 5.3 1914 667 9.1 7.9 6.8 2139 1214 10.1 9.1 8.1 4941 81 8.8 7.6 6.7 3169 292 8.7 7.8 6.9 2847 351 8.4 7.4 6.4 1913 1179 11.2 10.0 9.0 1811 714 11.7 10.6 9.9 1543 1032 10.0 9.0 8.1 1352 1039 9.1 7.4 5.9 3072 298 11.0 9.4 8.4 2804 499 10.0 8.6 7.4 2597 827 7.5 6.2 5.2 1888 444 8.6 7.2 5.8 3482 221 1 site, 20 more on CD-ROM 9.2 8.1 7.1 1855 1509 15 sites, 33 more on CD-ROM 8.5 7.5 6.8 4023 197 10.9 9.5 8.6 3627 388 11.5 10.0 8.7 3784 334 9.4 8.3 7.6 3898 306 9.1 7.7 6.5 3221 441 8.5 7.3 6.4 3757 239 11.0 9.7 8.3 3602 341 7.3 6.3 5.4 3682 370 9.3 8.0 7.1 3817 118 12.9 11.3 9.8 3461 415 10.5 9.1 8.1 3113 413 9.1 7.7 6.6 3847 283 12.7 11.0 9.7 2952 399 10.3 8.8 7.6 3940 157 9.5 8.4 7.6 3524 399 5 sites, 3 more on CD-ROM 9.4 8.4 7.6 27 3676 10.3 9.1 8.1 41 3967 9.2 8.1 7.2 13 3719 9.0 8.0 7.3 16 3695 8.3 7.2 6.4 42 3492 25 sites, 223 more on CD-ROM 11.5 10.2 9.1 3193 18 12.2 10.5 9.3 3577 9 9.8 8.7 7.9 3102 31 11.9 10.5 9.5 3060 21 10.2 8.8 7.7 2594 59
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017 (Click here to access additional station tables on CD-ROM.) Meaning of acronyms: DB: Dry bulb temperature, °C WB: Wet bulb temperature, °C MCWB: Mean coincident wet bulb temperature, °C
Station CARDIFF WEATHER CENTRE CHURCH LAWFORD CILFYNYDD CROSBY EDINBURGH EMLEY MOOR GLASGOW GRAVESEND-BROADNESS HAWARDEN KENLEY LECONFIELD LEEDS BRADFORD LEEDS WEATHER CENTRE LIVERPOOL JOHN LENNON LONDON HEATHROW LONDON WC CLERKENWELL MANCHESTER NORTHOLT NOTTINGHAM EAST MIDLANDS VALLEY ANGLESEY
Lat: Latitude, ° DP: Dew point temperature, °C MCDB: Mean coincident dry bulb temperature, °C
Heating DB
Cooling DB/MCWB 0.4% 1% 2% DB / MCWB DB / MCWB DB / MCWB 26.2 18.2 24.3 17.4 22.7 16.6 26.4 18.6 24.3 17.5 22.6 16.6 25.5 18.0 23.4 16.8 21.5 16.1 24.4 18.2 22.2 17.4 20.6 16.7 22.1 16.6 20.7 16.0 19.2 15.1 23.7 17.5 21.7 16.5 20.1 15.7 23.1 17.1 21.2 16.2 19.8 15.5 27.7 19.6 25.7 18.6 23.9 17.7 25.1 18.2 23.1 17.4 21.5 16.6 26.2 18.0 24.3 17.2 22.6 16.4 24.9 18.2 23.1 17.3 21.6 16.5 23.9 17.8 21.9 16.6 20.1 15.7 26.2 17.9 24.1 17.0 22.4 16.1 25.0 17.7 23.0 16.9 21.2 16.3 28.3 18.6 26.2 17.7 24.5 17.0 28.2 18.3 26.2 17.6 24.4 16.8 25.5 17.8 23.4 16.9 21.8 16.2 28.2 18.5 26.1 17.7 24.3 17.0 26.2 18.0 24.1 17.0 22.2 16.2 22.9 17.3 20.8 16.1 19.1 15.4
Long: Longitude, ° Elev: Elevation, m HR: Humidity ratio, g of moisture per kg of dry air WS: Wind speed, m/s HDD and CDD 18.3: Annual heating and cooling degree-days, base 18.3°C, °C-day
Evaporation WB/MCDB 0.4% 1% WB / MCDB WB / MCDB 19.2 24.5 18.2 22.6 19.4 24.3 18.3 22.8 18.8 23.8 17.6 21.7 19.2 22.8 18.2 21.0 17.6 20.9 16.7 19.5 18.3 22.1 17.2 20.7 18.1 21.7 17.1 20.0 20.6 26.0 19.5 24.0 19.2 23.3 18.2 22.0 19.0 24.2 18.1 22.6 19.0 23.2 18.1 21.9 18.4 22.2 17.3 20.8 18.7 24.3 17.8 22.8 18.6 23.3 17.7 21.7 19.6 26.0 18.7 24.2 19.4 25.6 18.6 24.1 18.6 23.5 17.8 22.1 19.6 25.8 18.6 24.0 19.0 24.4 18.0 22.5 17.8 21.3 16.9 19.3
Dehumidification DP/HR/MCDB 0.4% 1% DP / HR / MCDB DP / HR / MCDB 17.3 12.5 20.8 16.6 11.9 19.9 17.7 12.8 20.7 16.6 12.0 19.8 17.1 12.5 20.1 16.2 11.8 18.8 18.0 12.9 20.3 17.0 12.2 19.4 16.2 11.6 18.9 15.4 11.0 18.1 16.8 12.3 19.9 15.8 11.6 18.7 16.8 12.0 19.7 15.9 11.3 18.9 18.6 13.5 22.6 17.6 12.6 21.6 17.8 12.7 20.8 16.7 11.9 19.8 17.2 12.5 20.5 16.3 11.9 19.7 17.4 12.5 20.6 16.5 11.8 19.5 17.0 12.4 20.1 16.0 11.6 18.7 16.7 12.0 20.8 15.8 11.3 19.7 16.9 12.1 20.4 16.0 11.4 19.7 17.4 12.5 21.4 16.6 11.9 20.7 17.2 12.3 21.6 16.4 11.7 21.0 17.0 12.2 20.4 16.1 11.5 19.3 17.4 12.5 21.4 16.6 11.8 20.6 17.1 12.3 20.8 16.1 11.6 19.8 16.5 11.8 18.5 15.9 11.3 17.6
Lat
Long
Elev
51.483N 52.367N 51.633N 53.500N 55.950N 53.617N 55.872N 51.467N 53.178N 51.300N 53.867N 53.866N 53.800N 53.334N 51.478N 51.522N 53.354N 51.553N 52.831N 53.248N
3.183W 1.333W 3.300W 3.067W 3.373W 1.667W 4.433W 0.300E 2.978W 0.083W 0.433W 1.661W 1.550W 2.850W 0.461W 0.112W 2.275W 0.418W 1.328W 4.535W
52 106 194 9 41 259 8 3 14 170 7 208 47 24 25 39 78 38 93 11
99.6% -1.0 -4.3 -4.1 -3.4 -5.2 -3.3 -6.0 -2.4 -4.3 -3.2 -3.7 -3.8 -2.4 -2.2 -2.5 -0.6 -4.0 -3.8 -3.8 -1.6
99% 0.1 -2.8 -2.6 -1.7 -3.2 -2.3 -3.9 -1.2 -2.7 -2.0 -2.1 -2.2 -1.2 -1.1 -1.2 0.4 -2.2 -2.4 -2.1 -0.3
34.838S 56.031W 34.850S 56.200W
32 16
1.1 3.0
2.8 4.3
31.8 31.6
21.9 22.5
30.0 30.1
21.4 22.0
28.2 28.8
21.0 21.7
24.3 24.2
28.4 29.1
23.4 23.5
27.0 27.8
23.1 22.7
17.9 17.5
26.1 26.5
22.2 22.1
16.9 16.8
25.0 25.9
40.983N 71.583E 39.701N 66.984E 41.258N 69.281E
474 678 432
-9.0 -10.7 -10.1
-6.5 -7.5 -7.6
36.9 36.2 38.7
21.2 18.9 19.7
35.6 35.0 37.2
20.8 18.6 19.2
34.4 33.9 36.0
20.4 18.2 18.9
23.1 20.7 22.4
33.1 32.7 33.9
22.2 19.9 21.2
32.7 31.9 33.0
19.7 16.8 18.4
15.2 13.0 14.0
29.7 25.2 28.9
18.5 15.7 17.2
14.2 12.1 12.9
28.8 24.4 26.9
7.841N 72.440W 400 10.603N 66.991W 72
19.9 20.8
20.7 21.6
35.2 34.0
23.4 28.1
34.7 33.1
23.3 27.7
34.0 32.8
23.2 27.6
25.9 29.9
31.3 32.1
25.3 29.2
30.9 31.6
24.7 29.2
20.7 26.2
27.5 31.1
24.0 28.8
19.8 25.5
27.0 30.8
16.044N 108.199E 10 20.800N 106.633E 116 10.819N 106.652E 10 21.221N 105.807E 12
16.7 9.7 20.0 9.9
17.5 10.8 21.1 11.0
36.2 33.9 35.6 36.0
26.1 28.8 25.8 27.4
35.2 33.0 34.8 34.9
26.1 28.6 25.8 27.5
34.2 32.3 34.0 34.0
26.2 28.2 25.7 27.4
28.0 29.9 28.0 29.3
32.3 32.5 31.7 32.7
27.5 29.2 27.6 28.8
31.9 31.9 31.3 32.2
26.9 29.2 27.1 28.8
22.6 26.3 22.9 25.3
30.4 32.0 29.6 31.2
26.4 28.6 26.8 28.0
21.9 25.3 22.5 24.2
29.9 31.2 29.4 30.8
17.932S 31.093E 1490
6.2
7.4
31.1
16.5
30.0
16.4
29.0
16.4
20.2
25.3
19.8
24.7
19.1
16.7
21.4
18.6
16.2
21.1
Uruguay CARRASCO INTL MONTEVIDEO PRADO
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Uzbekistan NAMANGAN SAMARKAND TASHKENT INTL
Venezuela JUAN VICENTE GOMEZ INTL SIMON BOLIVAR INTL
Viet Nam DA NANG INTL HAI PHONG PHU LIEN HO CHI MINH TAN SON NHAT INTL NOI BAI INTL
Zimbabwe HARARE INTL
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Extreme Heat./Cool. Annual WS Degree-Days 1% 2.5% 5% HDD / CDD 18.3 11.7 10.2 9.0 2530 57 9.4 8.2 7.2 3102 30 11.5 9.9 8.7 3277 22 17.2 14.9 13.3 2906 19 12.4 10.8 9.5 3423 4 14.8 12.9 10.9 3494 14 12.7 11.0 9.6 3376 7 11.9 10.1 8.9 2612 74 10.3 9.1 8.1 3016 18 10.4 9.2 8.2 2977 42 12.1 10.5 9.3 3149 17 12.8 11.1 9.7 3434 12 12.9 11.0 9.4 2948 41 13.0 11.3 9.9 2870 28 10.3 9.1 8.1 2578 93 9.3 8.3 7.4 2322 123 11.1 9.8 8.7 3083 27 10.3 9.1 8.2 2747 71 12.4 10.8 9.5 3022 37 17.3 15.0 13.4 2869 10 2 sites, 11 more on CD-ROM 12.2 10.3 9.2 1206 474 10.2 8.6 7.6 1104 573 3 sites, 18 more on CD-ROM 7.2 5.5 4.3 2235 1116 9.5 8.3 7.0 2217 856 6.2 5.3 4.5 2112 1057 2 sites, 6 more on CD-ROM 11.9 10.4 9.5 0 3297 4.2 3.5 3.2 0 3376 4 sites, 25 more on CD-ROM 7.5 6.3 5.4 3 2922 6.5 5.3 4.5 179 2185 9.6 7.7 6.5 0 3598 7.0 6.1 5.3 174 2381 1 site, 1 more on CD-ROM 8.8 7.8 6.9 349 762
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Related Commercial Resources CHAPTER 15
FENESTRATION
Licensed for single user. © 2017 ASHRAE, Inc.
FENESTRATION COMPONENTS .......................................... 15.1 Glazing Units ........................................................................... 15.1 Framing ................................................................................... 15.2 Shading .................................................................................... 15.3 DETERMINING FENESTRATION ENERGY FLOW.............. 15.3 U-FACTOR (THERMAL TRANSMITTANCE) ........................ 15.4 Determining Fenestration U-Factors ...................................... 15.5 Surface and Cavity Heat Transfer Coefficients ....................... 15.6 Representative U-Factors for Doors ..................................... 15.13 SOLAR HEAT GAIN AND VISIBLE TRANSMITTANCE...... 15.14 Solar-Optical Properties of Glazing...................................... 15.14 Solar Heat Gain Coefficient .................................................. 15.19 Calculation of Solar Heat Gain ............................................. 15.32 SHADING AND FENESTRATION ATTACHMENTS ................................................................ 15.33
ENESTRATION is an architectural term that refers to the arrangement, proportion, and design of window, skylight, and door systems in a building. Fenestration can serve as a physical and/or visual connection to the outdoors, as well as a means to admit solar radiation for daylighting and heat gain to a space. Fenestration can be fixed or operable, and operable units can allow natural ventilation to a space and egress in low-rise buildings. Fenestration affects building energy use through four basic mechanisms: thermal heat transfer, solar heat gain, air leakage/ventilation/ exchange, and daylighting. Fenestration can be used to positively influence a building's energy performance by (1) using glazing and framing to minimize conductive heat loss, (2) using glazing and shading strategies to control solar heat gain to supplement heating and minimize cooling requirements, (3) specifying low-air-leakage fenestration products, (4) integrating fenestration into natural ventilation strategies that can reduce energy use for cooling and outdoor air requirements, and (5) using daylight to offset lighting requirements. Today’s designers and builders; minimum energy standards and codes; green building standards, codes, and rating programs; and energy efficiency incentive programs are seeking more from fenestration systems and giving credit for high-performing products. Window, skylight, and door manufacturers are responding with new and improved products to meet these demands. With the widespread use of simulation software, designing to improve thermal performance of fenestration products has become much easier. Through participation in rating and certification programs that require the use of this software, fenestration manufacturers can take credit for these improvements through certified ratings. A designer should consider architectural and code requirements, thermal performance, daylight performance, air leakage, energy and environmental impacts, economic criteria, and human comfort when selecting fenestration. Typically, a wide range of fenestration products is available that meet the specifications for a project. Refining the specifications to improve energy performance and enhance a living or work space can result in lower energy costs, increased productivity, and improved thermal and visual comfort.
F
The preparation of this chapter is assigned to TC 4.5, Fenestration.
1.
15.33 15.34 15.52 15.53 15.54 15.54 15.55 15.57 15.57 15.58 15.60 15.61 15.62 15.62 15.64
FENESTRATION COMPONENTS
Fenestration components include glazing material, either glass or plastic; framing, insulation, mullions, muntin bars, dividers, and opaque door slabs; and indoor and outdoor shading devices such as louvered blinds, drapes, roller shades, lightshelves, metal grills, and awnings. In this chapter, fenestration and fenestration systems refer to the basic assemblies and components of window, skylight, and door systems that are part of the building envelope.
1.1
GLAZING UNITS
Most fenestration currently manufactured using glass contains a glazing system that is packaged in the form of a glazing unit. A glazing unit consists of two or more glazing layers that are held apart by an edge seal. Figure 1 shows the construction of a typical doubleglazing unit. The most common glazing material is glass, although polymer (plastic) is sometimes used, either in the form of intermediate films bonded to glass, or as stand-alone glazing material popular in some skylight products. Both may be clear, tinted, coated, laminated,
Fig. 1 Construction Details of Typical Double-Glazing Unit
15.1 Copyright © 2017, ASHRAE
Shading................................................................................... Fenestration Attachments....................................................... VISUAL AND THERMAL CONTROLS ................................. AIR LEAKAGE ....................................................................... DAYLIGHTING...................................................................... Daylight Prediction ................................................................ Light Transmittance and Daylight Use .................................. SELECTING FENESTRATION.............................................. Annual Energy Performance.................................................. Condensation Resistance ....................................................... Occupant Comfort and Acceptance ....................................... Durability ............................................................................... Supply and Exhaust Airflow Windows ................................... Codes and Standards.............................................................. Symbols ..................................................................................
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Licensed for single user. © 2017 ASHRAE, Inc.
15.2 tempered, patterned, or obscured, and polymer types can be easily shaped, textured, and profiled using several processing options. Clear glazing material transmits more than 75% of the incident solar radiation and more than 85% of the visible light. Body-tinted glass containing a pigment is available in many colors, all of which differ in the amount of solar radiation and visible light they transmit and absorb. Some coated glazing materials are highly reflective (e.g., mirrors), whereas others have very low reflectance. Some spectrally selective glazing products include coatings that have a visible light transmittance more than double their solar transmittance; these are desirable for good daylighting while minimizing cooling loads. Coatings that reduce radiant heat exchange are called low-emissivity (low-e) coatings. Laminated glass is made of two panes of glass adhered together. The interlayer between the two panes of glass is typically plastic and may be clear, tinted, or coated. Tempered glass is designed for safety and shatters into pebble-sized pieces when broken. Patterned glass is a durable ceramic frit applied to a glass surface in a decorative pattern. Obscured glass is translucent and is typically used in privacy applications. Low-e coated glass is energy efficient, improves daylighting potential, and enhances occupant comfort. Thus, it is now used in the vast majority of fenestration products. Low-e coatings are typically applied to one of the protected internal surfaces of the glazing unit (surface #2 or #3 in Figure 1), but some manufacturers now offer double- or triple-glazed products with an additional low-e coating on the exposed room-side surface (surface #4 in Figure 1, or what would be #6 in triple glazing). Low-e coatings can also be applied to thin plastic films for use either as one of the middle layers in glazing units with three or more layers (stretched and held in place by a spacer system), or surface-applied film, where the exposed low-e coating is protected by a very thin, thermal-IR (TIR) transparent protective layer. A wide variety of low-e coatings are used today. Physically, they can be divided into either soft- or hard-coat types. Soft-coat low-e is fabricated using a sputtering process in a vacuum chamber. This is done as a post-processing phase after glass has been produced, cut, and stored. The resulting coating is very fragile, especially to the effects of moisture, so these coatings are normally protected inside the sealed glazing unit. Hard coatings are created using chemical vapor deposition and are applied to the glass while it is still being floated in its final phases of production. This creates a durable lowe coating that can be exposed to moisture and elements. From a performance standpoint, categories include (1) high versus low solar gain, (2) spectrally selective, (3) reflective, (4) absorptive, and (5) high light-to-solar-gain (LSG) ratio. Some of these categories are related and causal. As a rule of thumb, soft coats are often low solar gain, whereas hard coats are usually high solar gain. High-solar-gain coatings, which are more transparent to the whole solar spectrum, are used primarily on south facades in northern heating climates in the northern hemisphere (and, conversely, on north facades in southern heating climates in the southern hemisphere), so solar heat gain during the day can offset thermal heat loss during long, cold winter nights. When this coating is applied to indoor-facing glazing lites (although still facing the glazing cavity), it further increases the solar heat gain coefficient (SHGC), because a higher portion of absorbed solar radiation is transferred to the room (inward-flowing fraction). Their low emissivity helps reduce thermal IR radiation, resulting in increased thermal insulation. Low-solar-gain coatings, which are usually spectrally selective [blocking admission of the near infrared portion of the solar spectrum, also known as near IR (NIR) or solar IR], are typically applied in hot, cooling-dominated regions to reduce solar heat gains, or in buildings where solar heat gains are not desirable (e.g., commercial buildings with large fenestration areas or high internal heat gains). Such coatings are normally applied to the outdoor-facing glazing lite to reduce the inward-flowing fraction of SHGC. The glass surface’s low emissivity also reduces thermal IR radiation heat transfer
2017 ASHRAE Handbook—Fundamentals (SI) and therefore increases thermal resistance of the fenestration product. Although the majority of low-e coatings are applied to the glazing surface facing the glazing cavity, increasing numbers of durable coatings are being applied to the room-facing glazing surface. This further improves thermal performance (U-factor) by substantially reducing thermal IR radiation heat transfer on the room side of the window. However, in building spaces with higher indoor humidity levels in cold climates, this approach may significantly increase the risk of moisture condensation on glass surfaces. A new class of lowe surface-applied films may be applied as a retrofit measure to the indoor-facing glass surface; these films can perform similarly to low-e coated glass, so their use is more likely to be as an easy retrofit measure, removing the need to replace glass or sash. In addition to low-e, fill gases such as argon, krypton, and xenon are used in lieu of air in the gap between glass panes. These fill gases reduce convective heat transfer across the glazing cavity. Krypton and xenon also reduce gap width, because their optimal gap widths are nearly half that of air. The main requirements of the edge seal are to exclude moisture, provide a desiccant for the sealed space, and retain the glazing unit’s structural integrity. Further, the edge seal isolates the cavity between the glazing materials, thereby reducing the number of surfaces to be cleaned, and creating an enclosure suitable for nondurable low-e coatings and/or fill gases. The edge seal is composed of a spacer, single- or multilevel sealant, and desiccant. The spacer separates glazing layers and provides a surface for primary and secondary sealant adhesion. Several types of spacers are used, each of which provides different heat transfer properties, depending on spacer material and geometry. Heat transfer at the edge of the glazing unit is greater than at its center because of (1) heat flow through the spacer system and (2) convective flow, which creates an area of higher convective heat transfer as the gas turns from one glazing to the other. To minimize this heat flow, warm-edge spacers have been developed that reduce edge heat transfer by using materials (e.g., stainless steel, galvanized steel, tin-plated steel, polymers, foamed silicone) of lower thermal conductivity than the typical aluminum alloy from which spacers have often been made. Fusing or bending the corners of the spacer minimizes moisture and hydrocarbon vapor transmission from the gap space. Several different sealant configurations are used in glazing unit construction. In dual-seal construction, a primary seal minimizes moisture transmission and gas escape. A secondary seal provides structural integrity between the lites of the glazing unit, and ensures long-term adhesion and greater resistance to solvents, oils, and short-term water immersion. In typical dual-seal construction, the primary seal is made of compressed polyisobutylene (PIB), and the secondary seal is made of silicone, polysulfide, or polyurethane. Single-seal construction depends on a single sealant to adhere the glazing to the spacer and to minimize moisture transmission and gas escape. Single-seal construction is generally more cost effective than dual-seal systems. Dual-seal-equivalent (DSE) materials take advantage of advanced cross-linking polymers that provide low moisture transmission and structural properties equivalent to dualseal systems. Desiccants are used to absorb moisture trapped in the glazing unit during assembly or that gradually diffuses through seals after construction. Typical desiccants include molecular sieve, silica gel, or a matrix of both materials.
1.2
FRAMING
The three main categories of fenestration framing materials are wood, metal, and polymers. Wood has good structural integrity and insulating value but low resistance to weather, moisture, warpage, and organic degradation (from mold and insects). Metal is durable and has excellent structural characteristics, but it has very poor
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Fenestration
15.3
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 2 Various Framing Configurations for Residential Fenestration thermal performance. The metal of choice in fenestration is almost exclusively aluminum alloy, because of its ease of manufacture, low cost, and low mass, but aluminum alloy has a thermal conductivity roughly 1000 times that of wood or polymers. Steel is sometimes used. Although lower in its thermal conductivity, the overall U-factor of a steel frame is similar to that of an aluminum frame of the same geometry. The poor thermal performance of metal-frame fenestration can be improved with a thermal break (a nonmetal component that separates the metal frame exposed to the outdoors from the surfaces exposed to the indoors). However, to be most effective, there must be thermal breaks in all operable sashes as well as in the frame. Polymer frames are made of extruded vinyl [unplasticized PVC (uPVC)] or pultruded fiberglass (glass-reinforced polyester). Their thermal and structural performance is similar to that of wood. Vinyl frames for large fenestration must be reinforced, which degrades their thermal performance slightly and substantially increases their weight. Polymer frames are generally hollow and thus can also be filled with polyurethane insulation, which reduces convective and radiative heat transfer, thereby achieving a better thermal performance than wood. Manufacturers sometimes combine these materials as clad units (e.g., vinyl-clad aluminum, aluminum-clad wood, vinyl-clad wood) to increase durability, improve thermal performance, or improve aesthetics. In addition, curtain wall systems for commercial buildings may be structurally glazed, and the outdoor “framing” is simply rubber gaskets or silicone. Generally, the framing system categorizes residential fenestration, as shown by the examples of traditional basic types in Figure 2. The glazing system can be mounted either directly in the frame (a direct-glazed or direct-set fenestration, which is not operable) or in a sash that moves in the frame (for an operating fenestration). In operable fenestration, a weather-sealing system between the frame and sash reduces air and water leakage.
1.3
SHADING
Shading devices are available in a wide range of products that differ greatly in their appearance and energy performance. They include louvered shades (e.g., venetian blinds, vertical blinds, louvered shutters), roller shades, solar screens, cellular shades, awnings, roller shutters, draperies (curtains), roman shades, perforated panels, etc. Shades can be positioned as indoor, outdoor, or between-glazing products. When shades are applied after the fenestration unit was installed, a popular retrofit measure, these products are called fenestration attachments. When a fenestration unit is sold with matched shading as a complete assembly, the shade is called integral. Materials used include metal, wood, polymer, woven, and nonwoven fabrics.
The ability of shading devices to control solar gains depends mainly on the location of the device. Shades on the outdoor side of the glazing can effectively reduce solar heat gains, but need more frequent maintenance and are often difficult to adjust. Conversely, indoor devices are ubiquitous and easier to maintain and operate, but may not be as effective in providing significant solar heat gain control, depending on the glazing type, shade properties, and control (Barnaby et al. 2009; Lee and Selkowitz 1995; Moeseke et al. 2007; Shen and Tzempelikos 2012; Tzempelikos and Athienitis 2007; Wright et al. 2009a). Shading devices often do not produce any significant improvement in a fenestration system’s U-factor (Barnaby et al. 2009; Wright et al. 2009b). However, several classes of insulating shades (e.g., roller shutters, cellular shades, window quilts) can provide substantial insulating value and be used as an effective retrofit option to improve thermal resistance, especially if they are sealed tightly around the perimeter, and are not porous to air. Shading devices are well suited to deal with daylighting, privacy, views, glare, and thermal comfort issues. Some products, such as properly adjusted venetian blinds, are quite versatile in this respect. Motorized shading devices can be adjusted under changing outdoor conditions to reduce glare, maximize daylight, reduce internal temperatures (Newsham 1994; Reinhart 2004), or improve thermal comfort to building occupants (Bessoudo et al. 2010) while also providing increased hours of view to the outdoor (Rao and Tzempelikos 2012). Shading of vertical fenestration is not confined to the use of shading devices. Building elements such as window reveals, side fins, and overhangs can also offer effective shading. Metal grilles with fixed louvers mounted horizontally at the top of a fenestration can block solar gain while still letting some light through, thereby avoiding the negative impacts of solid structural overhangs that act as thermal bridges in the building envelope and reduce daylight. Light shelves can provide shading and also redirect sunlight deeper into the room to provide a more even spatial distribution of daylight in a space. Metal perforated panels are a suitable compromise between completely cutting out views and alleviating glare and overillumination issues. Outdoor vegetative shading is particularly effective in reducing solar heat gain while enhancing the outdoor scene, but may be hard to maintain and also not be persistent over multiple years.
2.
DETERMINING FENESTRATION ENERGY FLOW
Energy flows through fenestration via (1) conductive and convective heat transfer caused by the temperature difference between
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15.4
2017 ASHRAE Handbook—Fundamentals (SI)
outdoor and indoor air; (2) net long-wave (above 2500 nm) radiative exchange between the fenestration and its surroundings and between glazing layers; (3) short-wave (below 2500 nm) solar radiation incident on the fenestration product, either directly from the sun or reflected from the ground or adjacent objects; and (4) air leakage through the fenestration. Simplified calculations are based on the observation that temperatures of the sky, ground, and surrounding objects (and hence their radiant emission) correlate with the outdoor air temperature. This is not the case for clear skies, particularly at night, which can be much colder than ambient air. The radiative interchanges are then approximated by assuming that all the radiating surfaces (including the sky) are at the same temperature as the outdoor air. With this assumption, the basic equation for the steadystate energy flow Q through a fenestration is Q = UApf (tout – tin) + (SHGC)Apf Et + (AL)Apf Cp(tout – tin) (1)
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where Q = instantaneous energy flow, W U = overall coefficient of heat transfer (U-factor), W/(m2 ·K) Apf = total projected area of fenestration (product’s rough opening in wall or roof less installation clearances), m2 tin = indoor air temperature, °C tout = outdoor air temperature, °C SHGC = solar heat gain coefficient, dimensionless; also known as g-value in Europe Et = incident total irradiance, W/m2 AL = air leakage at current conditions, m3/(s·m2) = air density, kg/m3 Cp = specific heat of air, kJ/kg·K
Here, the first term on the right-hand side of Equation (1) represents heat transfer resulting from a temperature difference across the fenestration, the second term represents heat transfer caused by solar radiation, and the last term represents heat transfer caused by air leakage. The sections on U-factor (Thermal Transmittance), Solar Heat Gain Coefficient and Visible Transmittance, and Air Leakage discuss these topics, respectively. The main justification for Equation (1) is its simplicity, achieved by collecting all the linked radiative, conductive, and convective energy transfer processes into U and SHGC. Note, however, that the values of U and SHGC vary because (1) convective heat transfer rates vary as fractional powers of temperature differences or freestream speeds, (2) variations in temperature caused by weather or climate are small on the absolute temperature scale (K) that controls radiative heat transfer rates, (3) fenestration systems always involve at least two thermal resistances in series, and (4) solar heat gain coefficients depend on solar incident angle and spectral distribution. Using U and SHGC values taken from product ratings causes some error in calculating Q. Of the two, the larger error results from the use of normal-incidence SHGC.
3. U-FACTOR (THERMAL TRANSMITTANCE) In the absence of sunlight, air infiltration, and moisture condensation, the first term on the right side of Equation (1) represents the heat transfer rate through a fenestration system. Most fenestration systems consist of transparent multipane glazing units and opaque elements comprising the sash and frame (hereafter called frame). The glazing unit’s heat transfer paths are subdivided into center-ofglass, edge-of-glass, and frame contributions (denoted by subscripts cg, eg, and f, respectively). Consequently, the total rate of heat transfer through a fenestration system can be calculated knowing the separate contributions of these three paths. (When present, glazing dividers, such as decorative grilles and muntin bars, also affect heat transfer, and their contribution must be considered.) The overall
U-factor is estimated using area-weighted U-factors for each contribution by Ucg A cg + Ueg A eg + Uf Af U = -----------------------------------------------------------A pf
(2a)
Note that a product frame normally consists of several different areas, and that, for example, Uf Af is really Ufi Afi , where i represents each frame region (e.g., sill, jamb, head, meeting rail). Likewise, each “edge of glazing” region is adjacent to a different frame region, so Ueg Aeg is really Uegi Aegi . Equation (2a) is used in North America [e.g., for fenestration ratings by the National Fenestration Rating Council (NFRC)], Australia [by the Australian Fenestration Rating Council (AFRC; www.afrc.org.au)], India, and other countries. This approach is described as one of the two options in ISO Standard 15099:2003 and is the basis of the calculation used by the Lawrence Berkeley National Lab WINDOW program (LBNL 2016), a widely used software tool for modeling fenestration products’ solar-optical and thermal performance. ASHRAE-derived environmental conditions, as used by the NFRC, are applied when Equation (2a) is used. These environmental conditions are detailed in Table 2. When a fenestration product has glazed surfaces in only one direction, the sum of the areas equals the projected area Apf. Skylights, greenhouse/garden windows, bay/bow windows, etc., because they extend beyond the plane of the wall/roof, have greater surface area for heat loss than fenestration with a similar glazing option and frame material; consequently, U-factors for such products are greater than for the nonprojecting versions. When the slope of a multiglazed fenestration product varies from vertical, the gas in the glazing cavity begins to exhibit vortices that also increase heat flow per unit area. This results in a further U-Factor penalty for products rated at a slope (such as skylights). In Europe and in some other parts of the world, U-factor is calculated differently: instead of the area-based, edge-of-glass term of Equation (2a), a linear thermal transmittance term g is used to describe heat transfer near the perimeter of the glass area. This length-weighted method follows an alternative standard, ISO Standard 10077-2:2012, and is also described as second option in ISO Standard 15099. U = (Ucg Acg + g lg + Uf Af )/Apf
(2b)
where lg is the perimeter length of the entire vision area, delimited by the sightline. ISO Standards 10077-2 and 10077-1:2006 use simpler algorithms than ISO Standard 15099. In many fenestration systems, this leads to a different apparent U-factor, depending on which standard was followed. In a globalized economy, it is essential that designers and specifiers are aware of this lack of harmonization and that straight conversion between these standards is not possible.
Comparison Between Area-Weighted and Length-Weighted Methods Glazing units (double- and triple-pane, many with low-e coatings and noble gas fills) were modeled for U-factor and reviewed by Curcija (2005) and van Dijk (2003). The studies showed that ISO 10077 versus NFRC differences in glazing U-factor can be as great as 0.3 W/(m2·K). These discrepancies are larger than the precision with which U-factors are reported for ratings under the European and NFRC systems. Therefore, for some glazing systems a noticeable difference in U-factor will be seen. In addition to differences between center-of-glass U-factors obtained by the two methods, it is important to note that large differences can occur in frame U-factors; ISO Standard 15099/NFRC frame U-factors can be ~1 W/(m2 ·K) greater than their ISO Standard 10077-2 equivalents. Although such differences are diluted when a whole-system, area-weighted calculation is performed, it
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Fenestration
15.5
cannot be assumed that these differences are negligible in all cases. Use ISO Standard 10077-based U-factors with caution and, for code compliance purposes in the United States, Canada, Australia, and some other countries, seek certifications based on NFRC procedures and environmental conditions. Refer to building codes for detailed requirements concerning NFRC compliance.
3.1
DETERMINING FENESTRATION U-FACTORS
Center-of-Glass U-Factor
Edge-of-Glass U-Factor
For single glazing, U-factors depend strongly on indoor and outdoor film coefficients. The U-factor for single glazing is
The edge-of-glass area is typically taken to be a band 65 mm wide around the sightline. The width of this area is determined from the extent of two-dimensional heat transfer effects in current computer models, which are based on conduction-only analysis. In reality, because of convective and radiative effects, this area may extend beyond 65 mm (Beck et al. 1995; Curcija and Goss 1994; Wright and Sullivan 1995a), and depends on the type of glazing unit and its thickness. The edge region may also extend farther in opaque systems such as spandrel panels. The appropriate edge width to use can be determined using finite-element heat transfer modeling, as described by Curcija and Goss (1994). In low-conductivity frames, most heat flow at the edge-of-glass and frame area is through the spacer, so the type of spacer has a greater impact on the edge-of-glass and frame U-factor. In metal frames, the edge-of-glass and frame U-factor varies little with the type of spacer (metal or insulating) because there is a significant heat flow through the highly conductive frame near the edge-ofglass area.
1 Usingle glazing = ---------------------------------------------1 ho + 1 hi + L k
(3)
where ho, hi = outdoor and indoor respective glazing surface heat transfer coefficients, W/(m2 ·K) L = glazing thickness, m k = thermal conductivity of glass, W/(m2 ·K)
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argon, and krypton fill gases and for high (uncoated) and low (coated) values of surface emissivity. The optimum gas space width is 13 mm for air and argon, and 8 mm for krypton. Greater widths have no significant effect on Ucg. Greater glazing unit thicknesses decrease Uo because the length of the shortest heat flow path through the frame increases. A low-emissivity coating combined with krypton gas fill offers significant potential for reducing heat transfer in narrow-gap-width glazing units.
For other fenestration, values for Ucg at standard indoor and outdoor conditions depend on glazing construction features such as the number of glazing lights, gas space dimensions, orientation relative to vertical, emissivity of each surface, and composition of fill gas. Several computer programs can be used to estimate glazing unit heat transfer for a wide range of glazing construction. The National Fenestration Rating Council (NFRC) calls for LBNL WINDOW 7.4 (LBNL 2016) as a standard calculation method for center of glazing. Heat flow across the central glazed portion of a multipane unit must consider both convective and radiative transfer in the gas space, and may be considered one-dimensional. Convective heat transfer is estimated based on high-aspect-ratio, natural convection correlations for vertical and inclined air layers (El Sherbiny et al. 1982; Shewen 1986; Wright 1996a). Radiative heat transfer (ignoring gas absorption) is quantified using a more fundamental approach. Computational methods solving the combined heat transfer problem have been devised by Hollands and Wright (1982), Rubin (1982a, 1982b), and Rubin et al. (1998). Figure 3 shows the effect of gas space width on Ucg for vertical double- and triple-paned glazing units. U-factors are plotted for air,
Frame U-Factor Fenestration frame elements consist of all structural members exclusive of glazing units and include sash, jamb, head, and sill members; meeting rails and stiles; mullions; and other glazing dividers. Estimating the rate of heat transfer through the frame is complicated by the (1) variety of fenestration products and frame configurations, (2) different combinations of materials used for frames, (3) different sizes available, and, to a lesser extent, (4) glazing unit width and spacer type. Internal dividers or grilles have little effect on the fenestration U-factor, provided there is at least a 3 mm gap between the divider and each glazing.
Fig. 3 Center-of-Glass U-Factor for Vertical Double- and Triple-Pane Glazing Units Gas fill is assumed to be 90% argon, 10% air.
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15.6
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Representative Fenestration Frame U-Factors in W/(m2 ·K), Vertical Orientation Product Type/Number of Glazing Layers
Frame Material
Type of Spacer
Operable 1b
2c
Aluminum without All 13.51 12.89 thermal break Metal 6.81 5.22 Aluminum with Insulated N/A 5.00 thermal breaka Aluminum-clad wood/ Metal 3.41 3.29 Insulated N/A 3.12 reinforced vinyl Wood/vinyl Metal 3.12 2.90 Insulated N/A 2.78 Insulated fiberglass/ Metal 2.10 1.87 Insulated N/A 1.82 vinyl Structural glazing Metal Insulated
Fixed 3d
1b
2c
12.49 10.90 10.22 4.71 4.37 2.90 2.73 2.73 2.27 1.82 1.48
7.49 N/A 3.12 N/A 3.12 N/A 2.10 N/A
6.42 5.91 2.90 2.73 2.73 2.38 1.87 1.82
3d
Plant-Assembled Skylight
1b
1b
2c
2c
3d
Curtain Walle 1f
2g
3h
Sloped/Overhead Glazinge 1f
2g
3h
9.88 10.67 10.39
44.57 39.86
39.01 17.09 16.81
16.07 17.32 17.03
16.30
6.30 5.79 2.73 2.50 2.38 1.99 1.82 1.48
39.46 N/A 27.60 N/A 14.20 N/A
26.01 10.22 9.94 23.39 N/A 9.26 20.78 19.48 10.11 9.71
9.37 10.33 9.99 8.57 N/A 9.31
9.43 8.63
10.22 7.21 N/A 5.79
5.91 10.33 7.27 4.26 N/A 5.79
5.96 4.26
Note: This table should only be used as an estimating tool for early phases of design. aDepends strongly on width of thermal break. Value given is for 9.5 mm. bSingle glazing corresponds to individual glazing unit thickness of 3 mm (nominal). cDouble glazing corresponds to individual glazing unit thickness of 19 mm (nominal). dTriple glazing corresponds to individual glazing unit thickness of 35 mm (nominal).
Licensed for single user. © 2017 ASHRAE, Inc.
Garden Window
Computer simulations show that frame heat loss in most fenestration is controlled by a single component or controlling resistance, and only changes in this component significantly affect frame heat loss (EEL 1990). For example, the frame U-factor for thermally broken aluminum fenestration products is largely controlled by the depth of the thermal break material in the heat flow direction. For aluminum frames without a thermal break, the indoor film coefficient provides most of the resistance to heat flow. For vinyl- and wood-framed fenestrations, the controlling resistance is the shortest distance between the indoor and outdoor surfaces, which usually depends on the thickness of the sealed glazing unit. Carpenter and McGowan (1993) experimentally validated frame U-factors for a variety of fixed and operable fenestration product types, sizes, and materials using computer modeling techniques. Table 1 lists frame U-factors for a variety of frame and spacer materials and glazing unit thicknesses. Frame and edge U-factors are normally determined by two-dimensional, finite-element computer simulation.
Curtain Wall Construction A curtain wall is an exterior building wall that carries no roof or floor loads and consists entirely or principally of glass and other surfacing materials supported by a framework. A curtain wall typically has a metal frame. To improve the thermal performance of standard metal frames, manufacturers provide both traditional thermal breaks as well as thermally improved products. The traditional thermal break is poured and debridged (i.e., urethane is poured into a metal U-channel in the frame and then the bottom of the channel is removed by machine). For this system to work well, there must be a thermal break between indoors and outdoors for all frame components, including those in any operable sash. Skip debridging (incomplete pour and debridging used for increased structural strength) can significantly degrade the U-factor. Bolts that penetrate the thermal break also degrade performance, but to a lesser degree. Griffith et al. (1998) showed that stainless steel bolts spaced 300 mm on center increased the frame U-factor by 18%. The paper also concluded that, in general, the isothermal planes method referenced in Chapter 27 provides a conservative approach to determining U-factors. Thermally improved metal curtain wall products are now being used more widely. In these products, most of the metal frame tends to be located on the indoor side with only a metal cap exposed on the outdoor side. Plastic spacers isolate the glazing assembly from both
5.11 4.83 N/A 4.71
28.67 26.97 22.31 21.29 11.81 11.47
eGlass
thickness in curtainwall and sloped/overhead glazing is 6.4 mm. glazing corresponds to individual glazing unit thickness of 6 mm (nominal). glazing corresponds to individual glazing unit thickness of 25 mm (nominal). hTriple glazing corresponds to individual glazing unit thickness of 44 mm (nominal). N/A: Not applicable fSingle
gDouble
the outdoor metal cap and the indoor metal frame. These products can have significantly better thermal performance than standard metal frames, but it is important to minimize the number and area of the bolts that penetrate from outdoor to indoor. Curtain wall systems using structural silicone glazing (SSG) can also provide better thermal performance than dry-glazed bolted systems. Silicone sealants, with a typical thermal conductivity of 0.35 W/(m·K) and a dimension of 6 mm between the glass and aluminum are considered thermal breaks per NFRC Technical Document 100 (NFRC 2014a). Work by Carbary et al. (2009) demonstrates that SSG systems have lower U-factors than comparable nonbroken or thermally improved dry-glazed systems, resulting in frame temperatures closer to interior ambient conditions. A more recent development is the use of fiberglass as the framing material for curtain walls. Fiberglass provides the strength needed for taller buildings while having a U-factor equal to or lower than that of other nonmetal materials (e.g., wood, vinyl). An important consideration in any curtain wall system is durability. Sealants or gaskets that degrade or fail over time allow additional air infiltration, which negatively affects energy consumption. A durable system minimizes air infiltration and thereby energy consumption.
3.2
SURFACE AND CAVITY HEAT TRANSFER COEFFICIENTS
Part of the overall thermal resistance of a fenestration system derives from convective and radiative heat transfer between the exposed surfaces and the environment, and in the cavity between glazings. Surface heat transfer coefficients ho , hi , and hc at the outer and inner glazing surfaces, and in the cavity, respectively, combine the effects of radiation and convection. Wind speed and building orientation are important in determining ho. This relationship has long been studied, and many correlations have been proposed for ho as a function of wind speed. However, no universal relationship has been accepted, and limited field measurements at low wind speeds by Klems (1989) differ significantly from values used by others. Convective heat transfer coefficients are usually determined at standard temperature and air velocity conditions on each side. Wind speed can vary from less than 0.2 m/s for calm weather, free convection conditions, to over 29 m/s for storm conditions. A nominal value of 26 W/(m2 ·K) corresponding to a 5.5 m/s wind is often
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Fenestration
15.7
Table 2 Indoor Surface Heat Transfer Coefficient hi in W/(m2 ·K), Vertical Orientation (Still Air Conditions) Glazing IDa Glazing Type
Glazing Height, m
1 Single glazing 5
Double glazing with 12.7 mm airspace
23
Double glazing with e = 0.1 on surface 2 and 12.7 mm argon space
43
Triple glazing with e = 0.1 on surfaces 2 and 5 and 12.7 mm argon spaces
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Notes: aGlazing ID refers to fenestration assemblies in Table 4. bWinter conditions: room air temperature t = 21°C, outdoor i air temperature to = 18°C, no solar radiation
0.6 1.2 1.8 0.6 1.2 1.8 0.6 1.2 1.8 0.6 1.2 1.8
Winter Conditionsb
Summer Conditionsc
hi , Glazing Temp. Diff., W/(m2 ·K) Temp., °C °C
hi , Glazing Temp. Diff., Temp., °C °C W/(m2 ·K)
9 9 9 7 7 7 13 13 13 17 17 17
30 30 30 14 14 14 8 8 8 4 4 4
8.04 7.42 7.10 7.72 7.21 6.95 7.44 7.00 6.77 7.09 6.72 6.53
33 33 33 35 35 35 34 34 34 40 40 40
9 9 9 11 11 11 10 10 10 16 16 16
8.10 7.64 7.41 8.30 7.82 7.58 8.20 7.73 7.50 8.73 8.20 7.94
cSummer
conditions: room air temperature ti = 24°C, outdoor air temperature to = 32°C, direct solar irradiance ED = 748 W/m2 hi = hic + hiR = 1.46(T/L)0.25 + (T i4 – T g4 )/T, where T = Ti – Tg, K; L = glazing height, m; Tg = glazing temperature, K; = Stefan-Boltzmann constant; and = surface emissivity.
used to represent winter design conditions. At low wind speeds, ho varies with outdoor air and surface temperature, orientation to vertical, and air moisture content. The overall surface heat transfer coefficient can be as low as 6.8 W/(m2 ·K) (Yazdanian and Klems 1993). For natural convection and radiation at the indoor surface of a vertical fenestration product, surface coefficient hi depends on the indoor air and glazing surface temperatures and on the emissivity of the glazing surface. Table 2 shows the variation of hi for winter (ti = 21C) and summer (ti = 24C) design conditions, for a range of glazing system types and heights. Designers often use hi = 8.3 W/ (m2 ·K), which corresponds to ti = 21C, a glazing temperature of –9.4C, and emissivity of eg = 0.84 (uncoated glass). For summer conditions, the same value [hi = 8.3 W/(m2 ·K)] is normally used, and corresponds approximately to a glazing temperature of 35C, ti = 24C, and eg = 0.84. For winter conditions, this most closely approximates single glazing with clear glass that is 600 mm tall, but it overestimates the value as the glazing unit conductance decreases and height increases. For summer conditions, this value approximates all types of glass that are 600 mm tall but, again, is less accurate as glass height increases. If the indoor surface of the glass has a low-e coating, hi values are about halved at both winter and summer conditions. Heat transfer between the glazing surface and its environment is driven not only by local air temperatures but also by radiant temperatures to which the surface is exposed. The radiant temperature of the indoor environment is generally assumed to be equal to the indoor air temperature. This is a safe assumption where a small fenestration is exposed to a large room with surface temperatures equal to the air temperature, but it is not valid in rooms where the fenestration is exposed to other large areas of glazing surfaces (e.g., greenhouse, atrium) or to other cooled or heated surfaces (Parmelee and Huebscher 1947). The radiant temperature of the outdoor environment is frequently assumed to be equal to the outdoor air temperature. This assumption may be in error, because additional radiative heat loss occurs between a fenestration and the clear sky (Berdahl and Martin 1984). Therefore, for clear-sky conditions, some effective outdoor temperature to,e should replace to in Equation (1). For methods of determining to,e , see, for example, work by AGSL (1992). Note that a fully cloudy sky is assumed in ASHRAE design conditions. The air space in a glazing unit constructed using glass with no reflective coating on the air space surfaces has a coefficient hs of 7.4 W/(m2 ·K). When a reflective coating is applied to an air space
surface, hs can be selected from Table 3 by first calculating the effective air space emissivity es,e by Equation (4): 1 es,e = -------------------------------------1 eo + 1 ei – 1
(4)
where eo and ei are the hemispherical emissivities of the two air space surfaces. Hemispherical emissivity of ordinary uncoated glass is 0.84 over a wavelength range of 0.4 to 40 m. Table 4 lists computed U-factors, using winter design conditions, for a variety of generic fenestration products, based on ASHRAEsponsored research involving laboratory testing and computer simulations. In the past, test data were used to provide more accurate results for specific products (Hogan 1988). Computer simulations (with validation by testing) are now accepted as the standard method for accurate product-specific U-factor determination. The simulation methodologies are specified in NFRC Technical Document 100 (NFRC 2014a) and are based on algorithms published in ISO Standard 15099. The International Energy Conservation Code and various state energy codes in the United States, the National Energy Code in Canada, and ASHRAE Standards 90.1 and 90.2 all reference these standards. Fenestration must be rated in accordance with the NFRC standards for code compliance. Use of Table 4 should be limited to that of an estimating tool for the early phases of design. Values in Table 4 are for vertical installation and for skylights and other sloped installations with glazing surfaces sloped 20° from the horizontal. Data are based on center-of-glass and edge-of-glass component U-factors and assume that there are no dividers. However, they apply only to the specific design conditions described in the table’s footnotes, and are typically used only to determine peak load conditions for sizing heating equipment. Although these Ufactors have been determined for winter conditions, they can also be used to estimate heat gain during peak cooling conditions, because conductive gain, which is one of several variables, is usually a small portion of the total heat gain for fenestration in direct sunlight. Glazing designs and framing materials may be compared in choosing a fenestration system that needs a specific winter design U-factor. Table 4 lists 48 glazing types, with multiple glazing categories appropriate for sealed glazing units and the addition of storm sash to other glazing units. No distinction is made between flat and domed units such as skylights. For acrylic domes, use an average gas-space width to determine the U-factor. Note that garden window and sloped/pyramid/barrel vault skylight U-factors are approximately twice those of other similar products. Although this is partially because of the difference in slope in the case of sloped/pyramid/ barrel vault skylights, it is largely because these products project out
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15.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 3 Air Space Coefficients for Horizontal Heat Flow
Air Air Air Air Space Coefficient hs , W/(m2 ·K) Space Temp. Space Effective Emissivity es,e Thickness, Temp., Diff., 0.82 0.72 0.40 0.20 0.10 0.05 K °C mm 13
15
0
10
Licensed for single user. © 2017 ASHRAE, Inc.
30
50
10
15
5 15 30 40 50 5 15 30 40 50 5 15 30 40 50 5 15 30 40 50 5 15 30 40 50 5 30 50
5.0 5.1 5.7 6.0 6.3 5.7 5.7 6.1 6.4 6.7 6.1 6.2 6.5 6.8 7.0 7.2 7.3 7.4 7.6 7.8 8.4 8.5 8.5 8.6 8.8 5.5 5.7 6.1
4.6 4.7 5.3 5.6 5.9 5.2 5.3 5.7 6.0 6.2 5.6 5.7 6.0 6.2 6.5 6.6 6.6 6.8 6.9 7.2 7.7 7.7 7.8 7.9 8.0 5.1 5.3 5.7
3.3 3.5 4.0 4.3 4.6 3.7 3.8 4.2 4.5 4.7 4.0 4.0 4.3 4.6 4.8 4.6 4.6 4.7 4.9 5.1 5.2 5.2 5.3 5.4 5.5 3.9 4.0 4.4
2.6 2.7 3.2 3.6 3.8 2.8 2.9 3.3 3.5 3.8 3.0 3.0 3.3 3.5 3.8 3.3 3.3 3.5 3.6 3.9 3.7 3.7 3.8 3.9 4.0 3.1 3.2 3.6
2.2 2.3 2.8 3.2 3.4 2.3 2.4 2.8 3.1 3.3 2.4 2.5 2.8 3.0 3.3 2.7 2.7 2.8 3.0 3.2 2.9 2.9 3.0 3.1 3.2 2.7 2.9 3.2
2.0 2.1 2.7 3.0 3.2 2.1 2.2 2.6 2.8 3.1 2.2 2.2 2.5 2.8 3.0 2.4 2.4 2.5 2.7 2.9 2.5 2.6 2.6 2.7 2.8 2.5 2.7 3.1
from the surface of the wall or roof. For instance, the skylight surface area, which includes the curb, can vary from 13 to 240% greater than the rough opening area, depending on the size and mounting method. Unless otherwise noted, all multiple-glazed units are filled with dry air. Argon units are assumed to be filled with 90% argon (Elmahdy and Yusuf 1995). U-factors for CO2-filled units are similar to argon fills. For spaces up to 13 mm, argon/SF6 (sulfur hexafluoride) mixtures up to 70% SF6 are generally the same as argon fills. Use of krypton gas can provide U-factors lower than those for argon for glazing spaces less than 13 mm. Table 4 provides data for six values of hemispherical emissivity and for 6 and 13 mm gas space widths. Emissivity of various low-e glasses varies considerably between manufacturers and processes. When emissivity is between the listed values, interpolation may be used. When manufacturers’ data are not available for low-e glass, assume that glass with a pyrolytic (hard) coating has a maximum emissivity of 0.20 and that glass with a sputtered (soft) coating has a maximum emissivity of 0.10. Tinted glass does not change the winter U-factor. Also, some surface-coated reflective glass may have an emissivity less than 0.84. Values listed are for glass units using aluminum edge spacers. If an insulated or nonmetallic spacer is used, the U-factors are approximately 0.17 W/(m2 ·K) lower. Fenestration product types are subdivided first by vertical versus sloped installation and then into two general categories: manufactured and site assembled. Manufactured products are delivered as complete units to the site. These products are typically installed in low-rise residential and small commercial/institutional/industrial buildings. Use the operable category for vertical sliders, horizontal sliders, casement, awning, pivoted, and dual-action windows, and
Air Air Air Air Space Coefficient hs , W/(m2 ·K) Space Temp. Space Effective Emissivity es,e Thickness, Temp., Diff., 0.82 0.72 0.40 0.20 0.10 0.05 K °C mm 10
0
10
30
50
7
6
5
15 0 10 30 50 15 0 10 30 50 15 0 10 30 50
5 30 50 5 30 50 5 30 50 5 30 50 <50 <50 <50 <50 <50 <50 <50 <50 <50 <50 <50 <50 <50 <50 <50
6.2 6.3 6.6 6.7 6.8 7.0 7.8 7.9 8.0 9.1 9.1 9.2 6.5 7.3 7.8 9.0 10.3 7.1 7.9 8.4 9.6 11.0 7.8 8.7 9.2 10.5 11.9
5.7 5.8 6.1 6.2 6.3 6.5 7.2 7.2 7.3 8.3 8.4 8.4 6.1 6.8 7.3 8.4 9.5 6.7 7.4 7.9 9.0 10.2 7.4 8.2 8.7 9.9 11.2
4.3 4.4 4.6 4.6 4.6 4.8 5.2 5.2 5.3 5.9 5.9 6.0 4.9 5.3 5.6 6.3 7.1 5.4 5.9 6.2 7.0 7.8 6.2 6.7 7.1 7.8 8.7
3.3 3.4 3.7 3.5 3.6 3.8 3.9 4.0 4.0 4.3 4.4 4.4 4.1 4.4 4.6 5.1 5.6 4.6 5.0 5.2 5.7 6.2 5.4 5.8 6.0 6.6 7.2
2.9 3.0 3.2 3.0 3.1 3.2 3.3 3.3 3.4 3.6 3.6 3.6 3.7 3.9 4.1 4.4 4.8 4.2 4.5 4.7 5.1 5.5 5.0 5.3 5.5 5.9 6.4
2.6 2.7 3.0 2.8 2.8 3.0 3.0 3.0 3.1 3.2 3.2 3.3 3.5 3.7 3.8 4.1 4.4 4.0 4.3 4.4 4.7 5.1 4.8 5.1 5.2 5.6 6.0
for sliding and swinging glass doors. For picture windows, use the fixed category. For products that project out from the surface of the wall, use the garden window category. For skylights, use the sloped skylight category. Site-assembled products have frame extrusions that are assembled on site, and then glazing is added. These products are typically installed in high-rise residential and larger commercial/institutional/industrial buildings. Curtain walls are typically made up of vision (transparent) and spandrel (opaque) panels. Table 4 contains representative U-factors for the vision panel (including mullions) for these assemblies. The spandrel portion of curtain walls usually consists of a metal pan filled with insulation and covered with a sheet of glass or other weatherproof covering. Although the U-factor in the center of the spandrel panel can be quite low, the metal pan is a thermal bridge, significantly increasing the U-factor of the assembly. Two-dimensional simulation, validated by testing of a curtain wall having an aluminum frame with a thermal break, found that the U-factor for the edge of the spandrel panel (the 65 mm band around the perimeter adjacent to the frame) was 40% of the way toward the U-factor of the frame. The U-factor was 0.34 W/(m2·K) for the center of the spandrel, 2.56 for the edge of the spandrel, and 6.02 for the frame (Carpenter and Elmahdy 1994). Two-dimensional heat transfer analysis or physical testing is recommended to determine the U-factor of spandrel panels. Use the sloped/overhead glazing category for sloped glazing panels comparable to curtain walls. Physical testing of double-glazed units showed U-factors of 5.74 W/(m2·K) for a thermally broken aluminum pyramidal skylight and 7.4 W/(m2·K) for an aluminum-frame half-round barrel
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Fenestration
15.9 Table 4
U-Factors for Various Fenestration Products in W/(m2 ·K)i Vertical Installation
Product Type Frame Type ID 1 2 3 4 5 6 7
Licensed for single user. © 2017 ASHRAE, Inc.
8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48
Glazing Type
Glass Only Center of Glass
Edge of Glass
Operable (including sliding and swinging glass doors) Aluminum Aluminum Reinforced Without With Vinyl/ Thermal Thermal Aluminum Break Break Clad Wood
Single Glazing 3.2 mm glass 5.91 5.91 7.01 6 mm acrylic/polycarb 5.00 5.00 6.23 3.2 mm acrylic/polycarb 5.45 5.45 6.62 Double Glazing 6 mm airspace 3.12 3.63 4.62 13 mm airspace 2.73 3.36 4.30 6 mm argon space 2.90 3.48 4.43 13 mm argon space 2.56 3.24 4.16 Double Glazing, e = 0.60 on surface 2 or 3 6 mm airspace 2.95 3.52 4.48 13 mm airspace 2.50 3.20 4.11 6 mm argon space 2.67 3.32 4.25 13 mm argon space 2.33 3.08 3.98 Double Glazing, e = 0.40 on surface 2 or 3 6 mm airspace 2.78 3.40 4.34 13 mm airspace 2.27 3.04 3.93 6 mm argon space 2.44 3.16 4.07 13 mm argon space 2.04 2.88 3.75 Double Glazing, e = 0.20 on surface 2 or 3 3.24 4.16 6 mm airspace 2.56 13 mm airspace 1.99 2.83 3.70 6 mm argon space 2.16 2.96 3.84 13 mm argon space 1.70 2.62 3.47 Double Glazing, e = 0.10 on surface 2 or 3 6 mm airspace 2.39 3.12 4.02 13 mm airspace 1.82 2.71 3.56 6 mm argon space 1.99 2.83 3.70 13 mm argon space 1.53 2.49 3.33 Double Glazing, e = 0.05 on surface 2 or 3 6 mm airspace 2.33 3.08 3.98 13 mm airspace 1.70 2.62 3.47 6 mm argon space 1.87 2.75 3.61 13 mm argon space 1.42 2.41 3.24 Triple Glazing 6 mm airspace 2.16 2.96 3.78 13 mm airspace 1.76 2.67 3.46 6 mm argon space 1.93 2.79 3.60 13 mm argon space 1.65 2.58 3.36 Triple Glazing, e = 0.20 on surface 2, 3, 4, or 5 6 mm airspace 1.87 2.75 3.55 13 mm airspace 1.42 2.41 3.18 6 mm argon space 1.59 2.54 3.32 13 mm argon space 1.25 2.28 3.04 Triple Glazing, e = 0.20 on surfaces 2 or 3 and 4 or 5 6 mm airspace 1.65 2.58 3.36 13 mm airspace 1.14 2.19 2.95 6 mm argon space 1.31 2.32 3.09 13 mm argon space 0.97 2.05 2.81 Triple Glazing, e = 0.10 on surfaces 2 or 3 and 4 or 5 6 mm airspace 1.53 2.49 3.27 13 mm airspace 1.02 2.10 2.85 6 mm argon space 1.19 2.23 2.99 13 mm argon space 0.80 1.92 2.67 Quadruple Glazing, e = 0.10 on surfaces 2 or 3 and 4 or 5 6 mm airspaces 1.25 2.28 3.04 13 mm airspaces 0.85 1.96 2.71 6 mm argon spaces 0.97 2.05 2.81 13 mm argon spaces 0.68 1.83 2.57 6 mm krypton spaces 0.68 1.83 2.57
Wood/ Vinyl
Fixed
Aluminum Aluminum Reinforced Insulated Without With Vinyl/ Fiberglass/ Thermal Thermal Aluminum Wood/ Vinyl Break Break Clad Wood Vinyl
Insulated Fiberglass/ Vinyl
6.08 5.35 5.72
5.27 4.59 4.93
5.20 4.52 4.86
4.83 4.18 4.51
6.38 5.55 5.96
6.06 5.23 5.64
5.58 4.77 5.18
5.58 4.77 5.18
5.40 4.61 5.01
3.61 3.31 3.44 3.18
3.24 2.96 3.08 2.84
3.14 2.86 2.98 2.74
2.84 2.58 2.69 2.46
3.88 3.54 3.68 3.39
3.52 3.18 3.33 3.04
3.18 2.85 3.00 2.71
3.16 2.83 2.98 2.69
3.04 2.72 2.86 2.58
3.48 3.14 3.27 3.01
3.12 2.80 2.92 2.68
3.02 2.70 2.82 2.58
2.73 2.42 2.54 2.31
3.73 3.34 3.49 3.20
3.38 2.99 3.13 2.84
3.04 2.67 2.81 2.52
3.02 2.65 2.79 2.50
2.90 2.53 2.67 2.39
3.35 2.96 3.09 2.79
3.00 2.64 2.76 2.48
2.90 2.54 2.66 2.38
2.61 2.27 2.38 2.11
3.59 3.15 3.30 2.95
3.23 2.79 2.94 2.60
2.90 2.48 2.62 2.29
2.88 2.46 2.60 2.27
2.77 2.35 2.49 2.16
3.18 2.75 2.88 2.53
2.84 2.44 2.56 2.24
2.74 2.34 2.46 2.14
2.46 2.07 2.19 1.88
3.39 2.91 3.05 2.66
3.04 2.55 2.70 2.30
2.71 2.24 2.38 2.00
2.69 2.22 2.36 1.98
2.58 2.12 2.26 1.88
3.05 2.62 2.75 2.40
2.72 2.32 2.44 2.12
2.62 2.22 2.34 2.02
2.34 1.96 2.07 1.76
3.25 2.76 2.91 2.51
2.89 2.40 2.55 2.16
2.57 2.10 2.24 1.86
2.55 2.08 2.22 1.84
2.44 1.98 2.12 1.74
3.01 2.53 2.66 2.31
2.68 2.24 2.36 2.04
2.58 2.14 2.26 1.94
2.31 1.88 2.00 1.69
3.20 2.66 2.81 2.42
2.84 2.30 2.45 2.06
2.52 2.00 2.15 1.76
2.50 1.98 2.12 1.74
2.39 1.88 2.02 1.65
2.78 2.47 2.60 2.39
2.46 2.18 2.30 2.10
2.42 2.14 2.26 2.06
2.17 1.90 2.02 1.83
3.02 2.68 2.82 2.58
2.68 2.34 2.49 2.24
2.36 2.03 2.17 1.93
2.36 2.03 2.17 1.93
2.25 1.92 2.06 1.83
2.56 2.21 2.34 2.08
2.26 1.94 2.06 1.82
2.22 1.90 2.02 1.78
1.98 1.67 1.79 1.55
2.78 2.38 2.53 2.24
2.44 2.05 2.20 1.90
2.12 1.74 1.89 1.60
2.12 1.74 1.89 1.60
2.01 1.64 1.78 1.50
2.39 1.99 2.12 1.86
2.10 1.74 1.86 1.62
2.06 1.69 1.82 1.57
1.83 1.48 1.59 1.36
2.58 2.14 2.29 1.99
2.24 1.80 1.95 1.65
1.93 1.50 1.65 1.36
1.93 1.50 1.65 1.36
1.83 1.40 1.55 1.26
2.30 1.90 2.04 1.73
2.02 1.66 1.78 1.49
1.98 1.61 1.73 1.45
1.75 1.40 1.52 1.24
2.48 2.04 2.19 1.84
2.15 1.70 1.85 1.51
1.84 1.41 1.55 1.22
1.84 1.41 1.55 1.22
1.73 1.31 1.45 1.12
2.08 1.77 1.86 1.64 1.64
1.82 1.54 1.62 1.41 1.41
1.78 1.49 1.57 1.37 1.37
1.55 1.28 1.36 1.16 1.16
2.24 1.89 1.99 1.74 1.74
1.90 1.55 1.65 1.41 1.41
1.60 1.26 1.36 1.12 1.12
1.60 1.26 1.36 1.12 1.12
1.50 1.17 1.26 1.03 1.03
Notes: 1. All heat transmission coefficients in this table include film resistances and are based on winter conditions of –18°C outdoor air temperature and 21°C indoor air temperature, with 6.7 m/s outdoor air velocity and zero solar flux. Except for single glazing, small changes in the indoor and outdoor temperatures do not significantly affect overall U-factors. Coefficients are for vertical position except skylight values, which are for 20° from horizontal with heat flow up.
2. Glazing layer surfaces are numbered from outdoor to indoor. Double, triple, and quadruple refer to number of glazing panels. All data are based on 3 mm glass, unless otherwise noted. Thermal conductivities are: 0.917 W/(m·K) for glass, and 0.19 W/ (m·K) for acrylic and polycarbonate. 3. Standard spacers are metal. Edge-of-glass effects are assumed to extend over the 63.5 mm band around perimeter of each glazing unit.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.10
2017 ASHRAE Handbook—Fundamentals (SI) Table 4 U-Factors for Various Fenestration Products in W/(m2 ·K)i (Concluded) Vertical Installation
Garden Windows
Sloped Installation
Curtainwall
Glass Only (Skylights)
Licensed for single user. © 2017 ASHRAE, Inc.
Aluminum Aluminum Aluminum Without Without With Thermal Wood/ Thermal Thermal Structural Break Vinyl Break Break Glazing
Center of Glass
Edge of Glass
Manufactured Skylight
Site-Assembled Sloped/Overhead Glazing
Aluminum Aluminum Reinforced Without With Vinyl/ Thermal Thermal Aluminum Wood/ Break Break Clad Wood Vinyl
Aluminum Without Thermal Break
Aluminum With Thermal Break
Structural Glazing
ID
14.21 12.70 13.45
11.94 10.42 11.18
6.86 6.03 6.44
6.27 5.44 5.86
6.27 5.44 5.86
6.76 5.85 6.30
6.76 5.85 6.30
10.03 9.09 9.56
9.68 8.74 9.21
9.16 8.23 8.70
8.05 7.45 7.89
7.66 6.83 7.24
7.64 6.80 7.22
7.10 6.27 6.68
1 2 3
9.78 9.19 9.44 8.94
7.50 6.92 7.17 6.67
4.38 4.03 4.18 3.89
3.79 3.45 3.60 3.30
3.56 3.22 3.37 3.07
3.29 3.24 3.01 3.01
3.75 3.71 3.56 3.56
6.23 6.17 5.96 5.96
5.46 5.41 5.19 5.19
5.21 5.16 4.94 4.94
4.79 4.74 4.54 4.54
4.54 4.49 4.30 4.30
4.71 4.68 4.52 4.52
3.75 3.70 3.51 3.51
4 5 6 7
9.53 8.86 9.11 8.61
7.25 6.58 6.84 6.33
4.23 3.84 3.99 3.69
3.65 3.25 3.40 3.11
3.41 3.02 3.17 2.88
3.07 3.01 2.78 2.78
3.60 3.56 3.40 3.40
6.01 5.96 5.74 5.74
5.24 5.19 4.97 4.97
4.99 4.94 4.72 4.72
4.59 4.54 4.34 4.34
4.35 4.30 4.10 4.10
4.56 4.52 4.37 4.37
3.55 3.51 3.31 3.31
8 9 10 11
9.28 8.52 8.77 8.18
7.00 6.25 6.50 5.91
4.08 3.64 3.79 3.45
3.50 3.06 3.21 2.86
3.27 2.83 2.97 2.63
2.90 2.84 2.50 2.61
3.48 3.44 3.20 3.28
5.85 5.79 5.46 5.57
5.08 5.02 4.69 4.80
4.83 4.78 4.45 4.56
4.44 4.39 4.09 4.19
4.20 4.15 3.86 3.96
4.45 4.41 4.18 4.25
3.41 3.36 3.07 3.17
12 13 14 15
8.94 8.10 8.35 7.67
6.67 5.82 6.08 5.39
3.89 3.40 3.54 3.15
3.30 2.81 2.96 2.56
3.07 2.58 2.73 2.33
2.61 2.61 2.22 2.27
3.28 3.28 3.00 3.04
5.57 5.57 5.19 5.24
4.80 4.80 4.42 4.47
4.56 4.56 4.18 4.24
4.19 4.19 3.84 3.89
3.96 3.96 3.61 3.66
4.25 4.25 3.98 4.02
3.17 3.17 2.83 2.88
16 17 18 19
8.69 7.84 8.10 7.41
6.42 5.57 5.82 5.14
3.74 3.25 3.40 3.00
3.16 2.66 2.81 2.42
2.92 2.43 2.58 2.18
2.50 2.50 2.04 2.16
3.20 3.20 2.88 2.96
5.46 5.46 5.02 5.13
4.69 4.69 4.25 4.36
4.45 4.45 4.02 4.13
4.09 4.09 3.69 3.79
3.86 3.86 3.46 3.56
4.18 4.18 3.86 3.94
3.07 3.07 2.68 2.78
20 21 22 23
8.61 7.67 7.93 7.24
6.33 5.39 5.65 4.96
3.69 3.15 3.30 2.90
3.11 2.56 2.71 2.32
2.88 2.33 2.48 2.09
2.39 2.44 1.93 2.04
3.12 3.16 2.79 2.88
5.35 5.41 4.91 5.02
4.58 4.64 4.14 4.25
4.34 4.40 3.91 4.02
3.99 4.04 3.58 3.69
3.76 3.81 3.37 3.46
4.10 4.14 3.77 3.86
2.97 3.02 2.58 2.68
24 25 26 27
see note 7
see note 7
3.48 3.14 3.28 3.04
2.91 2.57 2.71 2.47
2.62 2.27 2.42 2.17
2.22 2.04 1.99 1.87
3.00 2.88 2.83 2.75
5.13 4.96 4.91 4.80
4.24 4.07 4.01 3.90
4.03 3.87 3.81 3.70
3.63 3.48 3.43 3.33
3.55 3.40 3.35 3.25
3.92 3.80 3.76 3.68
2.70 2.56 2.51 2.41
28 29 30 31
see note 7
see note 7
3.23 2.84 2.99 2.69
2.66 2.27 2.42 2.12
2.37 1.97 2.12 1.83
1.93 1.76 1.59 1.53
2.79 2.67 2.54 2.49
4.85 4.68 4.52 4.46
3.96 3.79 3.63 3.57
3.76 3.59 3.43 3.37
3.38 3.22 3.07 3.02
3.30 3.16 3.01 2.96
3.72 3.59 3.47 3.43
2.46 2.31 2.17 2.12
32 33 34 35
see note 7
see note 7
3.04 2.59 2.74 2.44
2.47 2.02 2.17 1.87
2.17 1.73 1.87 1.58
1.65 1.53 1.36 1.25
2.58 2.49 2.36 2.28
4.57 4.46 4.29 4.18
3.68 3.57 3.40 3.29
3.48 3.37 3.21 3.10
3.12 3.02 2.86 2.76
3.06 2.96 2.81 2.71
3.51 3.43 3.30 3.22
2.22 2.12 1.97 1.87
36 37 38 39
see note 7
see note 7
2.94 2.49 2.64 2.29
2.37 1.92 2.07 1.72
2.07 1.63 1.78 1.43
1.53 1.42 1.19 1.14
2.49 2.41 2.23 2.19
4.46 4.35 4.13 4.07
3.57 3.46 3.24 3.18
3.37 3.27 3.04 2.99
3.02 2.91 2.71 2.66
2.96 2.86 2.66 2.61
3.43 3.34 3.18 3.13
2.12 2.02 1.82 1.77
40 41 42 43
see note 7
see note 7
2.69 2.34 2.44 2.19 2.19
2.12 1.77 1.87 1.62 1.62
1.83 1.48 1.58 1.33 1.33
1.25 1.08 1.02 0.91 0.74
2.28 2.14 2.10 2.01 1.87
4.18 4.02 3.96 3.85 3.68
3.29 3.12 3.07 2.96 2.79
3.10 2.93 2.88 2.77 2.60
2.76 2.60 2.55 2.45 2.29
2.71 2.56 2.51 2.41 2.26
3.22 3.09 3.05 2.96 2.83
1.87 1.72 1.67 1.58 1.43
44 45 46 47 48
4. Product sizes are described in Figure 4, and frame U-factors are from Table 1. 8. U-factors in this table were determined using NFRC 100-91. They have not been 5. Use U = 3.40 W/(m2·K) for glass block with mortar but without reinforcing or updated to the current rating methodology in NFRC 100 (2014a). framing. 6. Use of this table should be limited to that of an estimating tool for early phases of design. 7. Values for triple- and quadruple-glazed garden windows are not listed, because these are not common products.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Fenestration
15.11
Frame Width, mm Frame Material
Licensed for single user. © 2017 ASHRAE, Inc.
Aluminum without thermal break Aluminum with thermal break Aluminum-clad wood/reinforcing vinyl Wood/vinyl Insulated fiberglass/vinyl Structural glazing
Operable
Fixed
Garden Window
38 53 71 71 79 N/A
33 33 41 41 46 N/A
44 N/A N/A 44 N/A N/A
Skylight
Curtain Wall
Sloped/Overhead Glazing
18 18 23 23 N/A N/A
57 57 N/A N/A N/A 57
57 57 N/A N/A N/A 64
Fig. 4 Frame Widths for Standard Fenestration Units vault (both normalized to a rough opening of 2.4 by 2.4 m). Until more conclusive results are available, U-factors for these systems can be estimated by multiplying the site-assembled sloped/overhead glazing values in Table 4 by the ratio of total product surface area (including curbs) to rough opening area. These ratios range from 1.2 to 2.0 for low-slope skylights, 1.4 to 2.1 for pyramid assemblies sloped at 45°, and 1.7 to 2.9 for semicircular barrel vault assemblies. U-factors in Table 4 are based on definitions of the six product types, frame sizes, and proportion of frame to glass area shown in Figure 4. Four of the products are manufactured type. Sizes are as defined in NFRC Technical Document 100 (2014a): operable and fixed (nonoperable) glazing units are 1.8 m2 in area, and the overall size corresponds to a 1200 by 1500 mm fenestration product. The garden window category is 1.8 m2 in projected area (3.24 m2 in surface area) and 1500 mm wide by 1200 mm high by 380 mm deep. The manufactured skylight category is a nominal 1.44 m2 in area, corresponding to a 1200 by 1200 mm skylight. The nominal dimensions of a roof-mounted skylight correspond to centerline spacing of roof framing members; consequently, the rough opening dimensions are 1180 by 1180 mm. The curtain wall and sloped/overhead glazing categories are a nominal 4 m2 in area, representing repeating 2000 by 2000 mm panels. The nominal dimensions correspond to centerline spacing of the head and sill and vertical mullions. Six frame types are listed (although not all for any one category) in order of improving thermal performance. The most conservative assumption is to use the frame category of aluminum frame without a thermal break (although there are products on the market that have higher U-factors). The aluminum frame with a thermal break is for frames having at least a 10 mm thermal break between the indoors and outdoors for all members including both the frame and the operable sash, if applicable. Some products are available with significantly wider thermal breaks, which achieve considerable improvement. The aluminum-clad wood/reinforced vinyl category represents vinyl-frame products, such as sliding glass doors or large windows that have extensive metal reinforcing within the frame and wood products with extensive metal, usually on the outdoor surface of the frame. Both of these factors provide short circuits, which degrade the thermal performance of the frame material. The wood/ vinyl frame category represents the improved thermal performance that is possible if the thermal short circuits from the previous frame
Table 5
Glazing U-Factors for Various Wind Speeds in W/(m2 ·K) Wind Speed, km/h 24
12
0
0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5
0.46 0.92 1.33 1.74 2.15 2.56 2.98 3.39 3.80 4.21 4.62 5.03 5.95
0.42 0.85 1.27 1.69 2.12 2.54 2.96 3.38 3.81 4.23 4.65 5.08 5.50
category do not exist. Insulated fiberglass/vinyl represents fiberglass or vinyl frames that do not have metal reinforcing and whose frame cavities are filled with insulation. For several site-assembled product types, there is a structural glazing frame category that represents products where sheets of glass are butt-glazed to each other using a sealant only, and framing members are not exposed to the exterior. For glazing with a steel frame, use aluminum frame values. For aluminum fenestration with wood trim or vinyl cladding, use the values for aluminum. Frame type refers to the primary unit; therefore, when storm sash is added over another fenestration product, use values given for the nonstorm product. To estimate the overall U-factor of a fenestration product that differs significantly from the assumptions given in Table 4 and/or Figure 4, first determine the area that is frame/sash, center-ofglass, and edge-of-glass (based on a 63.5 mm band around the perimeter of each glazing unit). Next, determine the appropriate component U-factors. These can be taken either from the standard values listed in italics in Table 4 for glass, from the values in Table 1 for frames, or from some other source such as test data or computed factors. Finally, multiply the area and the component Ufactors, sum these products, and then divide by the rough opening
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.12
2017 ASHRAE Handbook—Fundamentals (SI)
in the building envelope where this product will fit to obtain the overall U-factor Uo. Table 5 provides approximate data to convert the overall U-factor at one wind condition to a U-factor at another. Example 1. Estimate the design U-factor for a manufactured fixed fenestration product with a reinforced vinyl frame and double-glazing with a sputter-type low-e coating (e = 0.10). The gap is 13 mm wide and argon-filled, and the spacer is metal. The outdoor wind speed is 12 km/h. Solution: Locate the glazing system type in the first column of Table 4 (ID = 23), then find the appropriate product type (fixed) and frame type (reinforced vinyl). The U-factor listed (in the tenth column of U-factors) is 1.86 W/(m2 ·K). This U-factor is for 24 km/h outdoor wind speed. From Table 5, interpolate 1.86 in the 24 km/h column to the corresponding value in the 12 km/h column. U 12 km/h – 1.33 1.86 – 1.50 --------------------------- = -----------------------------------1.74 – 1.33 2.00 – 1.50
Licensed for single user. © 2017 ASHRAE, Inc.
U12 km/s = 1.63 W/(m2 ·K) Example 2. Estimate a representative U-factor for a wood-framed, 970 by 2080 mm swinging French door with eight 280 by 400 mm panes (true divided panels), each consisting of clear double-glazing with a 6.5 mm air space and a metal spacer. Solution: Without more detailed information, assume that the dividers have the same U-factor as the frame and that the divider edge has the same U-factor as the edge-of-glass. Calculate center-of-glass, edge-ofglass, and frame areas: Acg = 8[(280 – 130)(400 – 130)]/106 = 0.324 m2 Aeg = 8(280 400)/106 – 0.324 = 0.572 m2 Af = (970 2080)/106 – 8(280 400)/106 = 1.122 m2 Select center-of-glass, edge-of-glass, and frame U-factors. These component U-factors are 3.12 and 3.63 W/(m2 ·K) (from Table 4, glazing ID = 4, U-factor columns 1 and 2) and 2.90 W/(m2 ·K) (from Table 1, wood frame, metal spacer, operable, double-glazing), respectively. From Equation (2), U french
door
3.12 0.324 + 3.63 0.572 + 2.90 1.122 = --------------------------------------------------------------------------------------------------------------------- 0.97 2.08 2
= 3.14 W/(m ·K) Example 3. Estimate the overall average U-factor for a multifloor curtain wall assembly that is part vision glass and part opaque spandrel. The typical floor-to-floor height is 3.6 m, and the building module is 1.2 m as reflected in the spacing of the mullions both horizontally and vertically. For a representative section 1.2 m wide and 3.6 m tall, one of the modules is glazed and the other two are opaque. The mullions are aluminum frame with a thermal break 80 mm wide and centered on the module. The glazing unit is double glazing with a pyrolytic low-e coating (e = 0.40) and has a 13 mm gap filled with air and a metal spacer. The spandrel panel has a metal pan backed by R = 3.5 (m2 ·K)/W insulation and no intermediate reinforcing members. Solution: It is necessary to calculate the U-factor for the glazed module and for the opaque spandrel modules, and then to do an area-weighted average to determine the average U-factor for the overall curtain wall assembly. First, calculate the overall U-factor for the glazed module. Calculate center-of-glass, edge-of-glass, and frame areas. The glazed area is 1120 by 1120 mm (1200 mm module, 38 mm of mullions on each edge). Acg = (1120 – 130)(1120 – 130)/106 = 0.9801 m2 Aeg = (1120 Af = [(1200
1120)/106
1200)/106
– 0.9801 = 0.2743
m2
– (1120 1120)]/106 = 0.1856 m2
Select center-of-glass, edge-of-glass, and frame U-factors. These component U-factors are 2.27 and 3.04 W/(m2 ·K) (from Table 4, ID = 13, columns 1 and 2) and 9.94 W/(m2 ·K) (from Table 1, aluminum frame with a thermal break, metal spacer, curtain wall, double glazing), respectively. From Equation (2), U glazing
module
2.27 0.9801 + 3.04 0.2743 + 9.94 0.1856 = ------------------------------------------------------------------------------------------------------------------------------ 1.2 1.2 2
= 3.41 W/(m ·K) Then, calculate the overall U-factor for the two opaque spandrel modules. The center-of-spandrel, edge-of spandrel, and frame areas are the same as the glazed module. The frame U-factor is the same. Calculate the center-of-spandrel U-factor. In this particular case, the R-value of the insulation does not need to be rated, because there are no intermediate framing members penetrating it and providing thermal short circuits. When the resistance of the insulation [3.5 (m2 · K)/W] is added to the exterior air film resistance of 0.03 (m2 · K)/W and the interior air film resistance of 0.12 (m2 · K)/W (from Table 1, Chapter 26), the total resistance is 3.65 (m2 ·K)/W, and the U-factor is 1/3.65 = 0.274 W/(m2 ·K). The edge-of-spandrel U-factor is 40% of the way to the frame U-factor, which is 0.274 + [0.40(9.94 – 0.274)] = 4.14 W/(m2 ·K). Uopaque
spandrel module
0.274 0.9801 + 4.14 0.2743 + 9.94 0.1856 = --------------------------------------------------------------------------------------------------------------------------------- 1.2 1.2 2 = 2.26 W/(m ·K) Finally, calculate the overall average U-factor for the curtain wall assembly, including the one module of vision glass and the two modules of opaque spandrel. U curtain
wall
3.41 1.2 1.2 + 2.26 2 1.2 1.2 = --------------------------------------------------------------------------------------------------------------3 1.2 1.2 2
= 2.64 W/(m ·K) Note that even with double glazing having a low-e coating and with R = 3.5 (m2 ·K)/W in the opaque areas, this curtain wall with metal pans only has an overall R-value of approximately 0.38 (m2 ·K)/W. Example 4. Estimate the U-factor for a semicircular barrel vault that is 6 m wide, 3 m tall, and 10 m long mounted on a 150 mm curb. The barrel vault has an aluminum frame without a thermal break. The glazing is double with a 13 mm gap width filled with air and a low-e coating (e = 0.20). Solution: An approximation can be made by multiplying the U-factor for a site-assembled sloped/overhead glazing product having the same frame and glazing features by the ratio of the surface area (including the curb) of the barrel vault to the rough opening area in the roof that the barrel vault fits over. First, determine the surface area (including the curb) of the barrel vault: Area of the curved portion of the barrel vault = ( diameter/2) length = (3.14 6/2) 10 = 94.25 m2 Area of the two ends of the barrel vault = 2 ( radius2)/2 = r2 = 3.14 32 = 28.27 m2 Area of the curb = perimeter curb height = (6 + 10 + 6 + 10) 0.150 = 4.8 m2 Total surface area of the barrel vault = 94.25 + 28.27 + 4.8 = 127.3 m2 Second, determine the rough opening area in the roof that the barrel vault fits over: = length width = 6 10 = 60 m2 Third, determine the ratio of the surface area to the rough opening area: = 127.3/60 = 2.12 Fourth, determine the U-factor from Table 4 of a site-assembled sloped/ overhead glazing product having the same frame and glazing features.
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The U-factor is 3.51 W/(m2 ·K) (ID = 17, 12th column on the second page of Table 4). Fifth, determine the estimated U-factor of the barrel vault. Ubarrel vault = Usloped overhead glazing surface area/rough opening for the barrel vault = 3.51 2.12 = 7.44 W/(m2 ·K)
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3.3
REPRESENTATIVE U-FACTORS FOR DOORS
Doors are often an overlooked component in the thermal integrity of the building envelope. Although entry doors (swinging, revolving, etc.) represent a small portion of the building envelope of residential, commercial, and institutional buildings, their U-factor is usually many times higher than that of the walls or ceilings. In some commercial and industrial buildings, vehicular access doors (upward-acting doors) represent a significant area of heat loss. Table 6 contains representative U-factors for swinging doors determined through computer simulation (Carpenter and Hogan 1996). These are generic values, and product-specific values determined in accordance with standards should be used whenever available. NFRC Technical Document 100 (NFRC 2014a) gives procedures for evaluating the performance of entry and vehicular access doors. Tables 7 to 9 contain representative U-factors for revolving, emergency exit, garage, and aircraft hangar doors determined through testing (McGowan et al. 2006). Swinging doors can be divided into two categories: slab and stile-and-rail. A stile-and-rail door is a swinging door with a fullglass insert supported by horizontal rails and vertical stiles. The Table 6 Design U-Factors of Swinging Doors in W/(m2 ·K)
Door Type (Rough Opening = 970 2080 mm) Slab Doors Wood slab in wood frame a 6% glazing (560 200 lite) 25% glazing (560 910 lite) 45% glazing (560 1620 lite) More than 50% glazing Insulated steel slab with wood edge in wood frame b 6% glazing (560 200 lite) 25% glazing (560 910 lite) 45% glazing (560 1630 lite) More than 50% glazing Foam-insulated steel slab with metal edge in steel frame c 6% glazing (560 200 lite) 25% glazing (560 910 lite) 45% glazing (560 1630 lite) More than 50% glazing Cardboard honeycomb slab with metal edge in steel frame Stile-and-Rail Doors Sliding glass doors/French doors Site-Assembled Stile-and-Rail Doors Aluminum in aluminum frame Aluminum in aluminum frame with thermal break
Double Double Glazing Glazing with with e = 0.10, No Single 12.7 mm 12.7 mm Glazing Glazing Air Space Argon 2.61 — — —
• Rolling (also called roll-up) doors consist of small metal slats of approximately 65 mm in height that travel in vertical guides and roll up around a metal barrel to open. • Sectional (also called garage) doors consist of a series of approximately 460 to 815 mm high sections that travel in vertical tracks to open. • Folding (also called biparting) doors, commonly used in aircraft hangars, have two large sections that also travel in vertical tracks to open, but fold together when the door is fully open. There is a wide range in the design of insulated upward-acting doors. Factors affecting heat transfer include insulation thickness, section/slat design, and section/slat interface design (which may Table 7 Design U-factors for Revolving Doors in W/(m2·K)
2.73 2.61 2.50 3.29 2.61 2.38 3.92 2.61 2.21 Use Table 4 (operable)
0.91 — — —
stiles and rails are typically either solid wood members or extruded aluminum or vinyl, as shown in Figure 5. Most residential doors are slab type with solid wood, steel, or a fiberglass skin over foam insulation in a wood frame with aluminum sill. The edges of the steel skin door are normally wood to provide a thermal break. In commercial construction, doors are either steel skin over foam insulation in a steel frame (i.e., utility doors) or a full glass door made up of aluminum stiles, rails, and frame (i.e., entrance doors). The most important factors affecting door U-factor are material construction, glass size, and glass type. Frame depth, slab width, and number of panels have a minor effect on door performance. Side lites and double doors have U-factors similar to a single door of the same construction. For wood slab doors in a wood frame, the glazing area has little effect on the U-factor. For an insulated steel slab in a wood frame, however, glazing area strongly affects U-factor. Typical commercial insulated slab doors have a U-factor approximately twice that of residential insulated doors, the prime reason being thermal bridging of the slab edge and the steel frame. Stile-and-rail doors, even if thermally broken, have U-factors 50% higher than a fullglass commercial steel slab door. There are three generic types of upward-acting doors:
Type 3-wing 4-wing Open*
1.19 1.08 1.02 2.21 1.48 1.31 3.29 1.99 1.48 Use Table 4 (operable)
Size (Width Height)
U-Factor
2.44 2.13 m 3.28 2.44 m 2.13 1.98 m 2.13 2.29 m 2.08 2.13 m
4.46 4.53 3.56 3.63 7.49
*U-factor of open door determined using NFRC Technical Document 100-91. It has not been updated to current rating methodology in NFRC Technical Document 100 (2014a).
2.10 — — —
2.50 2.33 2.21 3.12 2.73 2.50 4.03 3.18 2.73 Use Table 4 (operable)
3.46
Use Table 4 (operable) — —
7.49 6.42
5.28 4.20
4.49 3.58
Notes: aThermally broken sill [add 0.17 W/(m2 ·K) for non-thermally broken sill] bNon-thermally broken sill cNominal U-factors are through center of insulated panel before consideration of thermal bridges around edges of door sections and because of frame.
Fig. 5 Details of Stile-and-Rail Door
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Table 8
Design U-factors for Double-Skin Steel Emergency Exit Doors in W/(m2 ·K) Core Insulation
Rough Opening Size
Thickness, mm
Type
0.9 2 m
1.8 2 m
35*
Honeycomb kraft paper Mineral wool, steel ribs Polyurethane foam Honeycomb kraft paper Mineral wool, steel ribs Polyurethane foam Honeycomb kraft paper Mineral wool, steel ribs Polyurethane foam Honeycomb kraft paper Mineral wool, steel ribs Polyurethane foam
3.23 2.50 1.92 3.25 2.30 1.77 3.38 2.67 2.10 3.38 2.47 1.95
2.97 2.05 1.60 3.06 1.90 1.50 3.11 2.21 1.77 3.22 2.08 1.69
44*
35
44
*With thermal break
Fig. 6
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Table 9 Design U-Factors for Double-Skin Steel Sectional, Tilt-Up, and Aircraft Hangar Doors in W/(m2 ·K) Insulation Thickness, mm Type 45 Polyurethane, thermally broken 35 Extruded polystyrene, steel ribs Expanded polystyrene, steel ribs 50 Extruded polystyrene, steel ribs Expanded polystyrene, steel ribs 75 Extruded polystyrene, steel ribs Expanded polystyrene, steel ribs 100 Extruded polystyrene, steel ribs Expanded polystyrene, steel ribs 150 Extruded polystyrene, steel ribs Expanded polystyrene, steel ribs 100 Expanded polystyrene Mineral wool, steel ribs Extruded polystyrene 150 Expanded polystyrene Mineral wool, steel ribs Extruded polystyrene — Uninsulated
Sectional or Tilt-Up a 2.7 2.1 m 1.59 2.21 2.04 1.87 1.76 1.59 1.53 1.42 1.36 1.19 1.19
6.53d
Optical Properties of a Single Glazing Layer
is only one glazing between the indoors and outdoors (+ position). U-factors are given in Table 7 for both positions.
Aircraft Hangar 22 3.7 m b
1.42 1.42 1.31 1.19 1.31 1.14 6.25
73 15.2 mc
0.91 0.91 0.85 0.74 0.734 0.68 6.98
Notes: aU-factor determined using NFRC Technical Document 100-91. Not updated to current rating methodology in NFRC (2014a) Technical Document 100-2014. bTypical size for a small private airplane (single- or twin-engine.) cTypical hangar door for a midsized commercial jet airliner. dU-factor determined for 3 3 m sectional door, but is representative of similar products of different sizes.
include a thermal break). For noninsulated upward-acting doors, there is very little difference between the center value and the total value: the value is essentially equal to that of single glazing. The center of an insulated door has a relatively low U-factor, but thermal bridging at the door frame and section interfaces can affect the door assembly U-factor. Center-of-door U-factors for vehicular access doors may be used in thermal calculations for buildings only if the door design or door size indicates little difference with respect to the total door assembly U-factor. Many commercial buildings use revolving entrance doors. Most of these doors are of similar design: single glazing in an aluminum frame without thermal break. The door, however, can be in two positions: closed (X-shaped as viewed from above) or open (+-shaped as viewed from above). At nighttime, these doors are locked in the X position, effectively creating a double-glazed system. During the daytime, the door revolves and is often left positioned so that there
4. SOLAR HEAT GAIN AND VISIBLE TRANSMITTANCE Fenestration solar heat gain has two components. First is directly transmitted solar radiation. The quantity of radiation entering the fenestration directly is governed by the solar transmittance of the glazing system, and is determined by multiplying the incident irradiance by the glazing area and its solar transmittance. The second component is the inward flowing fraction of absorbed solar radiation, radiation that is absorbed in the glazing and framing materials of the fenestration, some of which is subsequently conducted, convected, or radiated to the interior of the building. Visible transmittance is the solar radiation transmitted through fenestration weighted with respect to the photopic response of the human eye. It physically represents the perceived clearness of the fenestration, and is likely different from the solar transmittance of the same fenestration. The underlying physics behind solar heat gain and visible transmittance can be very complex, but a rudimentary understanding is required if technologies such as low-e coatings are to be discussed. Accurately calculating the solar heat gain and visible transmittance of a fenestration system, including the effects of angular and spectral dependence, in the presence of multiple glazing and shade layers, is very complex. Refer to ISO Standard 15099 or the ASHRAE Handbook Online supplemental features for this chapter for complete details of how to do this calculation. Software such as LBNL WINDOW 7 (LBNL 2016) incorporate these advanced calculations and can be used for more detailed fenestration analysis.
4.1
SOLAR-OPTICAL PROPERTIES OF GLAZING
Optical Properties of Single Glazing Layers Radiation passing from one medium into another is partly transmitted and partly reflected at the interface between the two media. Further, as this radiation passes through either medium, an additional fraction is absorbed because of the absorptivity of the material. Materials that do not absorb radiation completely, such as air or glass, are classified as being transparent or translucent. Translucent glazings exhibit sufficient light-diffusing properties that images of objects viewed through it are blurred. Opaque glazings transmit no perceptible light.
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If solar radiation incident on glazing is considered, the transmittance T, reflectance R, and absorptance A of the glazing layer contain the effects of multiple reflections between the two interfaces of the layer as well as the effects of absorption during the passage through the layer of each interreflection (Figure 6). For radiation incident on the front side of the glazing, the reflectance is called the front reflectance R f. The back reflectance Rb (not shown in Figure 6) is the reflectance of the layer for radiation incident on back side b. The transmittance, reflectance, and absorptance of a layer are formally defined as the fractions of incident flux that transmit, reflect, and are absorbed by the layer, respectively, including the effects of interreflection. Their sum equals unity, as shown in Equation (5).
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T+R+A=1
(5)
The layer has a thickness d and is characterized by transmissivity and reflectivity of each of the two surfaces and by the absorptivity of the glazing layer of thickness d. In general, and are characteristics of the interface between the material and the adjacent medium; they may in principle be different for the two surfaces (e.g., for a coated surface, or where a material layer is adjacent to another material rather than air). Physical arguments, however, dictate that T f and T b for the layer be the same (and the f and b superscripts are therefore omitted). R f and R b will be different given similar variations in coatings or adjacent materials, and examination of Equation (5) shows that A f and Ab may be different as well. Uncoated glass has the same front and back properties. Angular Variations. The interfacial properties and , and consequently layer properties T, R, and A, also depend on the incident angle of the radiation incident on the layer. Figure 7 shows the optical properties of common window glass as a function of incidence angle. This variation of properties is small for incident angles below 40° but becomes significant at larger angles. Chapter 14 provides details on calculating the direction and magnitude of solar flux that is incident on a glazing system. Rubin (1985) measured broadband transmittance and reflectances for many uncoated (monolithic) glass substrates, including tinted substrates at various thicknesses. These formed the basis for Figure 8, which compares the properties of glasses of different thickness and composition. As the incident angle increases from zero, transmittance decreases, reflectance increases, and absorptance first increases because of the lengthened optical path and then decreases as more incident radiation is reflected. Although the shapes of the
property curves are superficially similar, note that both the magnitude of the transmittance at normal incidence and the angle at which the transmittance changes significantly vary with glass type and thickness. Wong (2016) normalized the measured solar transmittance (Rubin 1985) for 6 mm substrates. Two distinct sets of regressions coefficients appear to be adequate to represent the angular variation for a wide range of clear and tinted substrates. The result of the normalization is shown in Figure 9. These two distinct sets of regression coefficients were incorporated into the then WINDOW4 program (Finlayson et al. 1994) and all subsequent versions of that software. Note that tinted substrates considered here do not have as strong a spectrally selectively behavior as the high-performance tints described in Carmody et al. (2004). The three curves all have slightly different shapes. For coated glasses or for multiple-pane glazing systems, this difference is more pronounced. One cannot assume that all glazings or glazing systems have a universal angular dependence. ADOPT (1996) is a two-parameter model that describes the angular-dependent optical properties of glazing systems. It has been validated (Karlsson et al. 2001) against actual measurements (Rubin et al. 1999). Models have been developed from the ADOPT project, and validation of its accuracies at discrete angles of incidence had been reported (Hutchins et al. 2001; Rubin et al. 1998). The Simple Window Model (Arasteh et al. 2009) offers correlations for single-, double-, and triple-glazed systems for use in building energy simulation. The normalized transmittance curve is shown in Figure 10 below. Angular performance is important when peak gains and annual energy performance are considered. In North America, peak summertime solar gains occur with east- and west-facing vertical
Fig. 8 Variations with Incident Angle of Solar-Optical Properties for (A) Clear 3 mm Glass, (B) Clear 6 mm Glass, and (C) 6 mm Typical Heat-Absorbing (Tinted) Glass
Fig. 7
Transmittance and Reflectance of Glass Plate
(Refractive index n = 1.55, thickness t = 3.2 mm, absorptance = 0.01/m)
Fig. 9 Normalized Solar Transmittance for Five Common Glass Substrates as Function of Incidence Angle in Degrees
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2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 10 Normalized Solar Transmittance, as Function of Incidence Angle in Degrees, for 10 Glazing Systems (Single-, Double-. and Triple-Pane) as Developed for Simple Window Model (Arasteh et al. 2009)
Fig. 11 Spectral Transmittances of Commercially Available Glazings
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(Kruis 2015)
glazings at angles of incidence ranging from about 25 to 55°. The peak solar gain for horizontal glazings occurs typically at relatively small angles of incidence (midday sun high in sky in summer). For north- and south-facing vertical glazings, peak summertime solar gains occur at angles of incidence greater than about 40°. Angles of incidence important for annual energy performance calculations range from 5° to over 80° for east- and west-facing vertical and for horizontal glazings. This range is only slightly diminished for south-facing glazings. For north-facing glazings, the direct beam solar gains are small and their angles of incidence range from 62 to 86° (McCluney 1994a). Spectral Variations. Many glazing systems have optical properties that are spectrally selective (i.e., they vary across the electromagnetic spectrum with wavelength ). Ordinary clear float glass possesses this property, but to a modest degree that is seldom of much concern in load calculations. Tinted and coated glass can exhibit strong spectral selectivity, a desirable property for certain applications, and this effect must be accounted for in solar heat gain determinations. Figure 11 (McCluney 1993) shows the normal incidence spectral transmittances of several common commercially available glazings. Figure 12 (McCluney 1996) shows the normal incidence spectral transmittances and outdoor reflectances of a variety of additional coated and tinted glasses, indicating the strong spectral selectivity now available from some glass and window manufacturers. Actual transmittance varies with the amount of iron or other absorbers in the glass. Uncoated glass with low iron content has a relatively constant spectral transmittance over the entire solar spectrum. Solar-Optical Property Data. Transmittance and reflectance ar e the basic measurable quantities for an isolated glazing layer in air. Measurements on glazing layers are typically made using a spectrophotometer at normal or near-normal incidence, and the properties at other angles must be inferred from these measurements. A systematic compilation of these measured properties (for most glazings manufactured in the United States) called the International Glazing Database (IGDB) is maintained by the National Fenestration Rating Council and is available on the Internet at www.nfrc.org or from windows.lbl.gov/materials/IGDB/default.htm (LBNL 2016; NFRC 2014b). For uncoated glazings, layer properties can also be determined from first principles [e.g., McCluney (1994b)]. Advances in modeling complex glazing layers (e.g., screens, woven shades, blinds, some types of drapes) have led to the establishment of the Complex Glazing Database (CGDB 2016). Designed to support the LBNL (2016) WINDOW program and similar software, the CGDB contains detailed information on the angular and spectral
Fig. 12 Spectral Transmittances and Reflectances of Strongly Spectrally Selective Commercially Available Glazings (McCluney 1996)
properties of such materials, for total and diffuse transmittance and reflectance. The CGDB is being expanded to include products from many international manufacturers. More information may be found in the section on Shading and Fenestration Attachments. Obtaining the necessary basic information about the solaroptical properties of coated glass requires spectrophotometric measurements. Alternatively, an approximation procedure is described by Finlayson and Arasteh (1993). Coated glazing properties should vary from these estimates by no more than ±20% at 60° incidence (Rubin et al. 1999). It is currently not practical to determine the solar-optical properties of coated glazings from first principles.
Optical Properties of Glazing Systems The optical properties of glazing systems (multiple glazing layers) are affected by interreflections between layers in addition to the specular and angular properties of the individual layers.
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Consequently, the effect of a particular layer on the overall properties may not only depend on its solar-optical properties, but also on its position in the assembly. It is therefore necessary to expand glazing layer considerations to apply to the overall properties of systems and subsystems of glazing layers. The properties of any subsystem can be calculated by using recursion relations (LBNL 2016). Spectral Averaging of Glazing System Properties. The solaroptical properties of a glazing are the wavelength-integrated (or total) transmittance, reflectance, and absorptance of the glazing to incident solar radiation. If the spectral optical properties T(), R(), and A() of the glazing and the spectral irradiance E() incident on the glazing are known, the solar optical properties can be calculated using ASTM Standards E971, E972, and E1084, as well as NFRC Technical Document 300 (NFRC 2014c).
max
E STD X d
min
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X = ---------------------------------------------------
(6)
max
E STD d min
Fig. 13 Solar Spectrum, Human Eye Response Spectrum, Scaled Blackbody Radiation Spectrum, and Idealized Glazing Reflectance Spectrum
X cos d
where X() = T(), R(), or A () ESTD() = standard solar distribution X = total T, R, or A to standard solar distribution
For multiple-layer glazing systems, the spectral averaging should in general be applied to the system spectral properties at each angle. Because all glazing layer properties are to some extent both angle and wavelength dependent, and because these equations are nonlinear in the glazing properties, this is the only procedure that is valid in principle. Many glazings do not have strong spectral selectivity over the solar spectrum, so their spectral optical properties can be considered constant, even if the source spectrum changes substantially. In these cases, the transmitted spectral irradiance can be determined by multiplying the incident irradiance by the solar transmittance obtained with the standard solar distribution. For special combinations of climate and location, it may be desirable to use variant solar spectra as weighting functions, which requires an appropriate spectral solar radiation model (Gueymard 2007). It is still difficult, and seldom necessary, to perform radiative calculations using a detailed, time-dependent solar spectrum and the spectral glazing properties. Such additional calculations might only be justified in the case of glazings with strong spectral selectivity (Gueymard and DuPont 2009). Angular Averaging of Glazing System Properties. It is relatively simple to account for angular dependence in beam solar radiation, because at a given time the radiation is incident from a single, easily determined direction. However, for diffuse solar and ground-reflected radiation, the situation is more complicated. In principle, energy flow through the glazing should equal the sum of individual energy flows caused by incident radiation from each direction. Although such calculations can be done for specific sky conditions using detailed sky data or models, the labor involved is worthwhile only for very specific purposes. Usually, a drastically simplifying assumption is made. Both sky and ground radiation are assumed to be ideally diffuse (i.e., to have a sky radiance that is independent of direction). Diffuse properties are then determined by integrating over all directions. See the section on Diffuse Radiation under Solar Heat Gain Coefficient for more details. In addition, the spectral dependence is assumed to be the same as for beam solar radiation.
hem
X D = -------------------------------------------- = 2 cos d
X cos d
2
(7)
0
hem
where X() = T(), R(), or A() = solid angle of integration XD = total T, R, or A of standard solar distribution
More careful consideration must be given to these quantities for tilted glazings or for direction-dependent shading (e.g., overhangs, venetian blinds) when greater accuracy is desired. Spectrally Selective Glazing and Glazing Systems. Spectrally selective glazing shows strong changes in its optical properties with variations in wavelength over the spectrum. The spectral range from 300 nm to over 50 m contains radiation from both the sun and sky incident on fenestration systems. The majority of this radiation is called short-wave or solar radiation. About 99% of the energy in the solar spectrum is between 0.3 to 3.5 m. The spectral range from 3.5 to over 50 m is called long-wave, infrared, or thermal radiation, which contains radiation from the sun and sky, but also from warm bodies both outside and inside the building. In Figure 13, the solar spectrum for an air mass m = 1.5 (corresponding to a solar altitude of 41.8°) represents short-wave radiation, with thermal radiation represented by a blackbody source at 24°C. The latter has been scaled up to better compare it with the shape of the solar spectrum. The reflectance spectrum shown is an idealization of typical glazing reflectance. Figure 13 clearly shows the separation of the solar spectrum from the long-wave spectrum characteristic of radiant emission from an indoor pane of a multiple-pane glazing system. Figure 13 also shows the human eye spectral response (called the human photopic visibility function). To the human eye, the glazing represented in the figure does not appear very reflective. It is also strongly transmitting for solar radiation, including the visible portion. The glazing is, however, nearly opaque to long-wave radiation, demonstrating that visual perception of a material is a poor indicator of its overall spectral characteristics. The glazing system reflectance depicted in Figure 13 is good for admitting solar radiation while preventing the escape of long-wave radiation emitted by surfaces inside the room, a good design for cold sunny days.
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15.18 Almost all window glass is opaque to long-wave radiation emitted by surfaces at temperatures below about 1200°C. This characteristic produces the greenhouse effect, by which solar radiation passing through fenestration is partially retained indoors by the following mechanism. Radiation absorbed by surfaces in the room is emitted as long-wavelength radiation, which cannot escape directly through the glass because of its opaqueness to radiation beyond 4.5 m. Instead, radiation from room surfaces is absorbed and reemitted to both sides as determined by several parameters, such as the indoor and outdoor film heat transfer coefficients, the surface emissivities, and other glazing properties. A good long-wave reflector can be a poor short-wave reflector and a good short-wave transmitter. Because of the conservation of energy (T + R + A = 1.0), high long-wave reflectance means low transmittance and absorptance. Kirchhoff’s law shows that low absorptance means low emissivity as well. This is the principle of operation of the high-solar-gain (or cold-climate) low-e coating on glass. Such a coating has high transmittance over the entire solar spectrum, producing high solar heat gain while being highly reflective to long-wave infrared radiation emitted by the indoor surfaces, reflecting this radiation inward. The term low-e refers to a low emissivity over the long-wavelength portion of the spectrum. Figure 14 shows hypothetical glazing systems with performance tuned to specific climates. In this case, the sharp reflectance edge that the ideal high-solar-gain, cold-climate low-e coating exhibits just past the end of the solar spectrum in Figure 13 is shifted closer to the edge of the visible portion of the spectrum, thereby increasing the solar near-infrared (NIR) reflectance of the glazing. This results in a drop in the hot-climate transmittance to the right of the visible portion of the spectrum. The effect is to reflect the near-infrared portion of the solar spectrum outdoors, reducing solar gain, while still admitting visible light in the wavelength region below about 0.8 m. This low-solar-gain, hot-climate coating also exhibits low emissivity over the long-wave spectrum, and is therefore also properly termed a low-e coating. To distinguish the cold- from the hot-climate version, a glazing with this type of spectral response is often termed selective low-e. This is something of a misnomer, because both hot- and coldclimate glazings are spectrally selective. Another term is high-solargain, low-e glazing system for cold climates, contrasted with low-solar-gain, low-e glazing system for hot climates. The reduced infrared transmittance for the hot-climate glazing is ideally achieved by high reflectance and low absorptance (meaning also low emissivity). It can also be done with high infrared absorptance, if the flow of absorbed solar radiation to the interior of the
Fig. 14 Demonstration of Two Spectrally Selective Glazing Concepts, Showing Ideal Spectral Transmittances for Glazings Intended for Hot and Cold Climates
2017 ASHRAE Handbook—Fundamentals (SI) building can be reduced, introducing a second approach to the construction of a hot-climate, low-solar-gain glazing system. In this case, the outer pane of a multiple-pane glazing system is made to have good visible transmittance but high absorptance over the solar infrared spectrum. To protect the interior of the building from the heat of this absorbed radiation, additional glazings, gas spaces, and cold-climate or low-solar-gain, low-e coatings are added. By this means, radiation, conduction, and convection of heat from the hot outer pane to the interior ones and to the interior of the building are reduced because of the coating, the gas space, and the additional panes. Such a glazing system for hot climates is insulated primarily not to protect the building from conductive heat losses in winter but to protect the interior from solar radiant heat absorbed by the hot outer pane in summer. Many manufacturers offer this kind of nonreflecting, spectrally selective glazing system for commercial and residential buildings having large cooling loads. Figure 14 shows that glazings intended for hot climates should have (1) high transmittance over the visible portion of the spectrum to let daylight in for both illumination and view and (2) low transmittance over all other portions of the spectrum to reduce solar heat gain. In contrast, glazings intended for very cold climates should have high transmittance over the whole solar spectrum, from 0.38 to over 3.5 m, for maximum admission of solar radiant heat gain and light. In addition, glazings for cold climates should have low transmittance over the long-wavelength portion of the spectrum to block radiant heat emitted by the relatively warm indoor surfaces of buildings, preventing its escape to the outdoors. Extreme spectral selectivity in glazing systems in the visible portion of the spectrum can produce an unwanted color shift in transmitted light. The color of transmitted light and its color-rendering properties should be considered in the design. When discussing spectral behavior of fenestration, it is important to understand the relationship between the solar spectrum and the resulting broadband solar-optical properties of glazing systems. A brief introduction is provided here. Calculation of a glazing system’s SHGC (see the section on Solar Heat Gain Coefficient) requires calculating broadband solar transmittance and absorptance. It is influenced by the solar spectrum and the spectral selectivity achieved through the combination of coatings and substrates that make up the glazing system. These two glazing attributes contribute strongly to spectral variations displayed by glazing systems. To understand any spectral selectivity associated with a glazing layer, it is also necessary to have a good understanding of the spectral distribution of solar radiation. Typically, this is known as a solar spectrum or weighting function, whose wavelengths range from 300 to 2500 nm. Although there is energy at thermal wavelengths much longer than 2500 nm, glass and polymer layers are opaque to such wavelengths. Of available solar energy, the direct normal irradiance (DNI) exhibits the strongest intensity. Accurate representation of the solar spectrum is essential to calculate broadband solar-optical properties of glazings, which in turn are used to calculate SHGC. It can also be used to compare the solar control characteristics of different glazing layers. Gueymard and Kambezidis (2004) reviewed ways to mathematically describe the atmospheric processes of scattering and absorption, to better consider the effect of atmospheric attenuation of extraterrestrial solar irradiation. This enables more realistic modeling of the DNI spectrum. Over the past decade, better understanding of atmospheric phenomena that govern the attenuation of extraterrestrial solar radiation has led to updates of reference solar spectra. Because high-quality observational data have become more readily available, many aspects of the scattering and absorption process can now be modeled with increasing confidence. Reference spectra range from the original ASTM Standard E891-1989 (RA 1992) spectrum as referred to in the WINDOW4 documentation (Finlayson et al. 1993), through
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15.19
the use of ISO Standard 9845 as discussed in Rubin et al. (1998) and finally ASTM Standard G173-2003 (RA 2008) in the latest NFRC (2014c) Technical Document 300-2014. Note that the solar spectrum referenced by NFRC Technical Document 300 (2014c) is identical to that used by the LBNL (2016) WINDOW program for NFRC certification purposes. At present, the DNI spectrum is used to calculate both broadband transmittance and reflectance. Increased availability of high-quality observational data means it is now possible to develop mathematical models of the attenuation of solar radiation caused by scattering and absorption by aerosols and gases in the atmosphere. Several updates of the solar spectra have been published as a result of better consideration of atmospheric phenomena that govern the attenuation of extraterrestrial solar radiation. ASTM Standards G173 and G197 are examples of recent standards based on version 2.9.2 of the Simple Model of the Atmospheric Radiative Transfer of Sunshine (SMARTS) atmospheric transmission code for the photovoltaic industry (Gueymard and Kambezidis 2004). Standards G173 and G197 use angles of tilt (from the horizontal) of 37 and 20°, respectively, which are different from those used in earlier standards such as ASTM Standard E891 and ISO Standard 9845. Standards G173 and G197 are regarded as superior solar spectra because of their ability to consider more atmospheric interactions. They have better resolution (2002 wavelengths of information) than older standards (120 wavelengths). Development and review of newly created solar spectra are ongoing, but there has not yet been a thorough analysis on the influence played by different spectra on the calculated broadband solar transmittance and reflectance. It has been found that some glazing systems (Gueymard and Kambezidis 2004; Lyons et al. 2010), especially those that exhibit strong spectral selectivity, are more sensitive to the choice of solar spectrum compared with glazings having less selective behavior (Gueymard and du Pont 2009). In addition, there has not been any clear guidance on how trapezoidal integration deals with improved resolutions introduced through the newer Standards G173 and G197 spectra (Finlayson et al. 1993; ISO Standard 15099; Rubin et al. 1998). Most spectral measurements for coated and monolithic glazing layers only have 111 and 441 records respectively, as required by the NFRC.
4.2
The SHGC is needed to determine the solar heat gain through a fenestration’s glazing system, and should be included along with Ufactor and other instantaneous performance properties in any manufacturer’s description of a fenestration’s energy performance.
Calculation of Solar Heat Gain Coefficient Because the optical properties A and T vary with the angle of
incidence and wavelength, the solar heat gain coefficient is also a function of these variables. In the most general way, the solar heat gain power per unit area q() and the solar heat gain coefficient SHGC(,) are defined as q = =
ED T , + N A , d ED SHGC , d
ED() = incident solar spectral irradiance T(,) = spectral transmittance of glazing system A(,) = total spectral absorptance of glazing system
Here, the angle- and wavelength-dependent solar heat gain coefficient is given by SHGC(, ) = T(, ) + N A(, )
(8)
in units of energy flux per unit area, W/m2. The inward-flowing fraction is thermal in origin; it depends on heat transfer properties of the assembly rather than on its optical properties. Absorbed solar radiation, including ultraviolet, visible, and infrared radiation from the sun and sky, is turned into heat inside the absorbing material. In fenestration, the glazing system temperature rises as a result to some approximately equilibrium value at which energy gains from absorbed radiation are balanced by equal losses. Absorbed solar radiation is dissipated through conduction, convection, and radiation. Some heat leaves the building, and the remainder goes indoors, adding to the directly transmitted solar radiation. The magnitude of the inward-flowing fraction depends on the nature of the air boundary layers adjacent to both sides of the glazing, including any gas between the panes of a multiple-pane glazing system (Ni is often used to distinguish the inward-flowing fraction from the outward-flowing fraction, No. However, because
(11)
Combined with Equation (6), this becomes the wavelengthaveraged solar heat gain coefficient:
max min
E D T , + N A , d
max
The concept of the solar heat gain coefficient is best shown for the case of a single glass pane in direct sunlight. If ED = EDN cos is the direct solar irradiance incident on the glass, with T the solar transmittance, A the solar absorptance, and N the inward-flowing fraction of the absorbed radiation, then the total solar gain (per unit area) qb that enters the space because of incident solar radiation is
(10)
where
SHGC = -------------------------------------------------------------------------------------------------
SOLAR HEAT GAIN COEFFICIENT
qb = ED(T + N A)
only the inward-flowing fraction is used here, the subscript i is dropped for clarity). The quantity in parenthesis in Equation (8) is called the solar heat gain coefficient (SHGC), also known as g-value or total solar energy transmittance in Europe. The total solar gain power per unit area (from direct beam radiation) can therefore be computed using the following equation: qb = ED SHGC (9)
(12)
ED d min
Equations (10) to (12) indicate the preferred way of determining the solar gain of glazing systems and calculating the solar heat gain coefficient. Computer programs such as LBNL (2016) WINDOW are available to assist in the calculation. In WINDOW, the overall system optical properties at a given incident angle are calculated for each wavelength and the results averaged following Equation (12). The ASTM Standard G173 spectrum for direct irradiance is generally used in the averaging. The wavelength-averaged properties (at a given incident angle) can then be used in Equation (9). This approach has been adopted by the National Fenestration Rating Council in NFRC (2014d) Technical Document 200 for rating, certifying, and labeling fenestration for energy performance and by the Canadian Standards Association (CSA Standard A440.2). The method is valid for strongly spectrally selective (as well as nonselective) glazing systems. When a glazing system is not strongly spectrally selective, the solar-weighted spectral broadband values of the optical properties can be used, and the integral over wavelength shown in Equations (10) and (12) is not needed. In this case, each glazing layer has its own individual inward-flowing fraction of the absorbed radiation for that layer. With the glazings numbered from the outdoors inward, and k the glazing index, the SHGC is given by
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2017 ASHRAE Handbook—Fundamentals (SI)
Nk Ak L
f
SHGC = T +
f
(13)
k=1
where T f = front transmittance of glazing system L = number of glazing layers Akf = absorptance of layer k as part of the composite L-layer system (Finlayson et al 1993; Rubin et al. 1998). This is different from the individual, stand-alone absorptance of layer k. Nk = inward-flowing fraction for layer k
The inward-flowing fractions can be calculated from simplified heat transfer models, using the following equation: N k = U R j – 1 j 1
(14)
j =k
where
Rj–1,j = sum of resistances from exterior air node up to center of kth Licensed for single user. © 2017 ASHRAE, Inc.
glazing layer
Thus, the inward-flowing fraction of layer k is essentially the Ufactor of the fenestration times the thermal resistance from the center of the kth layer to the outdoors. In more complicated multilayer glazing systems, it is advisable to perform a detailed heat transfer analysis of the system to determine the values of Nk, because the effective heat transfer coefficients and U depend (weakly) on the glazing layer temperatures and other environmental conditions [e.g., Finlayson et al. (1993), LBNL (2016), Wright (1995a)].
Diffuse Radiation For incident diffuse radiation, the hemispherical average solar heat gain coefficient must be used. This may be calculated by combining Equation (12) with Equation (7) as follows:
SHGC cos d
SHGC D = --------------------------------------------------------cos d hem
0
hem
Fig. 15 Components of Solar Radiant Heat Gain with Double-Pane Fenestration, Including Both Frame and Glazing Contributions The solar heat gain coefficient of the fenestration system can be calculated while accounting for solar gain through the opaque elements by area-weighting the solar heat gain coefficients of the glazing, frame, and M divider elements. Thus, SHGC g A g + SHGCf A f + A i SHGC i i=1 SHGC = ------------------------------------------------------------------------------------------------M Ag + Af + Ai M
i=1
where SHGCg , SHGCf , and SHGCi are the solar heat gain coefficients of the glazed area, frame, and ith divider, respectively. Ag, Af , and Ai are the corresponding projected areas. In some cases, it is useful to have an overall SHGC for the opaque elements only, which is defined by SHGCf Af + A i SHGC i M
(15)
2
= 2
i=1
SHGC op = ----------------------------------------------------------A op
SHGC cos d
Equivalently, T and A in Equation (13) can be hemispherically averaged using Equation (7) so that SHGC D = T + f
D
Nk A k D L
f
(16)
k=1
In any case, Nk is unaffected in averaging, because it does not depend on incident angle or wavelength. In contrast, T and A do depend on wavelength. The spectral distribution of diffuse light is notably different from that of direct radiation. However, under usual clear-sky conditions, the fractional amount of diffuse radiation is relatively low compared to that of its direct counterpart, so that the error introduced by neglecting this difference is low in general.
Solar Gain Through Frame and Other Opaque Elements Figure 15 shows the mechanisms by which fenestration provides solar gain. It is assumed that all of the directly transmitted solar radiation is absorbed at indoor surfaces, where it is converted to heat. Solar gain also enters a building through opaque elements such as the frame and any mullion or dividers that are part of the fenestration system, because part of the solar energy absorbed at the surfaces of these elements is redirected to the indoor side by heat transfer.
(17)
where
(18)
A op = A f + A i M
i=1
SHGCf can be estimated (ISO Standard 15099:2003; Wright 1995b) using Af s Uf SHGC f = f ----- ------------ hf A surf
(19)
s where f is the solar absorptance of the outdoor surface of the frame, Uf is the frame U-factor, and hf is the heat transfer coefficient (radiative plus convective) between the frame and the outdoor environment. The projected-to-surface area ratio (Af /Asurf ) corrects for the fact that Uf is based on projected area Af , whereas hf is based on the exposed outdoor frame surface area Asurf . SHGCi can be calculated in the same way: s Ui Ai SHGC i = i ----- --------------- hi Asurf, i
(20)
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15.21
The outdoor-side heat transfer coefficients hf and hi can be estimated using 3
h f or h i = h co + 4e f t out
(21)
where hco is the convective heat transfer coefficient between the frame (or divider) surface and the outdoor environment, ef is the emissivity (long-wave) of the outdoor frame (or divider) surface, tout is the outdoor temperature, and is the Stefan-Boltzmann constant.
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Solar Heat Gain Coefficient, Visible Transmittance, and Spectrally Averaged Solar-Optical Property Values Table 10 lists visible transmittance, solar transmittance, front and back reflectance, and solar heat gain coefficients for common glazing and window systems. The ID number for each entry in Table 10 refers to an ID number in Table 4, and the window systems therefore include windows with aluminum or metal frames and windows with other frames that have a lower conductivity (e.g., thermally broken aluminum, wood, vinyl, and fiberglass). As can be seen in Table 10, total window solar heat gain coefficient varies with type of operator, size of fenestration product, and type of frame. The glazing Tv , Tsol , R f, Rb, and SHGC values have been calculated using manufacturers’ spectral data following methods described in Finlayson and Arasteh (1993) and Wright (1995b), and using the 1.5 air mass spectrum found in ASTM Standard G173. Glazing values are given for 3 and 6 mm glass and vary with glass thickness and manufacturer. Values shown are averages and may vary by ±0.05. It is recommended that actual values be determined using detailed spectral data from NFRC (2014b). The front reflectance is that of the unit to the outdoors, and the back reflectance is that to the room side. Visible transmittances are provided in Table 10 for center-glazing values at normal incidence and for total window values at normal incidence. A rule of thumb is to select a glazing unit whose visible transmittance is 1.5 to 2.0 times greater than its solar heat gain coefficient, especially if daylighting strategies will be used in the building. For maximum light with minimum solar gain, some fenestration products have visible transmittances 2.5 times their SHGCs. For energy calculations on a daylit building, visible transmittance for the entire window should be used. The visible transmittance of a window can be calculated by multiplying the fraction of glazing area by the center-glazing visible transmittance. Solar heat gain coefficients are provided for center-glazing and total window values. Center-glazing solar heat gain coefficients are given at normal incidence (0°) and at 40°, 50°, 60°, 70°, and 80° incidence angles. For angles other than those listed, straight-line interpolation can be used between the two closest angles for which values are shown. Total window solar heat gain coefficients assume normal incidence. The operable and fixed window sizes in Table 4 were used. To calculate the frame area, frame heights shown in Figure 4 for aluminum and aluminum-clad wood/wood/vinyl were used. The frame area for aluminum windows is 11% for operable size, and 10% for fixed. The frame area for other frames is 20% for the operable size and 12% for fixed. The ratio of projected frame dimension (PFD) to frame surface area is assumed to be 1.0, based on Wright (1995b). In reality, for most frames this ratio is in the range 0.7 to 0.9, because frames cannot simultaneously be flush with both outdoor and indoor glass surfaces. The PFD is defined by the sight line, which is the intersection of the opaque frame (which could be indoors or outdoors) and the vision area. Frame solar heat gain coefficients used to determine the total window solar heat gain coefficients are calculated according to the section on Solar Gain Through Frame and Other Opaque Elements. Frame U-factors are taken from Table 1. Frame absorptance is assumed to be 0.5 (this is the default NFRC value for commercial frames; however, note that NFRC procedures stipulate 0.3 for
residential frames). The outdoor convective film coefficient is 15 W/(m2 ·K), corresponding to a wind speed of 2.8 m/s. For the aluminum window, the frame solar heat gain coefficient is 0.14 for the operable window and 0.11 for the fixed. For the other frames, the frame solar heat gain coefficient varies between 0.02 and 0.07 for the various lower-conductivity frame types. A frame solar heat gain coefficient of 0.04 is used for the operable window, and 0.03 for the fixed. These values correspond directly to the aluminumclad wood/reinforced vinyl frames. Solar transmittances and front and back reflectances are also center-glazing values and are given at normal incidence (0°) and at 40°, 50°, 60°, 70°, and 80° incidence angles. The effective inwardflowing fraction of absorbed radiation for the entire system (not layer-specific values) can be determined from Equation (9) by inserting the solar transmittance and corresponding SHGC. Example 5. Estimate the overall visible light transmittance for an operable wood casement window with clear, uncoated 6 mm double glazing. The operable window has 27% frame area with a wood frame. Solution: The center-glazing visible light transmittance is 0.78 (see Table 10, glazing ID = 5b, first column). The overall visible light transmittance is Tv = 0.27(0) + 0.73(0.78) = 0.57
Airflow Windows If properly managed, airflow between panes of a double-glazed window can improve fenestration performance. In normal use, a venetian blind is located between the glazing layers. Ventilation air from the room enters the double-glazed cavity, flows over the blind, and can be exhausted from the building or returned through the ducts to the central HVAC system. These systems can control window heat transfer under many different operating conditions. During sunny winter days, the blind acts as a solar air collector; heat removed by the moving air can be used elsewhere in the building. Furthermore, the window acts as a heat exchanger when sunlit so that the indoor glazing temperature nearly equals the room air temperature and improves thermal comfort (also see the section on Occupant Comfort and Acceptance). In the summer, the window can have a very low solar heat gain coefficient if the blinds are appropriately placed, because the majority of solar gains are removed from the window. Brandle and Boehm (1982) and Sodergren and Bostrom (1971) give details on airflow windows.
Skylights Skylight solar heat gain strongly depends on the configuration of the space below or adjacent to (i.e., in sloped applications) the skylight formed by the skylight curb and any associated light well. Five aspects must be considered: (1) transmittance and absorptance of the skylight unit, (2) transmitted solar flux that reaches the aperture of the light well, (3) whether that aperture is covered by a diffuser, (4) transmitted solar flux that strikes the walls of the light well, and (5) reflectance of the walls of the light well. Data for flat skylights, which may be considered as sloped glazings, are found in Tables 4 and 11. The following categories of skylights all admit daylight from the roof plane but differ markedly in their design. Domed Skylights. Solar and total heat gains for domed skylights can be determined by the same procedure used for windows. Table 11 gives SHGCs for plastic domed skylights at normal incidence (Shutrum and Ozisik 1961). Manufacturers’ literature has further details. Given the poorly defined incident angle conditions for domed skylights, it is best to use these values without correction for incident angle, together with the correct (angle-dependent) value of incident solar irradiance. Results should be considered approximate. In the absence of other data, these values may also be used to make estimates for skylights on slanted roofs.
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2017 ASHRAE Handbook—Fundamentals (SI)
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems
Operable
0.78 0.75 0.14 0.14 0.11 0.73 0.68 0.13 0.13 0.19 0.64 0.55 0.12 0.12 0.33 0.53 0.39 0.11 0.62 0.51 0.62 0.52 0.12 0.12 0.37 0.52 0.38 0.11 0.11 0.51 0.61 0.51 0.12 0.12 0.37 0.51 0.36 0.10 0.10 0.54 0.54 0.40 0.11 0.11 0.49
0.67 0.64 0.25 0.25 0.11 0.62 0.58 0.24 0.24 0.19 0.55 0.46 0.22 0.22 0.33 0.45 0.32 0.19 0.53 0.49 0.53 0.43 0.21 0.21 0.36 0.45 0.32 0.20 0.20 0.49 0.53 0.42 0.21 0.21 0.37 0.44 0.29 0.19 0.19 0.52 0.46 0.33 0.20 0.20 0.47
0.42 0.39 0.51 0.51 0.11 0.39 0.35 0.48 0.48 0.17 0.34 0.27 0.45 0.45 0.29 0.29 0.18 0.42 0.33 0.41 0.33 0.25 0.45 0.45 0.31 0.29 0.18 0.42 0.42 0.40 0.33 0.24 0.44 0.44 0.32 0.28 0.16 0.41 0.41 0.43 0.30 0.19 0.43 0.43 0.38
0.78 0.75 0.14 0.14 0.10 0.73 0.69 0.13 0.13 0.17 0.65 0.56 0.12 0.12 0.31 0.54 0.41 0.10 0.10 0.48 0.63 0.53 0.11 0.11 0.35 0.54 0.40 0.10 0.10 0.49 0.63 0.53 0.11 0.11 0.35 0.52 0.38 0.10 0.10 0.51 0.55 0.42 0.11 0.11 0.48
0.78 0.79 0.70 0.76
0.80 0.81 0.72 0.79
0.74 0.74 0.66 0.72
0.78 0.79 0.70 0.77
0.67 0.67 0.59 0.65
0.61 0.61 0.54 0.60
0.57 0.57 0.50 0.55
0.48 0.49 0.43 0.48
0.64 0.64 0.57 0.62
0.73 0.74 0.66 0.72
0.55 0.55 0.49 0.53
0.68 0.68 0.61 0.67
0.64 0.64 0.57 0.62
0.55 0.56 0.50 0.55
0.54 0.54 0.48 0.52
0.41 0.41 0.37 0.40
0.57 0.57 0.50 0.55
0.67 0.68 0.60 0.66
0.19 0.06 0.33 0.50 0.61 0.25 0.11 0.26 0.44 0.63
0.19 0.06 0.34 0.50 0.61 0.25 0.10 0.27 0.44 0.63
0.19 0.06 0.35 0.51 0.60 0.24 0.10 0.28 0.45 0.62
0.18 0.05 0.37 0.53 0.58 0.23 0.09 0.31 0.47 0.60
0.16 0.04 0.44 0.58 0.52 0.20 0.07 0.38 0.52 0.55
0.10 0.03 0.61 0.71 0.37 0.13 0.04 0.57 0.67 0.39
0.18 0.05 0.36 0.52 0.57 0.23 0.09 0.30 0.46 0.60
0.18 0.18 0.16 0.17
0.07 0.07 0.06 0.07
0.24 0.24 0.21 0.22
0.12 0.13 0.11 0.12
Fixed
0.82 0.80 0.10 0.10 0.10 0.78 0.73 0.09 0.09 0.18 0.68 0.59 0.08 0.08 0.32 0.57 0.43 0.07 0.66 0.50 0.66 0.56 0.08 0.08 0.36 0.56 0.42 0.07 0.07 0.51 0.66 0.56 0.08 0.08 0.37 0.55 0.40 0.07 0.07 0.54 0.57 0.44 0.07 0.07 0.49
Fixed
Hemis., Diffuse
0.84 0.82 0.08 0.08 0.10 0.80 0.75 0.08 0.08 0.17 0.71 0.62 0.07 0.07 0.31 0.59 0.45 0.06 0.68 0.49 0.68 0.58 0.07 0.07 0.35 0.58 0.44 0.06 0.06 0.50 0.68 0.58 0.07 0.07 0.36 0.57 0.42 0.06 0.06 0.52 0.59 0.46 0.06 0.06 0.48
Fixed
80.00
Operable
70.00
0.86 0.83 0.08 0.08 0.09 0.81 0.77 0.07 0.07 0.16 0.73 0.65 0.06 0.06 0.29 0.62 0.49 0.05 0.05 0.46 0.70 0.61 0.06 0.06 0.33 0.60 0.47 0.05 0.05 0.47 0.70 0.61 0.06 0.06 0.33 0.59 0.46 0.05 0.05 0.49 0.62 0.49 0.06 0.06 0.45
Fixed
60.00
Operable
50.00
Center Glazing Tv
Other Aluminum Frames
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
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Uncoated Single Glazing 1a 3
CLR
0.90
1b 6
CLR
0.88
1c 3
BRZ
0.68
1d 6
BRZ
0.54
1e 3
GRN
0.82
1f 6
GRN
0.76
1g 3
GRY
0.62
1h 6
GRY
0.46
1i 6
BLUGRN
0.75
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f Reflective Single Glazing 1j 6
SS on CLR 8%
0.08
1k 6
SS on CLR 14%
0.14
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
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Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
SS on GRN 14%
0.12
1n 6
TI on CLR 20%
0.20
1o 6
TI on CLR 30%
0.30
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
SHGC T Rf Rb
A1f
Operable
Operable
Operable
0.28 0.13 0.26 0.41 0.61 0.23 0.06 0.19 0.47 0.76 0.27 0.12 0.26 0.44 0.62 0.35 0.19 0.20 0.36 0.62
0.24 0.11 0.34 0.48 0.56 0.21 0.04 0.27 0.52 0.68 0.23 0.09 0.34 0.50 0.57 0.30 0.16 0.28 0.43 0.57
0.16 0.06 0.54 0.64 0.40 0.14 0.03 0.49 0.67 0.48 0.15 0.06 0.54 0.65 0.40 0.20 0.09 0.50 0.60 0.40
0.28 0.13 0.25 0.41 0.60 0.23 0.06 0.18 0.46 0.75 0.27 0.12 0.26 0.43 0.62 0.35 0.20 0.19 0.36 0.62
0.29 0.29 0.26 0.28
0.18 0.18 0.16 0.18
0.24 0.24 0.21 0.22
0.11 0.11 0.10 0.11
0.27 0.27 0.24 0.26
0.18 0.18 0.16 0.18
0.36 0.36 0.32 0.35
0.27 0.27 0.24 0.26
0.76 0.70 0.13 0.13 0.10 0.07 0.70 0.61 0.11 0.11 0.17 0.11 0.62 0.55 0.09 0.12 0.30 0.06 0.49 0.38 0.07 0.10 0.48 0.07 0.60 0.52 0.09 0.12 0.34 0.05 0.49 0.39 0.08 0.10 0.49 0.05
0.74 0.68 0.14 0.14 0.11 0.08 0.67 0.58 0.12 0.12 0.18 0.12 0.60 0.51 0.10 0.13 0.33 0.06 0.46 0.35 0.08 0.11 0.51 0.07 0.57 0.49 0.10 0.13 0.37 0.05 0.46 0.36 0.08 0.11 0.51 0.05
0.71 0.65 0.16 0.16 0.11 0.08 0.64 0.55 0.15 0.15 0.19 0.12 0.57 0.48 0.12 0.15 0.34 0.06 0.44 0.32 0.09 0.13 0.52 0.07 0.54 0.46 0.12 0.15 0.38 0.05 0.44 0.33 0.10 0.13 0.05 0.05
0.64 0.58 0.23 0.23 0.12 0.08 0.58 0.48 0.20 0.20 0.20 0.12 0.51 0.42 0.16 0.21 0.36 0.06 0.39 0.27 0.13 0.19 0.53 0.07 0.49 0.40 0.16 0.21 0.39 0.04 0.39 0.29 0.14 0.19 0.53 0.05
0.50 0.44 0.36 0.36 0.13 0.07 0.45 0.36 0.33 0.33 0.21 0.10 0.39 0.31 0.27 0.35 0.37 0.05 0.31 0.20 0.22 0.31 0.53 0.06 0.38 0.30 0.27 0.35 0.39 0.04 0.31 0.21 0.23 0.31 0.52 0.04
0.26 0.21 0.61 0.61 0.13 0.05 0.23 0.17 0.57 0.57 0.20 0.07 0.20 0.14 0.49 0.59 0.34 0.03 0.17 0.08 0.44 0.55 0.45 0.04 0.20 0.13 0.50 0.60 0.35 0.03 0.17 0.09 0.45 0.55 0.43 0.03
0.66 0.60 0.21 0.21 0.11 0.07 0.60 0.51 0.18 0.18 0.19 0.11 0.53 0.45 0.15 0.19 0.33 0.06 0.41 0.30 0.12 0.17 0.50 0.07 0.51 0.43 0.15 0.19 0.37 0.04 0.41 0.31 0.13 0.17 0.50 0.05
0.69 0.70 0.62 0.67
0.72 0.73 0.65 0.71
0.64 0.64 0.57 0.62
0.69 0.70 0.62 0.69
0.57 0.57 0.50 0.55
0.55 0.56 0.50 0.55
0.45 0.45 0.40 0.43
0.42 0.42 0.38 0.41
0.55 0.55 0.49 0.53
0.67 0.68 0.60 0.66
0.45 0.45 0.40 0.43
0.61 0.61 0.54 0.60
Fixed
0.30 0.14 0.23 0.39 0.63 0.24 0.06 0.16 0.45 0.78 0.28 0.13 0.24 0.42 0.64 0.37 0.21 0.17 0.34 0.64
Fixed
0.30 0.15 0.22 0.38 0.64 0.25 0.06 0.14 0.44 0.80 0.29 0.13 0.22 0.40 0.65 0.38 0.22 0.15 0.33 0.65
Fixed
0.31 0.15 0.21 0.38 0.64 0.25 0.06 0.14 0.44 0.80 0.29 0.14 0.22 0.40 0.65 0.39 0.23 0.15 0.32 0.63
Fixed
Hemis., Diffuse
1m 6
SHGC T Rf Rb
80.00
0.20
70.00
SS on CLR 20%
60.00
1l 6
50.00
Center Glazing Tv
Other Aluminum Frames
40.00
Licensed for single user. © 2017 ASHRAE, Inc.
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
Uncoated Double Glazing 5a 3
5b 6
5c 3
5d 6
5e 3
5f 6
CLR CLR
CLR CLR
BRZ CLR
BRZ CLR
GRN CLR
GRN CLR
0.81
0.78
0.62
0.47
0.75
0.68
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.24
2017 ASHRAE Handbook—Fundamentals (SI)
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
0.59
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
0.51 0.42 0.14 0.19 0.37 0.05 0.39 0.28 0.12 0.17 0.53 0.06 0.43 0.32 0.13 0.17 0.48 0.05 0.33 0.22 0.11 0.17 0.63 0.03
0.13 0.05 0.33 0.38 0.61 0.01 0.17 0.08 0.26 0.34 0.63 0.02 0.22 0.12 0.21 0.30 0.64 0.03 0.16 0.05 0.14 0.34 0.80 0.01 0.21 0.11 0.22 0.32 0.65 0.02 0.29 0.18 0.15 0.27 0.64 0.04
0.12 0.05 0.34 0.37 0.61 0.01 0.17 0.08 0.27 0.33 0.64 0.02 0.21 0.11 0.22 0.30 0.64 0.03 0.16 0.05 0.14 0.33 0.80 0.01 0.20 0.10 0.22 0.31 0.66 0.02 0.28 0.17 0.15 0.27 0.64 0.04
0.12 0.04 0.35 0.38 0.60 0.01 0.16 0.08 0.28 0.34 0.64 0.02 0.21 0.11 0.23 0.31 0.63 0.03 0.15 0.05 0.16 0.34 0.79 0.01 0.19 0.10 0.24 0.32 0.65 0.02 0.27 0.16 0.17 0.28 0.63 0.04
0.11 0.04 0.37 0.40 0.58 0.01 0.15 0.07 0.31 0.37 0.63 0.02 0.19 0.09 0.26 0.34 0.62 0.03 0.14 0.04 0.19 0.37 0.76 0.01 0.18 0.08 0.27 0.35 0.63 0.02 0.25 0.14 0.20 0.31 0.62 0.04
0.10 0.03 0.44 0.46 0.53 0.01 0.13 0.05 0.38 0.44 0.61 0.02 0.16 0.07 0.34 0.41 0.57 0.02 0.12 0.03 0.27 0.44 0.69 0.01 0.15 0.06 0.34 0.42 0.58 0.02 0.20 0.10 0.29 0.40 0.58 0.03
0.06 0.01 0.61 0.61 0.37 0.01 0.08 0.02 0.57 0.60 0.56 0.02 0.09 0.03 0.54 0.59 0.41 0.02 0.08 0.01 0.49 0.60 0.49 0.01 0.09 0.03 0.54 0.59 0.41 0.01 0.12 0.05 0.51 0.58 0.43 0.02
0.11 0.04 0.37 0.40 0.56 0.01 0.16 0.07 0.30 0.36 0.60 0.02 0.20 0.10 0.25 0.33 0.61 0.03 0.14 0.04 0.18 0.36 0.76 0.01 0.18 0.09 0.26 0.35 0.62 0.02 0.25 0.15 0.19 0.31 0.61 0.04
Fixed
0.20 0.12 0.48 0.59 0.37 0.03 0.16 0.07 0.43 0.55 0.47 0.03 0.17 0.10 0.46 0.55 0.42 0.03 0.14 0.06 0.43 0.55 0.50 0.01
Operable
0.37 0.29 0.26 0.34 0.41 0.05 0.29 0.18 0.21 0.31 0.56 0.05 0.32 0.22 0.24 0.31 0.50 0.04 0.25 0.15 0.21 0.31 0.62 0.02
Other Aluminum Frames Fixed
0.48 0.39 0.16 0.21 0.40 0.05 0.37 0.25 0.12 0.18 0.57 0.06 0.40 0.30 0.14 0.19 0.51 0.05 0.31 0.20 0.12 0.19 0.65 0.03
Fixed
0.54 0.45 0.11 0.15 0.39 0.06 0.42 0.29 0.08 0.13 0.56 0.07 0.45 0.34 0.10 0.14 0.50 0.06 0.35 0.24 0.08 0.13 0.65 0.03
Operable
0.57 0.48 0.09 0.13 0.37 0.06 0.44 0.32 0.07 0.11 0.54 0.07 0.47 0.37 0.08 0.11 0.49 0.06 0.37 0.26 0.07 0.11 0.65 0.03
Fixed
0.60 0.51 0.09 0.12 0.34 0.05 0.47 0.36 0.07 0.10 0.51 0.07 0.50 0.40 0.08 0.11 0.47 0.06 0.39 0.28 0.06 0.10 0.62 0.03
Operable
Hemis., Diffuse
HI-P GRN CLR
0.67
A1f A f2
80.00
5j 6
BLUGRN CLR
0.41
SHGC T Rf Rb
70.00
Licensed for single user. © 2017 ASHRAE, Inc.
5i 6
GRY CLR
0.56
60.00
5h 6
GRY CLR
50.00
5g 3
Center Glazing Tv
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
0.55 0.55 0.49 0.53
0.50 0.50 0.45 0.49
0.43 0.43 0.38 0.42
0.36 0.37 0.33 0.36
0.46 0.46 0.41 0.44
0.60 0.60 0.54 0.59
0.36 0.36 0.32 0.35
0.53 0.53 0.47 0.52
0.13 0.13 0.11 0.12
0.06 0.06 0.06 0.06
0.17 0.16 0.14 0.15
0.12 0.12 0.10 0.11
0.21 0.21 0.18 0.20
0.16 0.16 0.14 0.16
0.16 0.16 0.14 0.14
0.10 0.10 0.09 0.10
0.20 0.20 0.18 0.19
0.16 0.16 0.14 0.16
0.27 0.27 0.24 0.26
0.24 0.24 0.22 0.24
Reflective Double Glazing 5k 6
5l 6
5m 6
5n 6
5o 6
5p 6
SS on CLR 8%, CLR
SS on CLR 14%, CLR
SS on CLR 20%, CLR
SS on GRN 14%, CLR
TI on CLR 20%, CLR
TI on CLR 30%, CLR
0.07
0.13
0.18
0.11
0.18
0.27
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Fenestration
15.25
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
Operable
0.56 0.48 0.24 0.26 0.21 0.08 0.51 0.41 0.22 0.23 0.26 0.11
0.43 0.36 0.37 0.38 0.20 0.07 0.40 0.30 0.35 0.35 0.25 0.10
0.23 0.18 0.61 0.61 0.16 0.05 0.21 0.14 0.59 0.57 0.19 0.07
0.57 0.50 0.22 0.24 0.20 0.07 0.53 0.43 0.21 0.22 0.25 0.10
0.59 0.60 0.53 0.58
0.68 0.68 0.61 0.67
0.55 0.55 0.49 0.53
0.65 0.66 0.58 0.64
0.70 0.59 0.17 0.15 0.11 0.14 0.65 0.51 0.15 0.14 0.17 0.17 0.57 0.46 0.12 0.14 0.31 0.11 0.45 0.33 0.09 0.13 0.48 0.11 0.55 0.44 0.11 0.14 0.35 0.11 0.41 0.29 0.08 0.13 0.53 0.10 0.54 0.43 0.11 0.14 0.36 0.10
0.68 0.56 0.18 0.16 0.12 0.14 0.63 0.48 0.16 0.15 0.19 0.17 0.54 0.43 0.12 0.15 0.34 0.11 0.42 0.30 0.09 0.14 0.51 0.11 0.52 0.41 0.11 0.15 0.38 0.10 0.39 0.26 0.08 0.14 0.57 0.09 0.51 0.40 0.11 0.15 0.39 0.10
0.65 0.54 0.20 0.18 0.13 0.14 0.60 0.46 0.18 0.17 0.20 0.17 0.51 0.41 0.14 0.17 0.35 0.10 0.40 0.28 0.10 0.16 0.52 0.10 0.50 0.38 0.13 0.17 0.39 0.10 0.36 0.24 0.09 0.16 0.58 0.09 0.49 0.38 0.13 0.17 0.40 0.10
0.59 0.48 0.26 0.24 0.13 0.13 0.54 0.41 0.23 0.22 0.21 0.15 0.46 0.36 0.18 0.23 0.37 0.10 0.35 0.24 0.14 0.21 0.54 0.09 0.44 0.33 0.17 0.23 0.41 0.09 0.32 0.21 0.13 0.21 0.59 0.08 0.44 0.33 0.17 0.22 0.42 0.09
0.46 0.36 0.38 0.37 0.14 0.11 0.42 0.30 0.35 0.35 0.22 0.13 0.35 0.26 0.28 0.35 0.38 0.08 0.27 0.17 0.23 0.34 0.53 0.07 0.34 0.24 0.27 0.35 0.42 0.07 0.25 0.15 0.22 0.34 0.58 0.06 0.33 0.24 0.27 0.35 0.42 0.07
0.24 0.18 0.61 0.61 0.15 0.07 0.21 0.14 0.57 0.59 0.22 0.07 0.18 0.12 0.50 0.60 0.35 0.04 0.14 0.07 0.44 0.58 0.45 0.04 0.17 0.11 0.48 0.60 0.37 0.04 0.13 0.06 0.43 0.58 0.48 0.03 0.17 0.11 0.48 0.60 0.38 0.04
0.61 0.50 0.24 0.22 0.12 0.13 0.56 0.43 0.22 0.21 0.19 0.16 0.48 0.38 0.17 0.21 0.34 0.10 0.38 0.26 0.13 0.20 0.50 0.09 0.46 0.36 0.16 0.21 0.38 0.09 0.34 0.23 0.13 0.20 0.56 0.08 0.46 0.35 0.16 0.21 0.39 0.09
0.64 0.64 0.57 0.62
0.68 0.68 0.1 0.67
0.59 0.60 0.53 0.58
0.65 0.66 0.58 0.64
0.52 0.52 0.46 0.51
0.52 0.52 0.46 0.51
0.42 0.42 0.37 0.40
0.40 0.41 0.36 0.40
0.50 0.51 0.45 0.49
0.62 0.63 0.56 0.62
0.38 0.38 0.34 0.36
0.54 0.55 0.49 0.54
0.50 0.50 0.44 0.48
0.47 0.48 0.42 0.47
Fixed
0.61 0.54 0.18 0.20 0.21 0.08 0.57 0.46 0.17 0.18 0.26 0.11
Fixed
Hemis., Diffuse
0.64 0.56 0.16 0.18 0.21 0.07 0.59 0.48 0.15 0.16 0.26 0.11
Fixed
80.00
Operable
70.00
0.65 0.59 0.15 0.17 0.20 0.07 0.60 0.51 0.14 0.15 0.26 0.10
Fixed
60.00
Operable
50.00
Center Glazing Tv
Other Aluminum Frames
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
Low-e Double Glazing, e = 0.2 on surface 2 17a 3
Licensed for single user. © 2017 ASHRAE, Inc.
17b 6
LE CLR
LE CLR
0.76
0.73
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2 Low-e Double Glazing, e = 0.2 on surface 3 17c 3
17d 6
17e 3
17f 6
17g 3
17h 6
17i 3
CLR LE
CLR LE
BRZ LE
BRZ LE
GRN LE
GRN LE
GRY LE
0.76
0.73
0.58
0.45
0.70
0.61
0.53
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.26
2017 ASHRAE Handbook—Fundamentals (SI)
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
0.55
SHGC T Rf Rb
A1f A f2
0.241 T Rf Rb
A1f A f2 Low-e Double Glazing, e = 0.1 on surface 2 21a 3 LE CLR 0.76
21b 6
LE CLR
0.72
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2 Low-e Double Glazing, e = 0.1 on surface 3 21c 3 CLR LE 0.75
21d 6
21e 3
21f 6
CLR LE
BRZ LE
BRZ LE
0.72
0.57
0.45
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
0.33 0.21 0.14 0.20 0.56 0.08 0.37 0.26 0.13 0.20 0.51 0.09 0.28 0.17 0.11 0.20 0.65 0.06
0.65 0.59 0.15 0.17 0.20 0.07 0.60 0.51 0.14 0.15 0.26 0.10
0.64 0.56 0.16 0.18 0.21 0.07 0.59 0.48 0.15 0.16 0.26 0.11
0.62 0.54 0.18 0.20 0.21 0.08 0.57 0.46 0.17 0.18 0.26 0.11
0.56 0.48 0.24 0.26 0.21 0.08 0.51 0.41 0.22 0.23 0.26 0.11
0.43 0.36 0.37 0.38 0.20 0.07 0.40 0.30 0.35 0.35 0.25 0.10
0.23 0.18 0.61 0.61 0.16 0.05 0.21 0.14 0.59 0.57 0.19 0.07
0.60 0.48 0.26 0.24 0.12 0.14 0.56 0.42 0.24 0.20 0.19 0.16 0.48 0.37 0.18 0.23 0.34 0.11 0.39 0.27 0.12 0.19 0.51 0.10
0.58 0.45 0.27 0.24 0.13 0.15 0.55 0.40 0.24 0.20 0.20 0.17 0.46 0.34 0.17 0.23 0.37 0.12 0.37 0.24 0.12 0.20 0.54 0.10
0.56 0.43 0.28 0.26 0.14 0.15 0.52 0.37 0.25 0.22 0.21 0.17 0.44 0.32 0.19 0.25 0.38 0.12 0.35 0.22 0.13 0.22 0.55 0.10
0.51 0.37 0.32 0.29 0.14 0.16 0.48 0.32 0.29 0.26 0.22 0.17 0.40 0.27 0.22 0.29 0.39 0.12 0.31 0.19 0.16 0.25 0.56 0.10
0.40 0.27 0.42 0.38 0.15 0.16 0.38 0.24 0.38 0.34 0.23 0.16 0.31 0.20 0.30 0.37 0.39 0.11 0.24 0.13 0.24 0.34 0.55 0.09
0.22 0.13 0.62 0.58 0.15 0.10 0.20 0.11 0.58 0.55 0.22 0.10 0.17 0.08 0.50 0.57 0.35 0.07 0.13 0.05 0.44 0.55 0.46 0.05
Fixed
0.13 0.06 0.44 0.58 0.48 0.03 0.14 0.07 0.44 0.58 0.45 0.03 0.11 0.04 0.41 0.58 0.53 0.02
Operable
0.24 0.14 0.23 0.34 0.58 0.06 0.27 0.17 0.23 0.34 0.54 0.07 0.20 0.10 0.20 0.33 0.66 0.04
Other Aluminum Frames Fixed
0.31 0.20 0.14 0.22 0.59 0.07 0.35 0.23 0.14 0.21 0.54 0.09 0.26 0.15 0.11 0.21 0.68 0.06
Fixed
0.35 0.23 0.11 0.16 0.59 0.08 0.40 0.27 0.10 0.16 0.53 0.10 0.30 0.18 0.08 0.16 0.68 0.06
Operable
0.37 0.25 0.09 0.14 0.58 0.09 0.42 0.29 0.09 0.14 0.51 0.10 0.31 0.19 0.07 0.14 0.67 0.07
Fixed
0.39 0.27 0.09 0.13 0.55 0.09 0.45 0.32 0.09 0.13 0.48 0.11 0.34 0.22 0.07 0.13 0.64 0.08
Operable
Hemis., Diffuse
HI-P GRN LE
A1f A f2
80.00
6
0.62
SHGC T Rf Rb
70.00
Licensed for single user. © 2017 ASHRAE, Inc.
17l
BLUGRN LE
0.37
60.00
17k 6
GRY LE
50.00
17j 6
Center Glazing Tv
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
0.36 0.36 0.32 0.35
0.33 0.33 0.30 0.33
0.42 0.42 0.37 0.40
0.55 0.56 0.50 0.55
0.32 0.32 0.28 0.30
0.49 0.50 0.44 0.48
0.57 0.50 0.22 0.24 0.20 0.07 0.53 0.43 0.21 0.22 0.25 0.10
0.59 0.60 0.53 0.58
0.68 0.68 0.61 0.67
0.55 0.55 0.49 0.53
0.64 0.65 0.58 0.63
0.52 0.40 0.31 0.28 0.13 0.15 0.49 0.35 0.28 0.25 0.21 0.16 0.42 0.30 0.21 0.28 0.37 0.11 0.33 0.21 0.16 0.24 0.53 0.10
0.55 0.55 0.49 0.53
0.67 0.68 0.60 0.66
0.51 0.52 0.46 0.50
0.64 0.65 0.58 0.63
0.44 0.44 0.39 0.43
0.51 0.51 0.46 0.50
0.36 0.36 0.32 0.35
0.40 0.41 0.36 0.40
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Fenestration
15.27
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
21l 6
BLUGRN LE
0.37
0.62
HI-P GRN W/LE CLR 0.57
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
Operable
Operable
Operable
0.38 0.26 0.20 0.29 0.43 0.11 0.28 0.16 0.14 0.25 0.61 0.08 0.38 0.25 0.20 0.29 0.44 0.11 0.27 0.15 0.15 0.25 0.61 0.08 0.31 0.20 0.16 0.28 0.56 0.08 0.26 0.17 0.13 0.28 0.66 0.05
0.30 0.18 0.29 0.37 0.43 0.10 0.22 0.11 0.22 0.34 0.59 0.08 0.30 0.18 0.28 0.37 0.44 0.10 0.21 0.11 0.23 0.34 0.59 0.07 0.24 0.14 0.24 0.37 0.55 0.08 0.21 0.12 0.22 0.37 0.62 0.04
0.16 0.08 0.48 0.57 0.38 0.06 0.12 0.05 0.43 0.55 0.48 0.04 0.16 0.08 0.48 0.57 0.38 0.06 0.12 0.04 0.44 0.55 0.48 0.04 0.13 0.06 0.44 0.57 0.46 0.05 0.12 0.06 0.46 0.57 0.46 0.03
0.40 0.28 0.20 0.27 0.40 0.10 0.30 0.18 0.14 0.24 0.58 0.08 0.39 0.28 0.20 0.27 0.41 0.10 0.28 0.17 0.15 0.24 0.59 0.08 0.33 0.22 0.16 0.27 0.53 0.08 0.27 0.18 0.12 0.27 0.65 0.04
0.42 0.43 0.38 0.41
0.61 0.61 0.54 0.60
0.34 0.34 0.30 0.32
0.54 0.55 0.49 0.54
0.42 0.43 0.38 0.41
0.46 0.47 0.42 0.46
0.32 0.32 0.28 0.30
0.33 0.33 0.30 0.33
0.36 0.36 0.32 0.35
0.55 0.56 0.50 0.55
0.29 0.29 0.26 0.28
0.51 0.51 0.46 0.50
0.41 0.37 0.35 0.39 0.24 0.04 0.37 0.30 0.30 0.35 0.34 0.06 0.26 0.18 0.15 0.34 0.63 0.04
0.40 0.35 0.36 0.39 0.26 0.04 0.36 0.28 0.30 0.35 0.35 0.07 0.25 0.17 0.16 0.34 0.63 0.04
0.38 0.33 0.37 0.40 0.26 0.04 0.34 0.27 0.32 0.35 0.35 0.07 0.24 0.16 0.17 0.35 0.63 0.04
0.34 0.29 0.40 0.43 0.27 0.04 0.31 0.23 0.35 0.38 0.36 0.06 0.22 0.14 0.21 0.37 0.61 0.04
0.27 0.22 0.47 0.50 0.28 0.03 0.24 0.17 0.42 0.44 0.35 0.06 0.18 0.10 0.29 0.44 0.57 0.03
0.14 0.11 0.64 0.66 0.23 0.03 0.13 0.08 0.60 0.60 0.28 0.04 0.10 0.05 0.51 0.60 0.42 0.03
0.36 0.31 0.39 0.42 0.26 0.04 0.32 0.25 0.34 0.37 0.34 0.06 0.23 0.15 0.20 0.37 0.60 0.04
0.38 0.38 0.34 0.36
0.64 0.65 0.58 0.63
0.34 0.34 0.30 0.33
0.62 0.63 0.56 0.62
0.25 0.25 0.22 0.23
0.37 0.38 0.34 0.37
Fixed
0.42 0.30 0.17 0.25 0.42 0.11 0.31 0.19 0.11 0.22 0.61 0.09 0.42 0.30 0.17 0.25 0.43 0.11 0.30 0.18 0.12 0.22 0.61 0.08 0.34 0.23 0.13 0.25 0.56 0.08 0.29 0.19 0.09 0.24 0.67 0.05
Fixed
0.44 0.32 0.16 0.23 0.41 0.11 0.33 0.21 0.10 0.20 0.59 0.09 0.44 0.32 0.16 0.23 0.42 0.11 0.32 0.20 0.11 0.20 0.60 0.08 0.37 0.25 0.12 0.23 0.54 0.09 0.30 0.21 0.07 0.23 0.68 0.05
Fixed
0.46 0.36 0.17 0.23 0.38 0.10 0.36 0.24 0.11 0.19 0.56 0.09 0.46 0.35 0.16 0.23 0.39 0.10 0.34 0.23 0.11 0.20 0.58 0.08 0.39 0.28 0.12 0.23 0.51 0.08 0.31 0.22 0.07 0.23 0.67 0.04
Fixed
Hemis., Diffuse
21k 6
GRY LE
0.52
A1f A f2
80.00
21j 6
GRY LE
0.61
SHGC T Rf Rb
70.00
21i 3
GRN LE
0.68
60.00
Licensed for single user. © 2017 ASHRAE, Inc.
21h 6
GRN LE
50.00
21g 3
Center Glazing Tv
Other Aluminum Frames
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
Low-e Double Glazing, e = 0.05 on surface 2 25a 3
25b 6
25c 6
LE CLR
LE CLR
BRZ W/LE CLR
0.72
0.70
0.42
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.28
2017 ASHRAE Handbook—Fundamentals (SI)
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
0.53
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
SHGC T Rf Rb
A1f A f2
0.27 0.18 0.15 0.37 0.62 0.05 0.21 0.13 0.17 0.37 0.66 0.03 0.24 0.16 0.16 0.37 0.63 0.04 0.23 0.15 0.12 0.37 0.68 0.04
0.68 0.60 0.17 0.17 0.10 0.08 0.06 0.61 0.49 0.14 0.14 0.17 0.12 0.08 0.32 0.20 0.06 0.13 0.64 0.06 0.04
0.65 0.57 0.18 0.18 0.11 0.08 0.06 0.58 0.45 0.15 0.15 0.19 0.13 0.08 0.29 0.17 0.07 0.14 0.67 0.06 0.04
0.62 0.53 0.21 0.21 0.12 0.09 0.06 0.55 0.42 0.18 0.18 0.20 0.13 0.08 0.27 0.15 0.08 0.16 0.68 0.05 0.04
0.54 0.45 0.28 0.28 0.13 0.09 0.06 0.48 0.35 0.24 0.24 0.21 0.13 0.08 0.24 0.12 0.11 0.22 0.68 0.05 0.03
0.39 0.31 0.42 0.42 0.14 0.08 0.05 0.35 0.24 0.37 0.37 0.22 0.12 0.06 0.18 0.07 0.20 0.35 0.66 0.05 0.02
0.18 0.12 0.65 0.65 0.14 0.07 0.03 0.16 0.09 0.59 0.59 0.21 0.08 0.03 0.10 0.02 0.41 0.57 0.53 0.03 0.01
0.57 0.49 0.25 0.25 0.12 0.08 0.06 0.51 0.39 0.22 0.22 0.19 0.12 0.08 0.26 0.15 0.11 0.20 0.65 0.05 0.04
0.60 0.50 0.17 0.19 0.20 0.08 0.06
0.58 0.47 0.19 0.20 0.20 0.08 0.06
0.55 0.44 0.21 0.22 0.20 0.08 0.06
0.48 0.38 0.27 0.29 0.21 0.09 0.06
0.35 0.26 0.41 0.42 0.21 0.08 0.05
0.17 0.10 0.64 0.63 0.17 0.07 0.03
0.51 0.41 0.25 0.26 0.20 0.08 0.06
Fixed
0.12 0.06 0.48 0.60 0.43 0.03 0.09 0.04 0.49 0.60 0.45 0.02 0.11 0.05 0.49 0.60 0.44 0.03 0.11 0.05 0.46 0.60 0.47 0.02
Operable
0.21 0.13 0.25 0.44 0.59 0.04 0.16 0.09 0.26 0.44 0.62 0.03 0.18 0.11 0.26 0.44 0.60 0.04 0.18 0.10 0.22 0.44 0.64 0.03
Other Aluminum Frames Fixed
0.26 0.17 0.16 0.37 0.63 0.05 0.20 0.12 0.18 0.37 0.67 0.03 0.23 0.15 0.17 0.37 0.64 0.04 0.23 0.14 0.13 0.38 0.69 0.04
Fixed
0.28 0.20 0.12 0.35 0.64 0.05 0.22 0.14 0.15 0.35 0.68 0.03 0.25 0.17 0.14 0.35 0.65 0.04 0.25 0.16 0.09 0.35 0.71 0.04
Operable
0.30 0.21 0.10 0.34 0.64 0.05 0.23 0.15 0.13 0.34 0.69 0.03 0.26 0.18 0.12 0.34 0.66 0.04 0.26 0.17 0.07 0.34 0.72 0.04
Fixed
0.31 0.22 0.10 0.35 0.64 0.05 0.24 0.16 0.12 0.34 0.69 0.03 0.27 0.19 0.12 0.34 0.66 0.04 0.27 0.18 0.07 0.35 0.71 0.04
Operable
Hemis., Diffuse
HI-P GRN W/LE CLR
0.45
A1f A f2
80.00
25g 6
BLUE W/LE CLR
0.35
SHGC T Rf Rb
70.00
25f 6
GRY W/LE CLR
0.60
60.00
Licensed for single user. © 2017 ASHRAE, Inc.
25e 6
GRN W/LE CLR
50.00
25d 6
Center Glazing Tv
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
0.29 0.29 0.26 0.28
0.53 0.54 0.48 0.53
0.23 0.23 0.20 0.21
0.31 0.32 0.28 0.31
0.26 0.25 0.22 0.24
0.40 0.41 0.36 0.40
0.26 0.25 0.22 0.24
0.47 0.48 0.42 0.47
0.62 0.62 0.55 0.60
0.66 0.67 0.59 0.65
0.56 0.56 0.50 0.54
0.62 0.63 0.56 0.62
0.30 0.30 0.26 0.29
0.47 0.48 0.42 0.47
0.55 0.55 0.49 0.53
0.61 0.61 0.54 0.60
Triple Glazing 29a 3
29b 6
29c 6
CLR CLR CLR
CLR CLR CLR
HI-P GRN CLR CLR
0.74
0.70
0.53
SHGC T Rf Rb
A1f A f2 Af3
SHGC T Rf Rb
A1f A f2 Af3
SHGC T Rf Rb
A1f A f2 Af3 Triple Glazing, e = 0.2 on surface 2 32a 3
LE CLR CLR
0.68
SHGC T Rf Rb
A1f A f2 Af3
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Fenestration
15.29
Table 10 Visible Transmittance Tv , Solar Heat Gain Coefficient (SHGC), Solar Transmittance T , Front Reflectance R f , Back Reflectance Rb , and Layer Absorptance A nf for Glazing and Window Systems (Continued)
Operable
Operable
Operable
0.41 0.27 0.21 0.24 0.34 0.11 0.07
0.29 0.17 0.31 0.36 0.37 0.10 0.05
0.14 0.06 0.53 0.57 0.31 0.08 0.03
0.44 0.30 0.20 0.22 0.31 0.11 0.07
0.49 0.49 0.43 0.47
0.57 0.58 0.51 0.56
0.62 0.50 0.19 0.18 0.11 0.09 0.11 0.56 0.39 0.16 0.14 0.17 0.13 0.15
0.60 0.47 0.20 0.19 0.12 0.10 0.11 0.53 0.36 0.16 0.15 0.19 0.14 0.16
0.57 0.44 0.22 0.21 0.13 0.10 0.11 0.50 0.33 0.19 0.17 0.20 0.14 0.15
0.49 0.38 0.29 0.27 0.14 0.10 0.10 0.44 0.27 0.24 0.21 0.21 0.14 0.14
0.36 0.26 0.42 0.41 0.15 0.10 0.08 0.32 0.17 0.36 0.31 0.22 0.13 0.12
0.16 0.10 0.63 0.64 0.15 0.08 0.04 0.15 0.06 0.57 0.53 0.22 0.10 0.05
0.52 0.41 0.26 0.25 0.13 0.10 0.10 0.47 0.30 0.22 0.20 0.19 0.13 0.14
0.57 0.57 0.50 0.55
0.61 0.61 0.54 0.60
0.51 0.52 0.46 0.50
0.57 0.58 0.1 0.56
0.41 0.29 0.30 0.30 0.25 0.07 0.08 0.36 0.24 0.34 0.23 0.24 0.10 0.09
0.39 0.26 0.30 0.30 0.27 0.08 0.09 0.34 0.21 0.34 0.23 0.25 0.11 0.09
0.37 0.24 0.31 0.31 0.28 0.08 0.09 0.32 0.19 0.35 0.25 0.26 0.11 0.09
0.32 0.20 0.34 0.34 0.30 0.08 0.09 0.28 0.16 0.38 0.28 0.28 0.11 0.08
0.24 0.13 0.41 0.41 0.32 0.07 0.07 0.21 0.10 0.44 0.36 0.30 0.10 0.07
0.12 0.05 0.59 0.59 0.27 0.06 0.04 0.10 0.03 0.61 0.56 0.25 0.07 0.03
0.34 0.23 0.33 0.33 0.28 0.07 0.08 0.30 0.18 0.37 0.27 0.26 0.10 0.08
0.38 0.38 0.34 0.36
0.55 0.56 0.50 0.55
0.34 0.34 0.30 0.32
0.53 0.53 0.47 0.52
0.27 0.18 0.41 0.46 0.27 0.12 0.02 0.26 0.15 0.33 0.39 0.34 0.15 0.03
0.25 0.17 0.41 0.45 0.28 0.12 0.02 0.25 0.14 0.33 0.38 0.36 0.15 0.03
0.24 0.16 0.42 0.46 0.28 0.12 0.02 0.23 0.12 0.34 0.38 0.36 0.15 0.03
0.21 0.13 0.44 0.48 0.29 0.12 0.02 0.21 0.10 0.37 0.40 0.37 0.14 0.03
0.16 0.08 0.50 0.53 0.30 0.11 0.01 0.16 0.07 0.43 0.46 0.36 0.12 0.02
0.08 0.03 0.65 0.68 0.24 0.07 0.01 0.08 0.02 0.60 0.61 0.28 0.08 0.01
0.23 0.14 0.44 0.47 0.28 0.12 0.02 0.22 0.12 0.36 0.40 0.35 0.14 0.03
0.26 0.25 0.22 0.25
0.52 0.52 0.46 0.51
0.25 0.25 0.21 0.24
0.49 0.0 0.44 0.48
Fixed
0.47 0.33 0.17 0.19 0.31 0.11 0.08
Fixed
0.50 0.36 0.15 0.16 0.31 0.11 0.08
Fixed
0.53 0.39 0.14 0.16 0.28 0.11 0.08
Fixed
Hemis., Diffuse
A1f A f2 Af3
80.00
SHGC T Rf Rb
70.00
0.64
60.00
LE CLR CLR
50.00
32b 6
Center Glazing Tv
Other Aluminum Frames
40.00
ID
Glass Thick., mm
Other Aluminum Frames
Incidence Angles Normal 0.00
Glazing System
Total Window Tv at Normal Incidence
Operable
Total Window SHGC at Normal Incidence
Center-of-Glazing Properties
Triple Glazing, e = 0.2 on surface 5
Licensed for single user. © 2017 ASHRAE, Inc.
32c 3
32d 6
CLR CLR LE
CLR CLR LE
0.68
0.64
SHGC T Rf Rb
A1f A f2 Af3
SHGC T Rf Rb
A1f A f2 Af3 Triple Glazing, e = 0.1 on surface 2 and 5 40a 3
40b 6
LE CLR LE
LE CLR LE
0.62
0.59
SHGC T Rf Rb
A1f A f2 Af3
SHGC T Rf Rb
A1f A f2 Af3 Triple Glazing, e = 0.05 on surface 2 and 4 49
50
3
6
LE LE CLR
LE LE CLR
0.58
0.55
SHGC T Rf Rb
A1f A f2 Af3
SHGC T Rf Rb
A1f A f2 Af3
KEY: CLR = clear, BRZ = bronze, GRN = green, GRY = gray, BLUGRN = blue-green, SS = stainless steel reflective coating, TI = titanium reflective coating Reflective coating descriptors include percent visible transmittance as x%. HI-P GRN = high-performance green tinted glass, LE = low-emissivity coating
Tv = visible transmittance, T = solar transmittance, SHGC = solar heat gain coefficient, and H. = hemispherical SHGC ID #s refer to U-factors in Table 4, except for products 49 and 50.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.30
2017 ASHRAE Handbook—Fundamentals (SI) Table 11 Solar Heat Gain Coefficients for Domed Horizontal Skylights
Dome
Curb Light Solar Heat Visible Height, Width-toDiffuser Gain Trans(Translucent) mm Height Ratio Coefficient mittance
Clear = 0.86
Yes = 0.58
Clear = 0.86
None
Translucent = 0.52 Translucent = 0.27
None None
0 225 450 0 225 450 0 450 0 225 450
5 2.5 5 2.5 2.5 5 2.5
0.53 0.50 0.44 0.86 0.77 0.70 0.50 0.40 0.30 0.26 0.24
0.56 0.58 0.59 0.91 0.91 0.91 0.46 0.32 0.25 0.21 0.18
Licensed for single user. © 2017 ASHRAE, Inc.
Sources: Laouadi et al. (2003), Schutrum and Ozisik (1961).
Tubular Daylighting Devices (TDDs). Tubular daylighting devices collect and channel daylight (diffuse sky and sunbeam light) from the roof of a building into interior spaces (Figure 16). TDDs are also known as mirror light pipes or tubular skylights. They are an alternative to conventional skylights and have the advantages of energy savings (small area relative to the amount of useful light they admit), lower solar heat gains, relative ease of installation, and tube lengths (depending on the tube length-to-diameter ratio) of up to 13 m. TDD technologies undergo continuous development to meet the increasing needs for high standards of energy efficiency in buildings and glare-free lighting. The energy-saving potential of TDDs is well recognized (Allen 1997; Carter 2008; Laouadi 2004), and TDDs are encouraged or mandated, particularly in commercial buildings [CEC 2010; see also U.S. Green Building Council’s (USGBC) Leadership in Energy and Environmental Design (LEED) program]. A variant known as a hybrid TDD (HTDD) is available from some manufacturers. In an HTDD, more than one material or geometry is present along the length of the tube (e.g., a round tube that transitions to a square ceiling diffuser). The natural daylight that TDDs deliver is beneficial to the visual comfort, health, and well-being of building occupants (Boyce et al. 2003; Carter and Al-Marwaee 2009). The daylighting performance (light output) of a TDD depends on many location and material parameters, but a typical device can illuminate an area of up to 28 m2. The area of coverage depends on the height of the ceiling: the higher the ceiling, the more widely the light will be uniformly distributed. TDD Components. TDDs typically consist of (1) a collector on the roof to gather sunbeams and diffuse sky light, (2) a hollow pipe guide in the plenum/attic space to channel light downwards, and (3) a light diffuser at ceiling level to spread light indoors (see Figure 16). The collector is usually a single- or multilayer transparent glass or plastic dome projecting above the plane of the roof. It may also include passive or active geometries (e.g., static or tracking reflectors) or some optical devices (e.g., prisms, diffusing elements) to enhance the device’s lighting output, especially at low sun altitude angles (Laouadi 2013). Straight guides (tubes) are common, but nonlinear light guides with bent sections (Figure 16) are sometimes needed to suit the geometry of a building. The guide can be rigid or flexible, and is typically made from an aluminum sheet with highly reflective interior coatings, which are sometimes spectrally selective. This
Fig. 16
Generalized Tubular Daylighting Device (NRCC 2011)
measure reduces solar heat transmission while preserving light transmission. Coating materials with a reflectivity exceeding 99% are commercially available. Vertical pipes without bends are more efficient than bent guides for low incidence angles (<60°; e.g., summer noon), and less efficient for higher incidence angles (e.g., winter noon) (Laouadi et al. 2013a). Diffusers in the ceiling plane of TDDs are usually hemispherical or flat with single- or multilayer glass or poly(methyl methacrylate (PMMA, also known as acrylic) or polycarbonate. Diffusers typically use one of three types of glazing: prismatic, diffusing, or lensed. Prismatic glazing consists of arrays or layers of micro replicated shapes (conical prisms, ridges, wedges, etc.), which reflect, refract, and redirect the incoming light based on the angle of incidence. Typical prismatic diffusers are made of acrylic or polycarbonate sheets with conical prisms. Diffusing glazing consists of very fine particles, pigments, or films applied to clear glass or plastic sheets to make the glazing look opal (white). The scattering properties depend on the refractive index and the size of the microstructures used. Lensed glazing is comprised of arrays of diverging lenses (e.g., Fresnel lenses) that are very thin so that little light is lost by absorption. Performance Metrics. Various theoretical and empirical models have been developed to predict the daylighting performance of TDDs [e.g., CIE (2006), Zhang and Muneer (2000)]. ASHRAE research project RP-1415 developed and validated models for the prediction of the optical and thermal performance of complex TDDs (Laouadi et al. 2013b, 2013c, 2013d). Several metrics were proposed, including • Visible transmittance (VT): calculated or measured at a standard incidence angle, or averaged over a predetermined number of incidence angles. Angle of incidence is the angle between the entering ray and the axis of the upper tube section. • Normalized intensity distribution: an index that treats TDDs as lighting luminaire systems. TDD intensity distribution is important for lighting energy calculation and spatial arrangement of TDDs.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Fenestration
15.31 Table 12 Performance Characteristics of Typical TDDs Glazing of Collector/ Diffuser Visible transmittance* SHGC* (250 mm insulation above ceiling) SHGC* (250 mm insulation under roof) U-factor, W/(m2 K) (250 mm insulation above ceiling) U-factor (250 mm insulation under roof)
Single/ Double/ Double/ Single/ Single/ Single Double Single Double Triple Layer Layer Layer Layer Layer 0.77 0.32
0.63 0.24
0.72 0.27
0.66 0.28
0.58 0.24
0.34
0.33
0.32
0.35
0.34
3.54
2.16
3.53
2.17
1.55
7.62
4.70
4.70
7.60
7.58
*VT and SHGC calculated at solar incidence angle of 50° at due south; this angle is close to annual averaged incidence angle (48.84°) of NFRC Standard 203-2014.
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Fig. 17 Transmittance of Straight Tube (Light Pipe) as Function of Reflectivity and Aspect Ratio (Length/Diameter) • Daylighting area coverage (DAC): portion of the work plane surface area under a daylight illuminance equal or greater than the recommended task illuminance under CIE standard clear sky conditions. DAC is an important parameter for TDD product rating and selection, and spatial arrangement that guarantees capital cost reduction (with minimum number of TDDs). • Spacing ratio (SR): maximum distance between two pairs of TDDs, normalized by the ceiling height of buildings. SR is calculated based on the illuminance uniformity criterion on the work plane surface. • Solar heat gain coefficient (SHGC): normalized solar heat gains penetrating the indoor space through a TDD system when sunbeam light is incident on the TDD pipe aperture at a standard angle. • Thermal transmittance (U-factor): heat loss from a TDD to the outdoors per unit surface area and temperature difference between the indoor and outdoor environments under standard thermal conditions. Simplified Single-Angle Solar and Visible Analysis. The following simplified equations are presented only as a guide to solar-optical behavior of a TDD. Real-world TDD performance is a function of many location-specific and season-dependent factors (especially solar angle of incidence) which lead to the final, effective SHGC and VT being a somewhat complex weighted average. For a given input solar ray and angle of incidence, the visible (or solar) transmittance of a TDD is the product of the individual visible (or solar) transmittances of the collecting dome c, the tube t, and the ceiling diffuser d: TTDD = tc tt td Of the three terms on the right-hand side of the equation, tube transmittance tt is by far the most sensitive to the incidence angle of the ray (Figure 17) and the presence of any optical element in the collector, which may change the angle of the incident ray. The equation also assumes that multiple interreflections between TDD components make a negligible contribution to the overall system transmittance and throughput of light. This is approximately true for large aspect ratios. The solar heat gain coefficient of the TDD is the sum of the (solar) transmittance plus an indirect, thermal term that is analogous to the inward-flowing fraction of solar heat gain from a general glazing system. The fraction of original, outdoor solar power incident on the diffuser is Id /Is = tc tt
where Is and Id are the solar fluxes (W/m2) incident on the collector dome and diffuser respectively. The fraction of incident solar radiation energy absorbed in the diffuser is d of which approximately half is transferred downward by convection and radiation into the building. The other half is transferred upward to the column of air in the tube. The fraction of this power that is transferred inward is equal to 0.5tcttd. Thus the solar heat gain coefficient of the TDD system, and where all optical parameters are determined for the solar spectrum, is approximately SHGCTDD = TTDD + 0.5tc tt d = tc tt (td + 0.5d) The National Fenestration Rating Council (NFRC) sets the model sizes and environmental conditions for the product rating of TDDs. At the time of writing, NFRC ratings are based on physical testing only, after a former NFRC simulation procedure for U-factor of TDDs was retired in 2012. TDD ratings for U-factor follow NFRC (2014a) Technical Document 100. Ratings for SHGC follow NFRC (2014e) Technical Document 201-2014, whereas ratings for VT follow NFRC (2014f) Technical Document 203-2014, which in turn references ASTM Standard E1175-87 (RA 2015). Table 12 shows the computed optical and thermal characteristics of a typical TDD made up of a 3 mm PMMA collector and a 3 mm polycarbonate diffuser with an NFRC standard-size pipe of 350 mm diameter and aspect ratio of 2.14. The tube reflectitivity is 99% and 76% for the visible and solar spectrum, respectively. The NFRC Technical Document 100 boundary conditions (NFRC 2014a) are used for the U-factor calculation. The table shows that, although the number of layers in the collector or diffuser has only a small effect on visible transmittance, as expected it has a significant effect on thermal performance. The table does not show any best arrangement, but provides information that can help a designer select an arrangement that best suits a particular application in residential or commercial buildings. Practical Guidelines. When selecting TDDs, specifiers should consider the following: • Visible transmittance: products with the highest visible transmittance should be selected for daylighting. • Solar heat gain coefficient: ideally, for heating-dominated climates, TDDs would have moderate to high SHGC. However, most available TDDs have medium to low SHGC (0.40) by design, because ENERGY STAR (www.energystar.gov) requires SHGC 0.40 except in the northern U.S. climate zone. Some indicative, real-world solar heat gain coefficients are given in Table 12. All exceed the minimum performance defined by ENERGY STAR (i.e., rated SHGC is lower than the ENERGY STAR upper limit).
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2017 ASHRAE Handbook—Fundamentals (SI) Table 13
Solar Heat Gain Coefficients for Standard Hollow Glass Block Wall Panels
Type of Glass Blocka
Solar Heat Gain Coefficient In Sun In Shadeb
Description of Glass Block
Type I
Glass colorless or aqua A, D: Smooth B, C: Smooth or wide ribs, or flutes horizontal or vertical, or shallow configuration E: None Type IA Same as type I except ceramic enamel on A Type II Same as type I except glass fiber screen partition E Type III Glass colorless or aqua A, D: Narrow vertical ribs or flutes. B, C: Horizontal light-diffusing prisms, or horizontal light-directing prisms E: Glass fiber screen Type IIIA Same as type III except E: Glass fiber screen with green ceramic spray coating or glass fiber screen and gray glass or glass fiber screen with light-selecting prisms
Type IV Same as type I except reflective oxide coating on A aAll
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values are for 200 by 200 by 100 mm block, set in light-colored mortar. For 300 by 300 by 100 mm block, increase coefficients by 15%, and for 150 by 150 by 100 mm block, reduce coefficients by 15%.
• U-factor: the lower the U-factor, the better. ENERGY STARqualified TDDs must achieve a U-factor of 1.5 W/(m2 ·K) in the northern U.S. climate zone. This is relaxed somewhat as one moves to progressively warmer climates, leading to a threshold of 2.3 W/(m2 ·K) in the southern climate zone. Products that are insulated above the ceiling perform considerably better than those with insulation under the roof. • Tubular daylighting devices: TDDs should have multipane ceiling diffusers when insulation is placed at ceiling level (e.g., residential buildings) or multipane collectors when insulation is placed at roof level (e.g., commercial buildings) to reduce thermal losses. In residential applications, a multipane ceiling diffuser is the single most effective measure to reduce the overall U-factor of a TDD. This is because tubes typically penetrate an unconditioned attic zone. In winter, heat loss from tube cavity to attic zone can be severe, and disproportionately large in comparison with the TDD’s cross-sectional area, if the ceiling diffuser is only single-glazed. • Spacing ratio (SR): SR relates to the normalized luminous intensity distribution of TDDs. To minimize capital cost (number of TDDs), select TDDs with higher spacing ratios. For ideally diffusing TDDs, SR varies from 1 to 1.5 for illuminance uniformity (minimum to average ratio) of 0.95 to 0.8, respectively. Installation. Installation methods vary depending on roof type (asphalt shingles, roof tiles, membranes, metallic roofing, etc.). TDD assemblies must be installed to resist air and water infiltration, and damage from wind, hail, and snow load. Manufacturers’ instructions provide specific details. Attention to detail is required when the roof is replaced.
Glass Block Walls Glass block can be used for light transmission through outdoor walls when optical clarity for view is unnecessary. Table 13 describes a variety of glass block patterns and gives solar heat gain coefficients to be applied to solar irradiances so that approximate instantaneous solar heat gains can be calculated (Smith and Pennington 1964). Table 4, footnote 6, provides a representative U-factor for glass block. Note that the U-factor for glass block is poor when compared to double glazing with a low-emissivity coating. Convection and low-temperature radiative heat gain for all hollow glass block panels fall within a narrow range. Differences in SHGCs are largely the result of differences in transmittance of glass blocks for solar radiation. Solar heat gain coefficients for any
0.57
0.35
0.23 0.38
0.17 0.30
0.29
0.23
0.22
0.16
0.14
0.10
bFor
NE, E, and SE panels in shade, add 50% to values listed for panels in shade.
particular glass block pattern vary depending on orientation and time of day. The SHGC for western exposures in the morning (shaded) is depressed because of heat storage in the block, whereas the SHGC for eastern exposures in the afternoon (shaded) is elevated as stored heat is dissipated. Time lag effects from heat storage are estimated by using solar gains and air-to-air temperature differences for one hour earlier than the time for which the load calculation is made. Calorimeter tests of Type 1A glass block showed little difference in solar heat gains between glass block with either black or white ceramic enamel on the exterior of the block. White and black ceramic enamel surfaces represent the two extremes for reflecting or absorbing solar energy; therefore, glass block with enamel surfaces of other colors should have solar heat gain coefficients between these values. Because glass blocks are good examples of strongly angularly selective fenestrations, appropriate caution must be taken.
Plastic Materials for Glazing Generally, factors outlined for glass apply also to glazing materials such as acrylic, polycarbonate, polystyrene, or other plastic panels. If solar transmittance, absorptance, and reflectance are known, the SHGC can be calculated in the same way as for glass. These properties can be obtained from the manufacturer or be determined by simple laboratory tests. The National Fenestration Rating Council developed standards for testing the optical properties of glazing (NFRC 2014d, 2014g). In selecting plastic panels for glazing, concerns include possible deterioration from the sun, expansion and contraction because of temperature extremes, and possible damage from abrasion.
4.3
CALCULATION OF SOLAR HEAT GAIN
To calculate solar energy fluxes, first calculate the incident angle from the local standard time and the longitude. The direct normal solar irradiance EDN , diffuse sky irradiance Ed , ground-reflected radiation Er , and total incident irradiance Et can then be determined. Note that the latter two are assumed to be ideally diffuse radiation. Calculation methods for these parameters are described in Chapter 14. Solar energy flow through a fenestration may be divided into two parts: opaque (qop) and glazing (qs ) portions. The glazing solar energy flux qs can be split into that from incident beam radiation
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Fenestration
15.33 qd = qdt + qda
(27)
where qdt = glazing solar energy flux caused by transmitted incident diffuse radiation qda = glazing solar energy flux caused by inward heat flow of absorbed diffuse radiation
Glazing solar energy flux caused by diffuse incident radiation is calculated from qd = (Ed + Er)SHGCD
(28)
where the hemispherically averaged solar heat gain coefficient is calculated from Equation (15). Solar radiant and heat fluxes can be separately calculated from Fig. 18
qdt = (Ed + Er)T D
Instantaneous Heat Balance for Sunlit Glazing Material
(qb) and incident diffuse radiation (qd ), which includes both diffuse sky radiation and radiation scattered (reflected) from the ground:
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qs = qb + qd
qb = qbt + qba
(23)
where qbt = glazing solar energy flux caused by transmitted incident beam radiation qba = glazing solar energy flux caused by inward heat flow of absorbed beam radiation
The glazing solar energy flux caused by incident beam radiation is calculated from qb = EDN cosSHGC()
(24)
where the beam solar heat gain coefficient is given by Equation (12) or (13). If, instead, the solar radiant and heat fluxes are needed separately, calculate the glazing transmitted solar flux (solar radiation traveling in the incident direction) from qbt = EDN cosT()
(25)
and the inward-flowing absorbed solar flux (heat) from q ba = E DN cos N k A k L
f
(26)
k=1
Values of T () and A f() in these equations can be found in Table 10, and determination of Nk is discussed in the section on Calculation of Solar Heat Gain Coefficient. For diffuse radiation,
q da = E d + E r N k A k D L
(22)
The net heat balance that would occur for a sunlit glazing if there were no diffuse radiation is shown in Figure 18. This net heat balance does not include any of the heat flows resulting from indoor/ outdoor temperature differences. The heat balance is pictured as superimposed on the thermal effect. This superposition picture should not be carried too far, however, because the heat flows indicated in Figure 18 as resulting from convection and radiation depend in part on processes that are nonlinear with respect to temperature, so that in reality the two effects cannot be separated. To calculate them, the actual glazing (and other) temperatures are needed, not simply the incremental temperature rise caused by sunlight. Figure 18 shows that the glazing solar energy flow from beam radiation consists of two parts:
(29)
which is diffusely distributed solar radiation (note that effects of finite glazing size and thickness are neglected), and f
(30)
k=1
Opaque Fenestration Elements The opaque portion solar energy flux is calculated from qop = (EDN cos + Ed + Er)SHGCop
(31)
where SHGCop is obtained from Equation (18). Example 6. Calculate the solar energy flux through the glazing system ID 25a given = 60o, EDN = 600 W/m2, and Ed = 150 W/m2. Solution: The solar heat gain coefficients are SHGC(60°) = 0.34, and SHGCD = 0.36 (see Table 10, glazing ID 25a, 5th and 8th columns). qs = qb + qd = EDN cos(60)SHGC(60) + (Ed + Er)SHGCD = (600.0)cos(60)(0.34) + (150)(0.36) = 156 W/m2
5.
SHADING AND FENESTRATION ATTACHMENTS 5.1
SHADING
The most effective way to reduce the solar load on fenestration is to intercept direct radiation from the sun before it reaches the glazing system. Fenestration products fully shaded from the outdoors reduce solar heat gain by as much as 80%. Fenestration can be shaded by roof overhangs, vertical and horizontal architectural projections, awnings, heavily proportioned outdoor louvers, or a variety of vegetative shades, including trees, hedges, and trellis vines. In all outdoor shading structures, it is necessary to consider the structures’ geometry relative to changing sun position to determine the times and quantities of direct sunlight penetration. A detailed discussion of the effectiveness of outdoor shading is given in Ewing and Yellott (1976). The general effect of shading is to attenuate solar radiation. Some of the beam radiation may reach the fenestration unaffected by the shade, and this is accounted for by the unshaded fraction Fu. Assuming that the shade does not transmit or diffuse solar radiation, the solar heat gain of the fenestration can be approximated by modifying Equation (22): qs = Fu qb + qd,shaded
(32)
Here, the term qd,shaded indicates that a new SHGC must be determined to account for the fact that the shading device restricts the amount of sky-diffuse radiation on the fenestration system. More
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2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 19 Profile Angle for South-Facing Horizontal Projections
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complex models are required for situations where the shade is partially transmitting and diffusing in nature.
Overhangs and Glazing Unit Recess: Horizontal and Vertical Projections In the northern hemisphere, horizontal projections can considerably reduce solar heat gain on south, southeast, and southwest exposures during late spring, summer, and early fall. On east and west exposures during the entire year, and on south exposures in winter, the solar altitude is generally so low that, to be effective, horizontal projections must be excessively long. On the other hand, recessing the fenestration deeper back into the wall achieves the same effect as a horizontal projection. The ability of horizontal projections to intercept the direct component of solar radiation depends on their geometry and the profile or shadow-line angle (Figure 19), defined as the angular difference between a horizontal plane and a plane tilted about a horizontal axis in the plane of the fenestration until it includes the sun. The vertical profile angle can be calculated by tan = tan /cos
(33)
where
Fig. 20 Vertical and Horizontal Projections and Related Profile Angles for Vertical Surface Containing Fenestration Example 7. A window facing 30° south of west (wall azimuth = +60°) in a building at 33.65°N latitude, and 84.42°W longitude is 1841.5 mm wide and 6286.5 mm high. The depth of the horizontal projection is 2438 mm. At 3:00 PM on July 21, it is calculated that the hour angle H = 15 (13.27 – 12) = 19.03°; and the declination = 20.60°. The solar altitude is calculated to be: sin = cos(33.65) cos(20.60) cos(19.03) + sin(33.65) sin(20.60) = 68.7° The solar azimuth is cos = [sin(68.68) sin(33.65) – sin(20.60)]/[cos(68.68) cos(33.65)] = 57.1° Thus, the wall solar azimuth is = 57.1 – 60 = –2.9°. (a) Find the sunlit and shaded area of the window. (b) Find the depth of the projections necessary to fully shade the window. Solution: (a) Using Equation (34), the width of the vertical projection shadow is SW = 0 |tan(–2.9)| = 0 mm
= solar altitude angle = solar azimuth
Using Equation (33), the profile angle for the horizontal projection is
The shadow width SW and shadow height SH (Figure 20) produced by the vertical and horizontal projections (PV and PH), respectively, can be calculated using the surface solar azimuth and the vertical profile angle determined by Equation (33). SW = PV |tan |
(34)
SH = PH tan
(35)
When the surface solar azimuth is greater than 90° and less than 270°, the fenestration product is completely in the shade; thus, SW = W + Rw and sunlit area ASL = 0. The sunlit (ASL) and shaded (ASH) areas of the fenestration product are variable during the day and can be calculated for each moment using the following relations: ASL = [W – (SW – RW)][H – (SH – RH)]
(36)
ASH = A – ASL
(37)
where A is total fenestration product area. For software-based or multiple calculations, McCluney (1990) describes an algorithm that can be used to calculate the unshaded fraction of a window equipped with overhangs, awnings, or side fins.
tan = tan(68.7)/cos(2.9) = 68.7° Using Equation (35), the height of the horizontal projection shadow is SH = 96 tan(68.7) = 6255 mm Using Equations (36) and (37), the sunlit and shaded areas of the window are now ASL = [1841.5 – (0 – 0)][6286.5 – (6255 – 0)]/106 = 0.058 m2 ASH = (1841.5 6286.5)/10 6 – 0.058 = 11.519 m2 (b) The shadow length necessary to fully shade the given window SH( fs) and SW( fs ) from the horizontal and vertical projection are given by (see Figure 20) SH( fs) = 6286.5 + 0 = 6286.5 mm SW( fs) = 1841.5 + 0 = 1841.5 mm Thus, using Equations (34) and (35), PH( fs) = 6286.5 cot(68.7) = 2453.5 mm PW( fs) = 1841.5|cot(–2.9)| = 36 351.7 mm
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Fenestration For this example, because both horizontal and vertical projections do not need to fully shade the window, a horizontal projection of 2454 mm is satisfactory. Also, to accurately analyze the influence of external projections, an hour-by-hour calculation must be performed over the periods of the year for which shading is desired.
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5.2
FENESTRATION ATTACHMENTS
Fenestration attachments generally consist of items that can be used as part of a system to provide solar and daylighting control, as well as privacy, aesthetics, and comfort for building occupants. Attachments also include other devices that, though not intended for solar control, affect the solar and visual performance of the fenestration system. Attachments to the indoor side of fenestration can include horizontal louvers (venetian blinds), vertical louvers, roller shades, insect screens, and drapery. Between glazings of multiglazed fenestration, horizontal louvers and roller shades may be incorporated. On the outdoor side, insect screens can be added, as well as horizontal louvers in the plane of the fenestration. Fenestrations with shading devices have a degree of thermal and optical complexity far greater than that of unshaded fenestrations, and are referred to as complex fenestration. In unshaded fenestration, individual glazing layers can only communicate thermally with adjacent layers. This is not the case for complex fenestration. A fenestration layer such as a screen or louvered blind is not sealed, and allows convective heat transfer between nonadjacent layers. Similarly, shading layers are inherently diathermanous (i.e., they transmit both long- and short-wave radiation). Radiative heat transfer, therefore, can also occur between nonadjacent layers. For example, for a window with indoor venetian blinds, heat transfer occurs between the indoor glass and the blind, the indoor glass and the room, and between the blind and the room. Therefore, methods described previously for determining the Ufactor and inward-flowing fractions of fenestration systems cannot be applied to complex fenestration (Collins and Wright 2006; Wright 2008). Also, complex fenestration can have a nonspecular optical element. This is an element for which light (or short-wave infrared radiation) incident on the element from a single spatial direction does not emerge traveling in a single transmitted direction and/or a single reflected direction. Examples of nonspecular elements are shades, drapes, blinds, honeycombs, figured glass, ground glass, and other diffusers, lenses, prisms, and holographic glazings. Two methods have been developed that allow the analysis of complex fenestration. The first method was proposed by Klems (1994a, 1994b, 2001), and relies on measurement of the bidirectional transmittance and reflectance of each glazing layer, and on calorimetrically determined values of inward flowing fraction, as input to a matrix calculation. It is a physically based and highly accurate approach that is also computationally and experimentally intensive. For details of this approach, see Chapter 31 of the 2005 ASHRAE Handbook—Fundamentals. The second method, developed through ASHRAE-sponsored research (Wright et al. 2008), is an empirically based approach that uses readily available information about the system geometry and material properties, and is designed to fit into established fenestration analysis methodology. The methodology has been shown to accurately predict complex fenestration performance from easily obtained data regarding shade geometry and material. In contrast to these two methods, a simplified approach is presented in the following section for calculating the approximate SHGC for a selection of the more common shading elements and glazing systems.
15.35 Simplified Methodology Considering only the approximate total heat flux through the fenestration, measurements made on a fenestration under one set of conditions can often be extrapolated to other fenestrations and conditions to give an adequate answer. In this case, heat flux through the center-glass region is represented by q = EDN cos()SHGC()IAC(, ) + (Ed + Er)SHGCDIACD
(38)
where the solar heat gain coefficients are for the center-glass region of an unshaded glazing, and may be calculated using methods described previously, or obtained from Table 10. The indoor solar attenuation coefficient (IAC) represents the fraction of heat flow that enters the room, some energy having been excluded by the shading. Depending on the type of shade, it may vary angularly and with shade type and geometry. The IAC is defined as SHGC cg , shaded IAC = ------------------------------------------------------SHGC cg SHGC D cg , shaded IAC D = ------------------------------------------------- SHGC D cg
(39)
where is either the horizontal or vertical profile angle. IAC values presented in the following sections have been determined using the ASHWAT models (Wright et al. 2008), which have been validated, with calorimetric results showing prediction of fenestration performance to within 5% (Figure 21). Because shading layers generally have a small effect on the Ufactor of complex fenestration systems (Wright et al. 2008), in this simplified analysis, the effects of shading devices on U-factor are ignored. System U-factor is assumed to be similar to that of the same glazing (minus the shade) and can be determined from Table 4. Note that this simplified approach applies only to the SHGC of the center-glass region of the fenestration product. Results from this analysis must be combined with the methods provided in the Solar Heat Gain Coefficient and Solar Heat Gain sections.
Slat-Type Sunshades Slat-type sunshades consist of horizontal or vertically oriented louvers in located in the plane of the fenestration. They can be installed on the outdoor and indoor side of the fenestration, or between glazings in a multilayered glazing system. The transmitted solar radiation may consist of straight-through, transmitted diffuse, and reflected-through components. The geometry considered is shown in Figure 22, with slat width w, slat crown c, slat spacing s, and slat angle . The ratios of w/s and w/c are assumed constant at 1.2 and 16 respectively, which are representative of many commercially available products. The profile angle can represent either the vertical profile V or the horizontal profile H. The vertical profile angle is used for horizontal louvered shades, and is calculated using Equation (33). The horizontal profile angle is used for vertical louvered shades and is equal to the wall surface solar azimuth . Tables 14A to 14G presents IAC values at profile angles of 0 and 60o for various glazing and shade combinations. IAC varies with profile angle where the profile angle can be the vertical profile (for horizontal louvers) or horizontal profile (for vertical louvers). The variation of IAC with profile angle can be determined from IAC(,) = IAC0 + IACx min(1.2,60)/60
(40)
Barnaby et al. (2009), Collins et al. (2008), Huang et al. (2006), Kotey et al. (2009a), and Wright et al. (2008) contain more
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Fig. 21 Comparison of IAC and Solar Transmission Values from ASHWAT Model Versus Measurements (normal incidence; various shading layers attached to conventional double-glazed window) (Wright et al. 2009b) (a) Indoor shade IAC (60,40) = IAC 0 + IAC 60 – IAC 0 min (1.2, 60) 60 40 = 0.99 + 0.87 – 0.99 -----60 = 0.91 SHGC(,)cg, shaded = IAC(60,40) SHGC()cg = 0.91 0.34 = 0.31 SHGCD, cg, shaded = IACD SHGCD, cg = 0.93 0.36 = 0.33 (b) Louvers between glazings IAC (60,40) = IAC 0 + IAC 60 – IAC 0 min (1.2, 60) 60 40 = 0.97 + 0.68 – 0.97 -----60 = 0.78 SHGC(,)cg, shaded = = = SHGCD, cg, shaded = = =
Fig. 22 Geometry of Slat-Type Sunshades comprehensive discussions of models used to determine IACs of louvered sunshades. Example 8. Calculate the SHGC of glazing system ID 25a if a horizontally louvered shade is added (a) on the indoor side, (b) between the glazings, and (c) on the outdoor side. The shade material has a reflectance of 0.60 and the shades are installed in the open position (0°). = 60° and V = 40°. Also consider (d) vertical louvers located on the indoor side of the glazing with H = 40°. Solution: Use Equations (39) and (40) and values from Table 14E. From Example 6, SHGC(60°) = 0.34, and SHGCD = 0.36.
IAC(60,40) SHGC()cg 0.78 0.34 0.27 IACD SHGCD, cg 0.82 0.36 0.30
(c) Outdoor louvers IAC (60,40) = IAC 0 + IAC 60 – IAC 0 min (1.2, 60) 60 40 = 0.94 + 0.15 – 0.94 -----60 = 0.41 SHGC(,)cg, shaded = = = SHGCD, cg, shaded = = =
IAC(60,40) SHGC()cg 0.41 0.34 0.14 IACD SHGCD, cg 0.51 0.36 0.18
(d) Vertical louvers For the given conditions, results are the same as for part (a).
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Drapery Drapery fabrics can be classified in terms of their solar-optical properties as having specific values of fabric transmittance and reflectance. Fabric reflectance is the major factor in determining the ability of a fabric to reduce solar heat gain. Based on their appearance, draperies can also be classified by yarn color as dark, medium, and light and by weave as closed, semiopen, and open. The apparent color of a fabric is determined by the reflectance of the yarn itself. Drapery fabrics are classified into nine types, rated by openness and yarn reflectances (Figures 23 and 24). The solar-optical properties of drapery fabrics can be determined accurately by laboratory tests (Collins et al. 2011; Kotey et al. 2009b; Yellott 1963), and manufacturers can usually supply solar transmittance and reflectance values for their products. In addition to these properties, the openness factor (ratio of the open area between the fibers to the total area of the fabric) is a useful property that can be measured exactly (Keyes 1967; Pennington and Moore 1967). Visual estimations of openness and yarn reflectance, interpreted through Figures 23 and 24, are valuable in judging the effectiveness of drapes for (1) protection from excessive radiant energy from either sunlight or sun-heated glazing, (2) brightness control, (3) providing either outward view or privacy, and (4) sound control. To understand drapery layer solar-optical properties, the fullness of the drapery is needed. As simplified in Figure 25, the pleating of the drape is assumed to be square, with pleat depth w and width s. For 100% fullness, the width of fabric used is twice the width of the fenestration. If the drapery is hung flat, like a fenestration product shade, the fullness is 0%. Table 14G presents IAC and radiant heat transfer fraction FR values for typical glazing and shade combinations. For these types of shades, the IAC value is not strongly influenced by the incident
15.37 angle of irradiation; therefore, a constant value of IAC can be used. Kotey et al. (2009b, 2009c, 2011) and Wright et al. (2009a) contain more comprehensive discussions of models used to determine IACs of draperies. Example 9. Calculate the SHGC of glazing system ID 25a if a drapery is added on the indoor side. The fabric has an openness factor of 0.05 and a yarn reflectance of 0.60, and the drapery has 100% fullness. Solution: In Figure 24, these lines intersect in the area of designator IIIL. Fabric is closed and light in color, with probable fabric reflectance of 0.52 and fabric transmittance of 0.30. From Table 14G, both IAC and IACD = 0.68. SHGC(,)cg, shaded = IAC SHGC()cg
Fig. 24
Fig. 23 Designation of Drapery Fabrics
Drapery Fabric Properties
Fig. 25 Geometry of Drapery Fabrics
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2017 ASHRAE Handbook—Fundamentals (SI)
SHGCD, cg, shaded
= 0.68 0.34 = 0.23 = IACD SHGCD, cg = 0.68 0.36 = 0.25
Roller Shades and Insect Screens In general, both roller shades and insect screens are equivalent to drapery of 0% fullness. Appropriately, much of the methodology applied to drapery fabrics applies for these devices as well. Table 14G presents IAC for typical glazing and shade combinations. For these types of shades, the IAC value is not strongly influenced by the incident angle of irradiation; therefore, a constant value of IAC can be used. For a more comprehensive discussion of models used to determine IACs, see Kotey et al. (2008, 2009d) for roller shades and Kotey et al. (2009e) for insect screens.
Licensed for single user. © 2017 ASHRAE, Inc.
Example 10. Calculate the SHGC of glazing system ID 25a if a roller shade is added on the indoor side. The shade has an openness factor of 0.0 and a yarn reflectance of 0.65. Solution: From Table 14G, both IAC and IACD = 0.60. SHGC(,)cg, shaded = IAC(60,40) SHGC()cg = 0.60 0.34 = 0.20 SHGCD, cg, shaded = IACD SHGCD, cg = 0.60 0.36 = 0.22
6. VISUAL AND THERMAL CONTROLS The ideal fenestration system allows optimum lighting, heating, ventilation, and visibility; minimizes moisture and sound transfer between the outdoors and the indoors; and produces a satisfactory physiological and psychological environment. The controls of an optimum system react to varying climatological and occupant demands. Fixed controls may have operation or cost advantages, or both, but do not react to physical and psychological variations. Variable controls are, therefore, more effective in energy conservation and environmental satisfaction.
Operational Effectiveness of Shading Devices Shading devices vary in their operational effectiveness. Some devices, such as overhangs, light shelves, and tinted glazings, do not require operation, have long life expectancies, and do not degrade significantly over their effective life. Other types of shading devices, especially operable indoor shades, may have reduced effectiveness because of less than optimal operation and degradation of effectiveness over time. It is important to evaluate operational effectiveness when considering the actual heat rejection potential of shading devices. The performance of shading devices for reducing peak cooling loads and annual energy use should account for operational effectiveness or reliability in actual operation. Passive devices, such as architectural elements and glazing tinting, are considered 100% effective in operation. Glazing coatings and adherent films may degrade over time. Shade screens are removable and may be assumed to operate seasonally, but in any given population of users, some will remain in place all year long and some will not be installed or removed at optimum times. Automated shading devices controlled for optimum thermal operation are considered more effective than manual devices, but controls require ongoing maintenance, and some occupants may object to the lack of personal control with totally automated devices. Automated shading devices may also operate for nonthermal purposes such as glare and daylighting optimization, and this may reduce thermal effectiveness. Manually operated devices are subject to wide variation in use effectiveness,
and this diversity in effective use should be considered when evaluating performance.
Indoor Shading Devices Although thermal comfort of occupants may be paramount to the HVAC designer, other factors that should be considered, some of which may be more important to the user, include the following: Radiant Energy Protection. Unshaded fenestration products become sources of radiant heat by transmitting short-wave solar radiation and by emitting long-wave radiation to dissipate some of the absorbed solar energy. In winter, glazing temperatures usually fall below room air temperature, which may produce thermal discomfort to occupants near the fenestration. In summer, individuals seated near unshaded fenestration may experience discomfort from both direct solar rays and long-wave radiation emitted by sunheated glazing. In winter, loss of heat by radiation to cold glazing can also cause discomfort. Tightly woven, highly reflective drapes minimize such discomfort; drapes with high openness factors are less effective because they allow short- and long-wave radiation to pass more freely. Light-colored shading devices with maximum total surface usually provide the best protection because they absorb less heat and tend to lose heat readily by convection to the conditioned air. Outward Vision. Outward vision is normally desirable in both business and living spaces. Open-weave, dark-colored fabrics of uniform pattern allow maximum outward vision, whereas uneven pattern weaves reduce the ability to see out. A semiopen weave modifies the view without completely obscuring the outdoors. Tightly woven fabrics block outward vision completely. Privacy. Venetian blinds, either vertical or horizontal, can be adjusted and, when completely closed, afford full privacy. When draperies are closed, the degree of privacy is determined by their color and tightness of weave and the source of the principal illumination. To obscure the view so completely that not even shadows or silhouettes can be detected, fully opaque materials are used. Generally, the more brightly lit side of a partially shaded glazing is the most visible from the opposite side, making the indoors fairly private in daytime, but not at night. Brightness Control. Visual comfort is essential in many occupied areas, and freedom from glare is an important factor in performing tasks. Discomfort glare is produced by uneven brightness in occupied spaces, with areas or spots that are much brighter than surrounding surfaces. Fenestrations themselves, when they look out onto bright skies or brightly reflecting surfaces, can be glare sources if care is not taken to keep surrounding brightness comparable. A maximum brightness ratio of about 3 to 1 is sometimes quoted. This ratio can be moderated by indoor furnishings and wall coverings, which on average have moderately high diffuse reflectances and access to admitted daylight. Conversely, dark indoor surfaces, and those shaded from daylight illumination, accentuate the brightness difference between the fenestration and its surroundings. Indoor surface brightness can also be elevated by ample use of indoor electric lighting, but this can have adverse consequences for the building’s energy use. In general, larger fenestration apertures admit more sunlight, increasing indoor brightness without affecting the perceived brightness of the fenestration, all other factors being equal. An important guideline is that direct sunlight must not strike the eye, and reflected sunlight from bright or shiny surfaces is equally disturbing and even disabling. A tightly woven white fabric with high solar transmittance attains such brilliance when illuminated by direct sunshine that, by contrast with its surroundings, it creates excessive glare. Off-white colors should be used so their surface brightness is not too great. Venetian blinds allow considerable light to enter by interreflection between slats. When two shading devices are used, the one on the indoors (away from the fenestration product) should be darker and more open. With this arrangement, the indoor
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Louver Location Louver Reflection Indoor Side
0.15
Worsta
Excluded Beamb 45° Closed 0.50
Worsta 0° Excluded Beamb 45°
Licensed for single user. © 2017 ASHRAE, Inc.
Closed Indoor Side
0.80
Worsta 0° Excluded Beamb 45° Closed
Outdoor Side
0.15
Worsta 0° Excluded Beamb 45° Closed
Outdoor Side
0.50
1b
1c
0°
Indoor Side
1a
Worst a 0° Excluded Beamb 45°
1e
1f
1g
1h
1i
IAC 0 (IAC 60)/IACdiff , FR d 0.98 (0.97)/0.86 0.92 0.98 (0.78)/0.87 0.69 0.73 (0.78)/0.87 0.43 0.80 (0.74)/0.83 0.47 0.70 (0.70)/0.73 0.44
0.98 (0.97)/0.86 0.91 0.98 (0.79)/0.87 0.68 0.74 (0.79)/0.87 0.43 0.80 (0.75)/0.83 0.46 0.70 (0.70)/0.74 0.44
0.98 (0.96)/0.86 0.88 0.98 (0.80)/0.88 0.66 0.75 (0.80)/0.88 0.42 0.81 (0.76)/0.84 0.45 0.72 (0.72)/0.75 0.42
0.97 (0.95)/0.87 0.82 0.97 (0.82)/0.88 0.64 0.77 (0.82)/0.88 0.41 0.82 (0.78)/0.85 0.44 0.74 (0.74)/0.76 0.4
0.98 (0.96)/0.87 0.87 0.98 (0.80)/0.88 0.66 0.76 (0.80)/0.88 0.42 0.81 (0.77)/0.84 0.45 0.72 (0.72)/0.75 0.42
0.97 (0.95)/0.87 0.82 0.97 (0.82)/0.89 0.63 0.77 (0.82)/0.88 0.41 0.83 (0.79)/0.85 0.43 0.74 (0.74)/0.77 0.4
0.98 (0.96)/0.87 0.87 0.98 (0.80)/0.88 0.66 0.76 (0.80)/0.88 0.42 0.81 (0.77)/0.84 0.45 0.72 (0.72)/0.75 0.42
0.97 (0.95)/0.87 0.81 0.97 (0.82)/0.89 0.63 0.78 (0.82)/0.88 0.41 0.83 (0.79)/0.85 0.43 0.74 (0.74)/0.77 0.4
0.97 (0.95)/0.87 0.83 0.97 (0.82)/0.88 0.64 0.77 (0.82)/0.88 0.41 0.82 (0.78)/0.85 0.44 0.74 (0.74)/0.76 0.4
0.98 (0.96)/0.80 0.94 0.98 (0.70)/0.83 0.74 0.59 (0.70)/0.82 0.5 0.69 (0.58)/0.74 0.53 0.51 (0.49)/0.58 0.46
0.97 (0.96)/0.80 0.93 0.97 (0.70)/0.84 0.73 0.60 (0.70)/0.83 0.5 0.70 (0.59)/0.75 0.52 0.52 (0.50)/0.58 0.45
0.97 (0.96)/0.81 0.89 0.97 (0.72)/0.84 0.71 0.63 (0.72)/0.84 0.48 0.72 (0.62)/0.76 0.5 0.55 (0.53)/0.61 0.43
0.97 (0.95)/0.83 0.83 0.97 (0.75)/0.86 0.67 0.67 (0.75)/0.85 0.46 0.75 (0.66)/0.79 0.48 0.60 (0.58)/0.65 0.4
0.97 (0.96)/0.82 0.88 0.97 (0.73)/0.85 0.7 0.64 (0.73)/0.84 0.48 0.73 (0.63)/0.77 0.5 0.56 (0.54)/0.62 0.42
0.97 (0.95)/0.83 0.83 0.97 (0.76)/0.86 0.67 0.67 (0.76)/0.85 0.46 0.75 (0.67)/0.79 0.47 0.60 (0.59)/0.65 0.4
0.97 (0.96)/0.82 0.88 0.97 (0.73)/0.85 0.7 0.64 (0.73)/0.84 0.48 0.73 (0.63)/0.77 0.5 0.56 (0.54)/0.62 0.42
0.97 (0.95)/0.83 0.82 0.97 (0.76)/0.86 0.66 0.67 (0.76)/0.85 0.46 0.75 (0.67)/0.79 0.47 0.61 (0.59)/0.66 0.39
0.97 (0.95)/0.83 0.84 0.97 (0.75)/0.86 0.67 0.67 (0.75)/0.85 0.46 0.75 (0.66)/0.79 0.48 0.60 (0.58)/0.65 0.4
0.97 (0.96)/0.73 0.95 0.97 (0.60)/0.78 0.82 0.45 (0.60)/0.77 0.66 0.59 (0.42)/0.66 0.66 0.33 (0.29)/0.43 0.52
0.97 (0.96)/0.74 0.94 0.97 (0.61)/0.79 0.81 0.47 (0.61)/0.77 0.65 0.60 (0.43)/0.67 0.65 0.35 (0.31)/0.44 0.51
0.97 (0.95)/0.76 0.9 0.97 (0.64)/0.80 0.78 0.51 (0.64)/0.79 0.61 0.63 (0.48)/0.69 0.61 0.40 (0.37)/0.49 0.46
0.96 (0.95)/0.78 0.85 0.96 (0.68)/0.82 0.73 0.57 (0.68)/0.81 0.56 0.67 (0.54)/0.73 0.56 0.47 (0.44)/0.54 0.41
0.97 (0.95)/0.76 0.89 0.97 (0.65)/0.81 0.77 0.52 (0.65)/0.79 0.59 0.64 (0.49)/0.70 0.59 0.42 (0.38)/0.50 0.45
0.96 (0.95)/0.79 0.84 0.96 (0.69)/0.83 0.72 0.57 (0.69)/0.81 0.55 0.68 (0.55)/0.73 0.55 0.48 (0.45)/0.55 0.41
0.97 (0.95)/0.76 0.89 0.97 (0.65)/0.81 0.77 0.52 (0.65)/0.79 0.59 0.64 (0.49)/0.70 0.59 0.42 (0.38)/0.50 0.45
0.96 (0.95)/0.79 0.83 0.96 (0.69)/0.83 0.72 0.58 (0.69)/0.82 0.55 0.68 (0.56)/0.73 0.55 0.49 (0.46)/0.55 0.41
0.96 (0.95)/0.78 0.85 0.96 (0.68)/0.82 0.73 0.57 (0.68)/0.81 0.56 0.68 (0.54)/0.73 0.56 0.47 (0.44)/0.54 0.42
0.93 (0.89)/0.36 0.98 0.93 (0.05)/0.41 0.9 0.04 (0.05)/0.39 0.77 0.20 (0.04)/0.29 0.83 0.03 (0.03)/0.11 0.65
0.93 (0.89)/0.36 0.97 0.93 (0.06)/0.41 0.89 0.04 (0.06)/0.40 0.76 0.20 (0.04)/0.30 0.82 0.03 (0.04)/0.11 0.65
0.93 (0.89)/0.36 0.95 0.93 (0.06)/0.42 0.87 0.04 (0.06)/0.40 0.74 0.21 (0.04)/0.30 0.81 0.04 (0.04)/0.11 0.65
0.93 (0.89)/0.37 0.92 0.93 (0.06)/0.42 0.84 0.05 (0.06)/0.40 0.72 0.21 (0.05)/0.30 0.78 0.04 (0.05)/0.12 0.64
0.93 (0.89)/0.36 0.94 0.93 (0.06)/0.42 0.87 0.04 (0.06)/0.40 0.74 0.21 (0.04)/0.30 0.8 0.04 (0.04)/0.11 0.64
0.93 (0.89)/0.37 0.91 0.93 (0.06)/0.42 0.84 0.05 (0.06)/0.40 0.72 0.21 (0.05)/0.30 0.78 0.05 (0.05)/0.12 0.64
0.93 (0.89)/0.36 0.94 0.93 (0.06)/0.42 0.87 0.04 (0.06)/0.40 0.74 0.21 (0.04)/0.30 0.8 0.04 (0.04)/0.11 0.64
0.93 (0.89)/0.37 0.91 0.93 (0.07)/0.42 0.83 0.05 (0.07)/0.40 0.72 0.21 (0.05)/0.30 0.77 0.05 (0.05)/0.12 0.64
0.93 (0.89)/0.37 0.92 0.93 (0.06)/0.42 0.84 0.05 (0.06)/0.40 0.72 0.21 (0.05)/0.30 0.78 0.04 (0.05)/0.12 0.64
0.94 (0.95)/0.44 0.98 0.94 (0.15)/0.50 0.96 0.08 (0.15)/0.48 0.92 0.26 (0.07)/0.36 0.93 0.05 (0.03)/0.14 0.8
0.94 (0.95)/0.44 0.97 0.94 (0.15)/0.50 0.96 0.08 (0.15)/0.48 0.91 0.26 (0.07)/0.36 0.92 0.05 (0.03)/0.14 0.8
0.94 (0.95)/0.44 0.95 0.94 (0.15)/0.50 0.93 0.09 (0.15)/0.48 0.89 0.26 (0.07)/0.36 0.89 0.05 (0.03)/0.14 0.77
0.94 (0.95)/0.44 0.92 0.94 (0.15)/0.50 0.9 0.09 (0.15)/0.48 0.85 0.26 (0.07)/0.37 0.86 0.05 (0.04)/0.15 0.75
0.94 (0.95)/0.44 0.94 0.94 (0.15)/0.50 0.92 0.09 (0.15)/0.48 0.88 0.26 (0.07)/0.36 0.89 0.05 (0.04)/0.14 0.77
0.94 (0.95)/0.44 0.91 0.94 (0.15)/0.50 0.89 0.09 (0.15)/0.48 0.85 0.26 (0.07)/0.37 0.85 0.06 (0.04)/0.15 0.74
0.94 (0.95)/0.44 0.94 0.94 (0.15)/0.50 0.92 0.09 (0.15)/0.48 0.88 0.26 (0.07)/0.36 0.89 0.05 (0.04)/0.14 0.77
0.94 (0.95)/0.44 0.91 0.94 (0.15)/0.50 0.89 0.09 (0.15)/0.48 0.84 0.26 (0.07)/0.37 0.85 0.06 (0.04)/0.15 0.74
0.94 (0.95)/0.44 0.92 0.94 (0.15)/0.50 0.9 0.09 (0.15)/0.48 0.85 0.26 (0.07)/0.37 0.86 0.05 (0.04)/0.15 0.75
15.39
Closed
1d
Fenestration
Table 14A IAC Values for Louvered Shades: Uncoated Single Glazings Glazing ID:
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Outdoor Side
0.80
Worsta 0° Excluded Beamb 45° Closed
Sheer Notes:
aLouvers
bLouvers
100%
1a
1b
1c
1d
1e
1f
1g
1h
1i
0.95 (1.02)/0.54 0.98 0.95 (0.28)/0.61 0.98 0.17 (0.28)/0.59 0.97 0.34 (0.12)/0.46 0.97 0.08 (0.04)/0.21 0.93 0.7 0.5
0.95 (1.02)/0.54 0.98 0.95 (0.28)/0.61 0.97 0.17 (0.28)/0.59 0.96 0.34 (0.12)/0.46 0.96 0.08 (0.04)/0.21 0.92 0.71 0.49
0.95 (1.01)/0.54 0.95 0.95 (0.28)/0.61 0.95 0.17 (0.28)/0.58 0.94 0.34 (0.12)/0.46 0.94 0.09 (0.04)/0.20 0.89 0.72 0.47
0.95 (1.00)/0.54 0.92 0.95 (0.28)/0.61 0.91 0.17 (0.28)/0.58 0.9 0.34 (0.12)/0.46 0.9 0.09 (0.04)/0.21 0.86 0.74 0.45
0.95 (1.01)/0.54 0.95 0.95 (0.28)/0.61 0.94 0.17 (0.28)/0.58 0.93 0.34 (0.12)/0.46 0.93 0.09 (0.04)/0.20 0.89 0.72 0.47
0.95 (1.00)/0.54 0.91 0.95 (0.28)/0.61 0.91 0.17 (0.28)/0.58 0.89 0.34 (0.12)/0.46 0.9 0.09 (0.04)/0.21 0.85 0.74 0.44
0.95 (1.01)/0.54 0.95 0.95 (0.28)/0.61 0.94 0.17 (0.28)/0.58 0.93 0.34 (0.12)/0.46 0.93 0.09 (0.04)/0.20 0.89 0.72 0.47
0.95 (1.00)/0.54 0.91 0.95 (0.28)/0.61 0.91 0.17 (0.28)/0.59 0.89 0.34 (0.12)/0.46 0.89 0.09 (0.04)/0.21 0.85 0.74 0.44
0.95 (1.01)/0.54 0.92 0.95 (0.28)/0.61 0.91 0.17 (0.28)/0.59 0.9 0.35 (0.12)/0.46 0.9 0.09 (0.04)/0.21 0.86 0.74 0.45
15.40
Table 14A IAC Values for Louvered Shades: Uncoated Single Glazings (Continued) Glazing ID:
cGlazing
track so that profile angle equals negative slat angle and maximum direct beam is admitted. track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
dF R
cavity width equals original cavity width plus slat width. is radiant fraction; ratio of radiative heat transfer to total heat transfer, on room side of glazing system.
Table 14B IAC Values for Louvered Shades: Uncoated Double Glazings Glazing ID: Louver Location Louver Reflection Indoor Side
0.15
Worsta 0° Excluded Beamb
Closed 0.50
Worsta 0° Excluded Beamb 45° Closed
Indoor Side
0.80
Worsta 0° Excluded Beamb 45° Closed
Between Glazingsc
0.15
5b
5c
Worsta 0° Excluded Beamb 45°
5d
5e
5f
5g
5h
5i
0.98 (0.97)/0.93 0.78 0.98 (0.90)/0.94 0.62 0.86 (0.90)/0.93 0.4 0.89 (0.87)/0.91 0.42 0.83 (0.84)/0.85 0.38 0.98 (0.97)/0.89 0.79 0.98 (0.84)/0.91 0.65 0.76 (0.84)/0.90 0.45 0.82 (0.76)/0.85 0.46 0.70 (0.70)/0.75 0.38 0.97 (0.96)/0.84 0.8 0.97 (0.77)/0.87 0.7 0.66 (0.77)/0.86 0.53 0.75 (0.66)/0.79 0.53 0.57 (0.57)/0.64 0.39 0.95 (0.95)/0.69 0.88 0.95 (0.55)/0.72 0.78 0.49 (0.55)/0.71 0.65 0.58 (0.52)/0.65
0.99 (0.97)/0.92 0.83 0.99 (0.89)/0.93 0.65 0.85 (0.89)/0.93 0.41 0.88 (0.86)/0.90 0.44 0.82 (0.82)/0.84 0.4 0.98 (0.97)/0.87 0.84 0.98 (0.82)/0.90 0.68 0.73 (0.82)/0.89 0.46 0.80 (0.73)/0.84 0.48 0.66 (0.66)/0.71 0.4 0.97 (0.96)/0.82 0.86 0.97 (0.73)/0.85 0.74 0.61 (0.73)/0.84 0.57 0.71 (0.60)/0.76 0.57 0.51 (0.50)/0.59 0.43 0.95 (0.97)/0.67 0.91 0.95 (0.52)/0.70 0.79 0.46 (0.52)/0.69 0.65 0.56 (0.49)/0.63
0.98 (0.97)/0.93 0.78 0.98 (0.90)/0.94 0.62 0.86 (0.90)/0.93 0.4 0.89 (0.87)/0.91 0.42 0.83 (0.84)/0.85 0.38 0.98 (0.97)/0.89 0.79 0.98 (0.84)/0.91 0.65 0.76 (0.84)/0.90 0.45 0.82 (0.76)/0.85 0.46 0.70 (0.70)/0.75 0.38 0.97 (0.96)/0.84 0.8 0.97 (0.77)/0.87 0.7 0.66 (0.77)/0.86 0.53 0.75 (0.66)/0.79 0.53 0.57 (0.57)/0.65 0.39 0.95 (0.95)/0.69 0.88 0.95 (0.55)/0.72 0.78 0.49 (0.55)/0.71 0.65 0.59 (0.52)/0.65
0.98 (0.97)/0.93 0.79 0.98 (0.90)/0.93 0.62 0.86 (0.90)/0.93 0.4 0.89 (0.87)/0.91 0.42 0.83 (0.84)/0.85 0.38 0.98 (0.97)/0.88 0.8 0.98 (0.84)/0.91 0.65 0.75 (0.84)/0.90 0.45 0.82 (0.76)/0.85 0.46 0.70 (0.70)/0.74 0.38 0.97 (0.96)/0.84 0.81 0.97 (0.77)/0.87 0.71 0.65 (0.77)/0.86 0.54 0.74 (0.65)/0.79 0.54 0.57 (0.56)/0.64 0.4 0.95 (0.96)/0.69 0.88 0.95 (0.55)/0.72 0.78 0.49 (0.55)/0.71 0.65 0.58 (0.52)/0.65
IAC 0 (IAC 60)/IACdiff , FRd 0.99 (0.98)/0.92 0.87 0.99 (0.88)/0.93 0.67 0.84 (0.88)/0.93 0.42 0.88 (0.85)/0.90 0.45 0.81 (0.81)/0.83 0.41 0.98 (0.97)/0.86 0.88 0.98 (0.80)/0.89 0.71 0.70 (0.80)/0.88 0.48 0.78 (0.70)/0.82 0.5 0.63 (0.63)/0.69 0.42 0.97 (0.96)/0.80 0.9 0.97 (0.71)/0.84 0.78 0.57 (0.71)/0.83 0.6 0.68 (0.56)/0.74 0.6 0.47 (0.45)/0.55 0.46 0.97 (0.98)/0.66 0.93 0.97 (0.50)/0.70 0.81 0.43 (0.50)/0.69 0.66 0.54 (0.47)/0.62
0.99 (0.98)/0.92 0.84 0.99 (0.89)/0.93 0.65 0.84 (0.89)/0.93 0.41 0.88 (0.85)/0.90 0.44 0.82 (0.82)/0.84 0.4 0.98 (0.97)/0.87 0.86 0.98 (0.82)/0.90 0.69 0.72 (0.82)/0.89 0.47 0.80 (0.72)/0.83 0.49 0.66 (0.65)/0.71 0.4 0.97 (0.96)/0.81 0.87 0.97 (0.73)/0.85 0.75 0.60 (0.73)/0.84 0.58 0.71 (0.60)/0.76 0.58 0.51 (0.50)/0.59 0.43 0.97 (0.99)/0.67 0.91 0.97 (0.51)/0.71 0.8 0.45 (0.51)/0.69 0.66 0.55 (0.48)/0.63
0.99 (0.97)/0.92 0.84 0.99 (0.89)/0.93 0.65 0.84 (0.89)/0.93 0.41 0.88 (0.85)/0.90 0.44 0.82 (0.82)/0.84 0.4 0.98 (0.97)/0.87 0.85 0.98 (0.81)/0.90 0.69 0.72 (0.81)/0.89 0.47 0.79 (0.72)/0.83 0.48 0.65 (0.65)/0.71 0.4 0.97 (0.96)/0.81 0.87 0.97 (0.73)/0.85 0.75 0.60 (0.73)/0.84 0.58 0.70 (0.59)/0.76 0.58 0.50 (0.49)/0.58 0.43 0.96 (0.97)/0.67 0.91 0.96 (0.52)/0.70 0.8 0.46 (0.52)/0.69 0.66 0.56 (0.49)/0.63
0.98 (0.97)/0.93 0.79 0.98 (0.90)/0.93 0.62 0.86 (0.90)/0.93 0.4 0.89 (0.87)/0.91 0.42 0.83 (0.84)/0.85 0.38 0.98 (0.97)/0.88 0.8 0.98 (0.84)/0.91 0.65 0.75 (0.84)/0.90 0.45 0.82 (0.76)/0.85 0.46 0.70 (0.70)/0.74 0.38 0.97 (0.96)/0.84 0.81 0.97 (0.77)/0.87 0.7 0.65 (0.77)/0.86 0.54 0.74 (0.65)/0.79 0.54 0.57 (0.56)/0.64 0.4 0.95 (0.96)/0.69 0.88 0.95 (0.55)/0.72 0.78 0.49 (0.55)/0.71 0.65 0.58 (0.52)/0.65
0.99 (0.97)/0.92 0.83 0.99 (0.89)/0.93 0.65 0.85 (0.89)/0.93 0.41 0.88 (0.86)/0.90 0.44 0.82 (0.82)/0.84 0.4 0.98 (0.97)/0.87 0.84 0.98 (0.82)/0.90 0.68 0.73 (0.82)/0.89 0.46 0.80 (0.73)/0.84 0.48 0.66 (0.66)/0.71 0.4 0.97 (0.96)/0.82 0.86 0.97 (0.73)/0.85 0.74 0.61 (0.73)/0.84 0.57 0.71 (0.60)/0.76 0.57 0.51 (0.50)/0.59 0.43 0.95 (0.97)/0.67 0.91 0.95 (0.52)/0.70 0.79 0.46 (0.52)/0.69 0.65 0.56 (0.49)/0.63
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
45°
Indoor Side
5a
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Closed Between Glazingsc
0.50
Worsta 0° Excluded Beamb 45° Closed
Between Glazingsc
0.80
Worsta 0° Excluded Beamb 45°
Licensed for single user. © 2017 ASHRAE, Inc.
Closed Outdoor Side
0.15
Worsta 0° Excluded Beamb 45° Closed
Outdoor Side
0.50
Worsta 0° Excluded Beamb 45° Closed
Outdoor Side
0.80
Worsta 0° Excluded Beamb 45° Closed
aLouvers bLouvers
5b 0.7 0.44 (0.47)/0.51 0.65 0.97 (1.02)/0.67 0.92 0.97 (0.50)/0.72 0.83 0.39 (0.50)/0.70 0.71 0.52 (0.40)/0.62 0.74 0.33 (0.33)/0.43 0.66 0.97 (1.04)/0.68 0.93 0.97 (0.50)/0.74 0.88 0.36 (0.50)/0.72 0.81 0.51 (0.33)/0.61 0.81 0.25 (0.22)/0.37 0.73 0.93 (0.89)/0.36 0.94 0.93 (0.05)/0.41 0.88 0.03 (0.05)/0.39 0.77 0.19 (0.03)/0.29 0.82 0.02 (0.02)/0.10 0.66 0.94 (0.98)/0.44 0.94 0.94 (0.14)/0.50 0.93 0.07 (0.14)/0.48 0.9 0.25 (0.06)/0.36 0.9 0.04 (0.03)/0.14 0.81 0.95 (1.07)/0.55 0.94 0.95 (0.29)/0.61 0.94 0.16 (0.29)/0.59 0.93 0.34 (0.12)/0.47 0.93 0.08 (0.04)/0.21 0.9
track so that profile angle equals negative slat angle and maximum direct beam is admitted. track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
5c 0.7 0.44 (0.47)/0.52 0.65 0.96 (1.00)/0.67 0.92 0.96 (0.51)/0.72 0.83 0.40 (0.51)/0.70 0.71 0.53 (0.41)/0.62 0.73 0.35 (0.35)/0.44 0.66 0.96 (1.02)/0.69 0.93 0.96 (0.51)/0.74 0.87 0.37 (0.51)/0.72 0.8 0.52 (0.35)/0.62 0.8 0.27 (0.25)/0.39 0.71 0.93 (0.89)/0.36 0.93 0.93 (0.05)/0.41 0.88 0.03 (0.05)/0.39 0.76 0.20 (0.03)/0.29 0.82 0.02 (0.03)/0.10 0.65 0.94 (0.97)/0.44 0.93 0.94 (0.15)/0.50 0.92 0.08 (0.15)/0.48 0.89 0.25 (0.06)/0.36 0.9 0.04 (0.03)/0.14 0.8 0.95 (1.04)/0.55 0.93 0.95 (0.28)/0.61 0.94 0.16 (0.28)/0.59 0.92 0.33 (0.12)/0.47 0.93 0.08 (0.04)/0.21 0.89
5d 0.69 0.47 (0.50)/0.54 0.64 0.95 (0.98)/0.69 0.89 0.95 (0.54)/0.73 0.8 0.44 (0.54)/0.72 0.69 0.56 (0.45)/0.64 0.72 0.39 (0.39)/0.47 0.65 0.95 (1.01)/0.70 0.9 0.95 (0.55)/0.75 0.84 0.42 (0.55)/0.74 0.76 0.55 (0.40)/0.64 0.77 0.32 (0.30)/0.43 0.69 0.93 (0.89)/0.36 0.9 0.93 (0.05)/0.41 0.84 0.03 (0.05)/0.39 0.73 0.20 (0.04)/0.29 0.79 0.03 (0.03)/0.11 0.64 0.94 (0.96)/0.44 0.9 0.94 (0.15)/0.50 0.89 0.08 (0.15)/0.48 0.85 0.25 (0.07)/0.36 0.86 0.04 (0.03)/0.14 0.76 0.95 (1.02)/0.54 0.9 0.95 (0.28)/0.61 0.9 0.16 (0.28)/0.59 0.89 0.33 (0.12)/0.46 0.89 0.08 (0.04)/0.21 0.86 cGlazing dF R
5e 0.69 0.45 (0.48)/0.52 0.64 0.95 (0.99)/0.68 0.92 0.95 (0.52)/0.72 0.82 0.41 (0.52)/0.71 0.71 0.53 (0.42)/0.62 0.73 0.36 (0.36)/0.45 0.66 0.96 (1.02)/0.69 0.93 0.96 (0.52)/0.74 0.87 0.38 (0.52)/0.72 0.79 0.53 (0.36)/0.62 0.8 0.28 (0.26)/0.40 0.71 0.93 (0.89)/0.36 0.93 0.93 (0.05)/0.41 0.87 0.03 (0.05)/0.39 0.76 0.20 (0.03)/0.29 0.81 0.03 (0.03)/0.10 0.65 0.94 (0.96)/0.44 0.93 0.94 (0.15)/0.50 0.92 0.08 (0.15)/0.48 0.88 0.25 (0.06)/0.36 0.89 0.04 (0.03)/0.14 0.79 0.95 (1.04)/0.55 0.93 0.95 (0.28)/0.61 0.93 0.16 (0.28)/0.59 0.92 0.33 (0.12)/0.46 0.92 0.08 (0.04)/0.21 0.89
5f 0.68 0.48 (0.51)/0.54 0.64 0.95 (0.98)/0.69 0.89 0.95 (0.55)/0.73 0.8 0.45 (0.55)/0.72 0.69 0.56 (0.46)/0.64 0.71 0.39 (0.40)/0.48 0.65 0.95 (1.00)/0.70 0.9 0.95 (0.55)/0.75 0.84 0.42 (0.55)/0.74 0.76 0.55 (0.41)/0.64 0.77 0.33 (0.31)/0.43 0.69 0.93 (0.89)/0.36 0.89 0.93 (0.06)/0.41 0.84 0.03 (0.06)/0.39 0.73 0.20 (0.04)/0.29 0.78 0.03 (0.03)/0.11 0.64 0.94 (0.95)/0.44 0.9 0.94 (0.15)/0.50 0.88 0.08 (0.15)/0.48 0.85 0.25 (0.07)/0.36 0.85 0.05 (0.03)/0.14 0.76 0.95 (1.02)/0.54 0.9 0.95 (0.28)/0.61 0.9 0.16 (0.28)/0.59 0.88 0.33 (0.12)/0.46 0.88 0.08 (0.04)/0.21 0.85
5g 0.69 0.45 (0.48)/0.52 0.64 0.95 (0.99)/0.68 0.92 0.95 (0.52)/0.72 0.82 0.41 (0.52)/0.71 0.71 0.53 (0.42)/0.62 0.73 0.36 (0.36)/0.45 0.66 0.96 (1.02)/0.69 0.93 0.96 (0.52)/0.74 0.87 0.38 (0.52)/0.72 0.79 0.53 (0.36)/0.62 0.8 0.28 (0.26)/0.40 0.71 0.93 (0.89)/0.36 0.93 0.93 (0.05)/0.41 0.87 0.03 (0.05)/0.39 0.76 0.20 (0.03)/0.29 0.81 0.03 (0.03)/0.10 0.65 0.94 (0.96)/0.44 0.93 0.94 (0.15)/0.50 0.92 0.08 (0.15)/0.48 0.88 0.25 (0.06)/0.36 0.89 0.04 (0.03)/0.14 0.79 0.95 (1.04)/0.55 0.93 0.95 (0.28)/0.61 0.93 0.16 (0.28)/0.59 0.92 0.33 (0.12)/0.46 0.92 0.08 (0.04)/0.21 0.89
5h 0.68 0.48 (0.51)/0.54 0.64 0.95 (0.98)/0.69 0.89 0.95 (0.55)/0.73 0.8 0.45 (0.55)/0.72 0.69 0.56 (0.46)/0.64 0.71 0.40 (0.40)/0.48 0.65 0.95 (1.00)/0.70 0.89 0.95 (0.55)/0.75 0.84 0.42 (0.55)/0.74 0.76 0.56 (0.41)/0.64 0.77 0.33 (0.31)/0.43 0.69 0.93 (0.89)/0.36 0.89 0.93 (0.06)/0.41 0.84 0.04 (0.06)/0.40 0.73 0.20 (0.04)/0.30 0.78 0.03 (0.03)/0.11 0.64 0.94 (0.95)/0.44 0.89 0.94 (0.15)/0.50 0.88 0.08 (0.15)/0.48 0.84 0.25 (0.07)/0.36 0.85 0.05 (0.03)/0.14 0.76 0.95 (1.02)/0.54 0.89 0.95 (0.28)/0.61 0.89 0.16 (0.28)/0.59 0.88 0.33 (0.12)/0.46 0.88 0.08 (0.04)/0.21 0.85
cavity width equals original cavity width plus slat width. is radiant fraction; ratio of radiative heat transfer to total heat transfer, on room side of glazing system.
5i 0.69 0.47 (0.50)/0.54 0.64 0.95 (0.98)/0.69 0.89 0.95 (0.54)/0.73 0.8 0.44 (0.54)/0.72 0.69 0.56 (0.45)/0.64 0.72 0.39 (0.39)/0.47 0.66 0.95 (1.01)/0.70 0.9 0.95 (0.55)/0.75 0.84 0.42 (0.55)/0.74 0.77 0.55 (0.40)/0.64 0.77 0.32 (0.30)/0.43 0.69 0.93 (0.89)/0.36 0.9 0.93 (0.05)/0.41 0.84 0.03 (0.05)/0.39 0.73 0.20 (0.04)/0.29 0.79 0.03 (0.03)/0.11 0.64 0.94 (0.96)/0.44 0.9 0.94 (0.15)/0.50 0.89 0.08 (0.15)/0.48 0.85 0.25 (0.07)/0.36 0.86 0.04 (0.03)/0.14 0.76 0.95 (1.03)/0.54 0.9 0.95 (0.28)/0.61 0.9 0.16 (0.28)/0.59 0.89 0.34 (0.12)/0.46 0.89 0.08 (0.04)/0.21 0.86
15.41
Notes:
5a 0.7 0.42 (0.45)/0.50 0.65 0.97 (1.01)/0.67 0.94 0.97 (0.49)/0.71 0.84 0.38 (0.49)/0.70 0.72 0.51 (0.38)/0.61 0.75 0.32 (0.32)/0.42 0.67 0.97 (1.04)/0.68 0.94 0.97 (0.49)/0.73 0.89 0.35 (0.49)/0.72 0.82 0.50 (0.32)/0.60 0.83 0.24 (0.20)/0.36 0.74 0.93 (0.89)/0.35 0.95 0.93 (0.05)/0.41 0.9 0.03 (0.05)/0.39 0.78 0.19 (0.03)/0.29 0.84 0.02 (0.02)/0.10 0.66 0.94 (0.98)/0.44 0.95 0.94 (0.14)/0.50 0.94 0.07 (0.14)/0.48 0.91 0.25 (0.06)/0.36 0.92 0.04 (0.02)/0.14 0.82 0.95 (1.08)/0.55 0.95 0.95 (0.29)/0.62 0.95 0.16 (0.29)/0.59 0.94 0.34 (0.12)/0.47 0.95 0.08 (0.04)/0.21 0.92
Fenestration
Table 14B IAC Values for Louvered Shades: Uncoated Double Glazings (Continued) Glazing ID:
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Louver Location Louver Reflection Indoor Side
0.15
17a
17b
17c
17d
17e
17f
17g
17h
17i
17j
17k
IAC 0 (IAC 60)/IACdiff , FR d
Worsta
0.99 (0.98)/0.94 0.99 (0.98)/0.94 0.99 (0.98)/0.94 0.99 (0.98)/0.94 0.99 (0.98)/0.94 0.99 (0.98)/0.95 0.99 (0.98)/0.94 0.99 (0.98)/0.95 0.99 (0.98)/0.94 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.86 0.83 0.83 0.79 0.81 0.76 0.8 0.75 0.8 0.75 0.76
0°
0.99 (0.91)/0.95 0.99 (0.91)/0.95 0.99 (0.91)/0.95 0.99 (0.92)/0.95 0.99 (0.92)/0.95 0.99 (0.93)/0.95 0.99 (0.92)/0.95 0.99 (0.93)/0.95 0.99 (0.92)/0.95 0.99 (0.93)/0.95 0.99 (0.93)/0.95 0.66 0.64 0.65 0.63 0.63 0.61 0.63 0.6 0.63 0.6 0.61
15.42
Table 14C IAC Values for Louvered Shades: Coated Double Glazings with 0.2 Low-e Glazing ID:
Excluded Beamb 0.87 (0.91)/0.94 0.88 (0.91)/0.95 0.88 (0.91)/0.95 0.89 (0.92)/0.95 0.88 (0.92)/0.95 0.90 (0.93)/0.95 0.88 (0.92)/0.95 0.90 (0.93)/0.95 0.88 (0.92)/0.95 0.90 (0.93)/0.95 0.90 (0.93)/0.95 0.41 0.41 0.41 0.41 0.41 0.4 0.41 0.4 0.41 0.4 0.4
Indoor Side
0.50
45°
0.90 (0.88)/0.92 0.91 (0.89)/0.93 0.91 (0.89)/0.93 0.92 (0.89)/0.93 0.91 (0.89)/0.93 0.92 (0.90)/0.93 0.91 (0.89)/0.93 0.92 (0.90)/0.93 0.91 (0.89)/0.93 0.92 (0.90)/0.93 0.92 (0.90)/0.93 0.44 0.43 0.44 0.43 0.43 0.42 0.43 0.42 0.43 0.42 0.42
Closed
0.85 (0.85)/0.87 0.86 (0.86)/0.88 0.85 (0.86)/0.88 0.86 (0.87)/0.88 0.86 (0.86)/0.88 0.87 (0.88)/0.89 0.86 (0.86)/0.88 0.87 (0.88)/0.89 0.86 (0.86)/0.88 0.87 (0.88)/0.89 0.87 (0.88)/0.89 0.4 0.39 0.4 0.39 0.39 0.37 0.39 0.37 0.39 0.37 0.37
Worsta
0.98 (0.98)/0.88 0.98 (0.98)/0.89 0.98 (0.98)/0.89 0.98 (0.98)/0.90 0.98 (0.98)/0.90 0.98 (0.98)/0.91 0.98 (0.98)/0.90 0.98 (0.98)/0.91 0.98 (0.98)/0.90 0.98 (0.98)/0.91 0.98 (0.98)/0.91 0.87 0.84 0.84 0.8 0.82 0.76 0.81 0.76 0.81 0.76 0.77
0°
0.98 (0.83)/0.91 0.98 (0.85)/0.91 0.98 (0.85)/0.91 0.98 (0.86)/0.92 0.98 (0.85)/0.92 0.98 (0.87)/0.93 0.98 (0.85)/0.92 0.98 (0.87)/0.93 0.98 (0.85)/0.92 0.98 (0.87)/0.93 0.98 (0.87)/0.93 0.7 0.68 0.68 0.66 0.67 0.63 0.66 0.63 0.66 0.63 0.64
Excluded Beamb 0.74 (0.83)/0.90 0.76 (0.85)/0.91 0.76 (0.85)/0.91 0.79 (0.86)/0.92 0.77 (0.85)/0.91 0.80 (0.87)/0.92 0.77 (0.85)/0.91 0.80 (0.87)/0.92 0.77 (0.85)/0.91 0.81 (0.87)/0.92 0.80 (0.87)/0.92 0.47 0.46 0.46 0.45 0.45 0.44 0.45 0.44 0.45 0.44 0.44
0.80
0.81 (0.74)/0.85 0.83 (0.76)/0.86 0.83 (0.76)/0.86 0.84 (0.79)/0.87 0.83 (0.77)/0.86 0.86 (0.81)/0.88 0.83 (0.77)/0.87 0.86 (0.81)/0.88 0.83 (0.77)/0.87 0.86 (0.81)/0.88 0.86 (0.81)/0.88 0.49 0.48 0.48 0.47 0.47 0.45 0.47 0.45 0.47 0.45 0.45
Closed
0.67 (0.67)/0.73 0.70 (0.70)/0.75 0.70 (0.70)/0.75 0.73 (0.73)/0.78 0.71 (0.71)/0.76 0.75 (0.75)/0.79 0.72 (0.71)/0.76 0.76 (0.75)/0.79 0.72 (0.71)/0.76 0.76 (0.76)/0.80 0.75 (0.75)/0.79 0.41 0.39 0.4 0.38 0.39 0.37 0.38 0.37 0.38 0.37 0.37
Worsta
0.98 (0.97)/0.82 0.98 (0.97)/0.83 0.98 (0.97)/0.83 0.98 (0.97)/0.85 0.98 (0.97)/0.84 0.98 (0.97)/0.86 0.98 (0.97)/0.84 0.98 (0.97)/0.87 0.98 (0.97)/0.84 0.98 (0.97)/0.87 0.98 (0.97)/0.87 0.88 0.85 0.85 0.82 0.83 0.78 0.83 0.77 0.83 0.77 0.78
0°
0.98 (0.74)/0.86 0.98 (0.76)/0.87 0.98 (0.76)/0.87 0.98 (0.79)/0.88 0.98 (0.77)/0.88 0.98 (0.81)/0.89 0.98 (0.77)/0.88 0.98 (0.81)/0.89 0.98 (0.77)/0.88 0.98 (0.81)/0.89 0.98 (0.81)/0.89 0.76 0.74 0.74 0.71 0.72 0.68 0.72 0.68 0.72 0.68 0.68
Excluded Beamb 0.61 (0.74)/0.84 0.64 (0.76)/0.86 0.64 (0.76)/0.86 0.68 (0.79)/0.87 0.66 (0.77)/0.86 0.71 (0.81)/0.88 0.66 (0.77)/0.87 0.71 (0.81)/0.89 0.66 (0.77)/0.87 0.72 (0.81)/0.89 0.71 (0.81)/0.88 0.59 0.56 0.56 0.54 0.55 0.51 0.54 0.51 0.54 0.51 0.51
Between Glazingsc
0.15
45°
0.71 (0.60)/0.77 0.74 (0.64)/0.79 0.74 (0.64)/0.79 0.77 (0.68)/0.81 0.75 (0.66)/0.80 0.79 (0.71)/0.83 0.75 (0.66)/0.80 0.79 (0.71)/0.83 0.75 (0.66)/0.80 0.79 (0.72)/0.83 0.79 (0.71)/0.83 0.59 0.56 0.56 0.54 0.55 0.51 0.54 0.51 0.54 0.51 0.51
Closed
0.51 (0.50)/0.59 0.55 (0.54)/0.63 0.55 (0.55)/0.63 0.60 (0.60)/0.67 0.57 (0.57)/0.65 0.64 (0.64)/0.70 0.58 (0.57)/0.65 0.64 (0.64)/0.70 0.58 (0.57)/0.65 0.64 (0.64)/0.70 0.64 (0.64)/0.70 0.44 0.42 0.42 0.4 0.41 0.38 0.4 0.38 0.4 0.38 0.38
Worsta
0.97 (1.02)/0.78 0.98 (1.02)/0.78 0.96 (0.96)/0.59 0.96 (0.96)/0.60 0.95 (0.95)/0.60 0.95 (0.95)/0.62 0.95 (0.95)/0.60 0.95 (0.95)/0.62 0.95 (0.95)/0.60 0.95 (0.95)/0.62 0.95 (0.95)/0.62 0.92 0.9 0.91 0.89 0.9 0.87 0.9 0.87 0.9 0.87 0.87
0°
0.97 (0.66)/0.80 0.98 (0.67)/0.81 0.96 (0.39)/0.63 0.96 (0.40)/0.64 0.95 (0.41)/0.64 0.95 (0.44)/0.65 0.95 (0.41)/0.64 0.95 (0.44)/0.65 0.95 (0.41)/0.64 0.95 (0.44)/0.65 0.95 (0.44)/0.65 0.8 0.79 0.79 0.78 0.79 0.77 0.79 0.77 0.79 0.77 0.77
Excluded Beamb 0.59 (0.66)/0.79 0.60 (0.67)/0.80 0.33 (0.39)/0.62 0.34 (0.40)/0.62 0.35 (0.41)/0.62 0.38 (0.44)/0.64 0.36 (0.41)/0.63 0.39 (0.44)/0.64 0.36 (0.41)/0.63 0.39 (0.44)/0.64 0.38 (0.44)/0.64 0.66 0.66 0.66 0.66 0.65 0.65 0.65 0.65 0.65 0.65 0.65 45° 0.67 (0.62)/0.74 0.68 (0.63)/0.75 0.46 (0.36)/0.54 0.46 (0.37)/0.55 0.47 (0.38)/0.55 0.50 (0.41)/0.57 0.47 (0.38)/0.56 0.50 (0.41)/0.57 0.47 (0.38)/0.56 0.50 (0.41)/0.58 0.49 (0.41)/0.57 0.69 0.69 0.7 0.7 0.7 0.69 0.7 0.69 0.7 0.69 0.69
Between Glazingsc
0.50
Closed
0.57 (0.60)/0.64 0.58 (0.61)/0.65 0.32 (0.34)/0.40 0.33 (0.35)/0.41 0.34 (0.36)/0.42 0.37 (0.40)/0.44 0.35 (0.37)/0.42 0.38 (0.40)/0.45 0.35 (0.37)/0.42 0.38 (0.40)/0.45 0.37 (0.39)/0.44 0.65 0.65 0.64 0.64 0.64 0.64 0.64 0.64 0.64 0.64 0.64
Worsta
0.97 (1.03)/0.75 0.97 (1.03)/0.75 0.96 (1.01)/0.61 0.96 (1.01)/0.62 0.96 (1.00)/0.62 0.95 (0.99)/0.64 0.95 (0.99)/0.62 0.95 (0.98)/0.64 0.95 (0.99)/0.62 0.95 (0.98)/0.64 0.95 (0.99)/0.64 0.92 0.91 0.91 0.89 0.91 0.87 0.9 0.87 0.9 0.87 0.88
0°
0.97 (0.61)/0.79 0.97 (0.62)/0.79 0.96 (0.41)/0.66 0.96 (0.42)/0.67 0.96 (0.43)/0.67 0.95 (0.45)/0.68 0.95 (0.43)/0.67 0.95 (0.46)/0.68 0.95 (0.43)/0.67 0.95 (0.46)/0.69 0.95 (0.45)/0.68 0.83 0.82 0.83 0.82 0.82 0.8 0.82 0.8 0.82 0.8 0.8
Excluded Beamb 0.49 (0.61)/0.77 0.50 (0.62)/0.78 0.31 (0.41)/0.65 0.32 (0.42)/0.65 0.33 (0.43)/0.65 0.36 (0.45)/0.67 0.33 (0.43)/0.66 0.37 (0.46)/0.67 0.33 (0.43)/0.66 0.37 (0.46)/0.67 0.36 (0.45)/0.67 0.7 0.7 0.72 0.72 0.71 0.7 0.71 0.7 0.71 0.7 0.7
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Indoor Side
45°
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Between Glazingsc
0.80
17a
17b
17c
17d
17e
17f
17g
17h
17i
17j
17k
45°
0.61 (0.49)/0.69 0.62 (0.50)/0.70 0.45 (0.31)/0.55 0.46 (0.32)/0.56 0.47 (0.33)/0.56 0.49 (0.37)/0.58 0.47 (0.34)/0.56 0.50 (0.37)/0.58 0.47 (0.34)/0.56 0.50 (0.37)/0.58 0.49 (0.37)/0.58 0.73 0.72 0.75 0.74 0.74 0.72 0.74 0.72 0.74 0.72 0.72
Closed
0.42 (0.42)/0.52 0.43 (0.43)/0.53 0.25 (0.25)/0.35 0.26 (0.26)/0.36 0.28 (0.27)/0.37 0.31 (0.31)/0.40 0.28 (0.28)/0.38 0.32 (0.32)/0.41 0.28 (0.28)/0.38 0.32 (0.32)/0.41 0.31 (0.31)/0.40 0.66 0.66 0.67 0.66 0.66 0.66 0.66 0.65 0.66 0.65 0.66
Worsta
0.97 (1.05)/0.72 0.97 (1.05)/0.72 0.97 (1.05)/0.65 0.97 (1.05)/0.66 0.96 (1.04)/0.66 0.96 (1.02)/0.67 0.96 (1.03)/0.66 0.95 (1.02)/0.67 0.96 (1.03)/0.66 0.95 (1.02)/0.68 0.96 (1.02)/0.67 0.94 0.92 0.92 0.9 0.91 0.88 0.91 0.88 0.91 0.88 0.88
0°
0.97 (0.55)/0.77 0.97 (0.56)/0.78 0.97 (0.45)/0.71 0.97 (0.46)/0.72 0.96 (0.46)/0.72 0.96 (0.49)/0.73 0.96 (0.47)/0.72 0.95 (0.49)/0.73 0.96 (0.47)/0.72 0.95 (0.50)/0.73 0.96 (0.49)/0.73 0.88 0.86 0.88 0.86 0.87 0.84 0.86 0.83 0.86 0.83 0.84
Fenestration
Table 14C IAC Values for Louvered Shades: Coated Double Glazings with 0.2 Low-e (Continued) Glazing ID:
Excluded Beamb 0.40 (0.55)/0.75 0.41 (0.56)/0.76 0.31 (0.45)/0.69 0.32 (0.46)/0.70 0.33 (0.46)/0.70 0.37 (0.49)/0.71 0.34 (0.47)/0.70 0.37 (0.49)/0.71 0.34 (0.47)/0.70 0.37 (0.50)/0.71 0.37 (0.49)/0.71 0.8 0.78 0.82 0.81 0.8 0.77 0.8 0.77 0.8 0.77 0.77
Licensed for single user. © 2017 ASHRAE, Inc.
Outdoor Side
0.15
45°
0.55 (0.37)/0.65 0.56 (0.39)/0.65 0.47 (0.28)/0.58 0.48 (0.29)/0.58 0.49 (0.31)/0.59 0.51 (0.34)/0.61 0.49 (0.31)/0.59 0.51 (0.35)/0.61 0.49 (0.31)/0.59 0.52 (0.35)/0.61 0.51 (0.34)/0.61 0.81 0.79 0.83 0.81 0.81 0.77 0.8 0.77 0.8 0.77 0.78
Closed
0.29 (0.25)/0.41 0.30 (0.27)/0.42 0.21 (0.17)/0.33 0.22 (0.18)/0.34 0.24 (0.20)/0.35 0.27 (0.25)/0.38 0.24 (0.21)/0.36 0.28 (0.25)/0.39 0.24 (0.21)/0.36 0.28 (0.26)/0.39 0.27 (0.25)/0.38 0.72 0.71 0.75 0.73 0.72 0.7 0.72 0.7 0.72 0.69 0.7
Worsta
0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.36 0.93 (0.89)/0.35 0.95 0.93 0.93 0.91 0.92 0.88 0.91 0.88 0.91 0.88 0.88
0°
0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.05)/0.41 0.9 0.88 0.89 0.87 0.87 0.84 0.87 0.83 0.87 0.83 0.84
Excluded Beamb 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.03 (0.05)/0.39 0.02 (0.04)/0.39 0.03 (0.05)/0.39 0.02 (0.04)/0.39 0.03 (0.05)/0.39 0.03 (0.05)/0.39 0.8 0.79 0.8 0.78 0.77 0.74 0.77 0.74 0.77 0.74 0.74
Outdoor Side
0.50
45°
0.19 (0.02)/0.28 0.19 (0.02)/0.29 0.19 (0.02)/0.28 0.19 (0.02)/0.29 0.19 (0.02)/0.29 0.19 (0.03)/0.29 0.19 (0.03)/0.29 0.20 (0.03)/0.29 0.19 (0.03)/0.29 0.20 (0.03)/0.29 0.19 (0.03)/0.29 0.84 0.83 0.84 0.82 0.82 0.79 0.81 0.78 0.81 0.78 0.79
Closed
0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.10 0.02 (0.02)/0.09 0.02 (0.03)/0.10 0.02 (0.02)/0.09 0.02 (0.03)/0.10 0.02 (0.02)/0.10 0.67 0.66 0.67 0.66 0.66 0.65 0.66 0.65 0.66 0.65 0.65
Worsta
0.94 (0.98)/0.43 0.94 (0.98)/0.43 0.94 (0.99)/0.44 0.94 (0.98)/0.44 0.94 (0.97)/0.43 0.94 (0.95)/0.43 0.94 (0.96)/0.43 0.94 (0.95)/0.43 0.94 (0.96)/0.43 0.94 (0.95)/0.43 0.94 (0.96)/0.43 0.95 0.93 0.93 0.91 0.92 0.88 0.91 0.88 0.91 0.88 0.88
0°
0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.49 0.94 (0.14)/0.49 0.94 (0.14)/0.49 0.94 (0.14)/0.49 0.94 (0.14)/0.49 0.94 (0.14)/0.49 0.94 (0.14)/0.49 0.94 0.92 0.92 0.9 0.91 0.87 0.91 0.87 0.91 0.87 0.88
Excluded Beamb 0.07 (0.14)/0.47 0.07 (0.14)/0.47 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.47 0.07 (0.14)/0.47 0.07 (0.14)/0.47 0.07 (0.14)/0.47 0.07 (0.14)/0.47 0.08 (0.14)/0.47 0.07 (0.14)/0.47 0.91 0.9 0.9 0.88 0.88 0.85 0.88 0.84 0.88 0.84 0.85
Outdoor Side
0.80
45°
0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.92 0.9 0.91 0.88 0.89 0.85 0.88 0.85 0.88 0.85 0.85
Closed
0.04 (0.02)/0.13 0.04 (0.02)/0.13 0.04 (0.02)/0.13 0.04 (0.02)/0.13 0.04 (0.02)/0.13 0.04 (0.03)/0.13 0.04 (0.02)/0.13 0.04 (0.03)/0.14 0.04 (0.02)/0.13 0.04 (0.03)/0.14 0.04 (0.03)/0.14 0.84 0.82 0.83 0.81 0.81 0.77 0.8 0.77 0.8 0.77 0.77
Worsta
0.95 (1.07)/0.55 0.95 (1.07)/0.55 0.95 (1.08)/0.55 0.95 (1.08)/0.55 0.95 (1.04)/0.55 0.95 (1.02)/0.54 0.95 (1.04)/0.54 0.95 (1.02)/0.54 0.95 (1.04)/0.54 0.95 (1.02)/0.54 0.95 (1.03)/0.54 0.95 0.93 0.93 0.91 0.92 0.88 0.91 0.88 0.91 0.88 0.88
0°
0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.29)/0.62 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 0.93 0.93 0.91 0.92 0.88 0.91 0.88 0.91 0.88 0.88
Excluded Beamb 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.29)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.58 0.16 (0.28)/0.59 0.16 (0.28)/0.58 0.16 (0.28)/0.59 0.16 (0.28)/0.58 0.16 (0.28)/0.59 0.94 0.92 0.92 0.9 0.91 0.87 0.9 0.87 0.9 0.87 0.87
aLouvers bLouvers
0.34 (0.12)/0.47 0.34 (0.12)/0.47 0.34 (0.12)/0.47 0.34 (0.12)/0.47 0.33 (0.12)/0.46 0.33 (0.12)/0.46 0.33 (0.12)/0.46 0.33 (0.12)/0.46 0.33 (0.12)/0.46 0.33 (0.12)/0.46 0.33 (0.12)/0.46 0.94 0.92 0.92 0.9 0.91 0.88 0.91 0.87 0.91 0.87 0.88
Closed
0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.20 0.08 (0.04)/0.21 0.08 (0.04)/0.20 0.08 (0.04)/0.21 0.08 (0.04)/0.20 0.08 (0.04)/0.20 0.92 0.9 0.91 0.88 0.89 0.85 0.88 0.85 0.88 0.85 0.85
track so that profile angle equals negative slat angle and maximum direct beam is admitted. track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
cGlazing cavity width equals original cavity width plus slat width. dF is radiant fraction; ratio of radiative heat transfer to total heat transfer, R
on room side of glazing system.
15.43
Notes:
45°
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Louver Location Louver Reflection Indoor Side
0.15
21a
21b
21c
21d
21e
21f
21g
21h
21i
21j
21k
IAC 0 (IAC 60)/IACdiff , FR d
Worsta
0.99 (0.98)/0.94 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.85 0.82 0.82 0.8 0.8 0.75 0.79 0.75 0.79 0.75 0.76
0°
0.99 (0.92)/0.95 0.99 (0.92)/0.95 0.99 (0.93)/0.96 0.99 (0.93)/0.96 0.99 (0.93)/0.96 0.99 (0.93)/0.96 0.99 (0.93)/0.96 0.99 (0.94)/0.96 0.99 (0.93)/0.96 0.99 (0.94)/0.96 0.99 (0.93)/0.96 0.66 0.64 0.64 0.63 0.63 0.61 0.63 0.6 0.63 0.6 0.61
15.44
Table 14D IAC Values for Louvered Shades: Coated Double Glazings with 0.1 Low-e Glazing ID:
Excluded Beamb 0.89 (0.92)/0.95 0.89 (0.92)/0.95 0.90 (0.93)/0.95 0.90 (0.93)/0.96 0.90 (0.93)/0.95 0.91 (0.93)/0.96 0.90 (0.93)/0.96 0.91 (0.94)/0.96 0.90 (0.93)/0.96 0.91 (0.94)/0.96 0.91 (0.93)/0.96 0.41 0.41 0.41 0.41 0.4 0.4 0.4 0.4 0.4 0.4 0.4
Indoor Side
0.50
45°
0.92 (0.89)/0.93 0.92 (0.90)/0.93 0.92 (0.90)/0.93 0.93 (0.91)/0.94 0.92 (0.91)/0.94 0.93 (0.91)/0.94 0.93 (0.91)/0.94 0.93 (0.91)/0.94 0.93 (0.91)/0.94 0.93 (0.91)/0.94 0.93 (0.91)/0.94 0.44 0.43 0.43 0.43 0.43 0.42 0.42 0.41 0.42 0.41 0.42
Closed
0.86 (0.87)/0.88 0.87 (0.87)/0.89 0.87 (0.88)/0.89 0.88 (0.88)/0.90 0.88 (0.88)/0.89 0.89 (0.89)/0.90 0.88 (0.88)/0.90 0.89 (0.89)/0.90 0.88 (0.88)/0.90 0.89 (0.89)/0.90 0.89 (0.89)/0.90 0.4 0.39 0.39 0.39 0.38 0.37 0.38 0.37 0.38 0.37 0.37
Worsta
0.99 (0.98)/0.89 0.99 (0.98)/0.90 0.99 (0.98)/0.91 0.99 (0.98)/0.91 0.99 (0.98)/0.91 0.99 (0.98)/0.92 0.99 (0.98)/0.91 0.99 (0.98)/0.92 0.99 (0.98)/0.91 0.99 (0.98)/0.92 0.99 (0.98)/0.92 0.86 0.84 0.83 0.81 0.81 0.76 0.8 0.76 0.8 0.76 0.76
0°
0.99 (0.85)/0.92 0.99 (0.86)/0.92 0.99 (0.87)/0.93 0.99 (0.87)/0.93 0.99 (0.87)/0.93 0.99 (0.89)/0.94 0.99 (0.87)/0.93 0.99 (0.89)/0.94 0.99 (0.87)/0.93 0.99 (0.89)/0.94 0.99 (0.89)/0.94 0.7 0.68 0.67 0.66 0.66 0.63 0.66 0.63 0.66 0.63 0.64
Excluded Beamb 0.77 (0.85)/0.91 0.79 (0.86)/0.92 0.79 (0.87)/0.92 0.81 (0.87)/0.92 0.80 (0.87)/0.92 0.82 (0.89)/0.93 0.80 (0.87)/0.92 0.82 (0.89)/0.93 0.80 (0.87)/0.92 0.82 (0.89)/0.93 0.82 (0.89)/0.93 0.47 0.46 0.46 0.45 0.45 0.44 0.45 0.44 0.45 0.44 0.44
0.80
0.83 (0.77)/0.86 0.84 (0.79)/0.87 0.85 (0.79)/0.88 0.86 (0.81)/0.88 0.85 (0.80)/0.88 0.87 (0.82)/0.89 0.86 (0.80)/0.88 0.87 (0.83)/0.89 0.86 (0.80)/0.88 0.87 (0.83)/0.89 0.87 (0.82)/0.89 0.49 0.48 0.48 0.47 0.47 0.45 0.47 0.45 0.47 0.45 0.45
Closed
0.71 (0.70)/0.75 0.73 (0.73)/0.77 0.74 (0.73)/0.78 0.75 (0.75)/0.79 0.75 (0.75)/0.79 0.77 (0.77)/0.81 0.75 (0.75)/0.79 0.78 (0.78)/0.81 0.75 (0.75)/0.79 0.78 (0.78)/0.81 0.77 (0.77)/0.81 0.41 0.39 0.39 0.38 0.39 0.37 0.38 0.37 0.38 0.37 0.37
Worsta
0.98 (0.97)/0.84 0.98 (0.97)/0.85 0.98 (0.97)/0.86 0.98 (0.97)/0.87 0.98 (0.97)/0.86 0.98 (0.97)/0.88 0.98 (0.97)/0.86 0.98 (0.97)/0.88 0.98 (0.97)/0.86 0.98 (0.97)/0.88 0.98 (0.97)/0.88 0.88 0.85 0.84 0.82 0.82 0.77 0.81 0.77 0.81 0.77 0.78
0°
0.98 (0.76)/0.87 0.98 (0.78)/0.88 0.98 (0.79)/0.89 0.98 (0.81)/0.89 0.98 (0.80)/0.89 0.98 (0.83)/0.90 0.98 (0.81)/0.89 0.98 (0.83)/0.90 0.98 (0.81)/0.89 0.98 (0.83)/0.90 0.98 (0.83)/0.90 0.76 0.74 0.73 0.71 0.72 0.68 0.71 0.68 0.71 0.68 0.68
Excluded Beamb 0.64 (0.76)/0.86 0.67 (0.78)/0.87 0.69 (0.79)/0.88 0.71 (0.81)/0.89 0.70 (0.80)/0.88 0.73 (0.83)/0.90 0.70 (0.81)/0.88 0.74 (0.83)/0.90 0.70 (0.81)/0.88 0.74 (0.83)/0.90 0.73 (0.83)/0.90 0.59 0.56 0.56 0.54 0.55 0.52 0.54 0.52 0.54 0.52 0.52
Between Glazingsc
0.15
45°
0.74 (0.64)/0.79 0.76 (0.67)/0.81 0.77 (0.68)/0.81 0.79 (0.71)/0.83 0.78 (0.70)/0.82 0.80 (0.74)/0.84 0.78 (0.70)/0.82 0.81 (0.74)/0.84 0.78 (0.70)/0.82 0.81 (0.74)/0.84 0.81 (0.74)/0.84 0.59 0.57 0.56 0.54 0.55 0.52 0.55 0.52 0.55 0.51 0.52
Closed
0.54 (0.53)/0.62 0.59 (0.58)/0.66 0.60 (0.59)/0.67 0.63 (0.62)/0.69 0.62 (0.61)/0.68 0.66 (0.66)/0.72 0.62 (0.62)/0.69 0.67 (0.66)/0.72 0.62 (0.62)/0.69 0.67 (0.67)/0.72 0.66 (0.66)/0.72 0.45 0.42 0.42 0.4 0.41 0.38 0.41 0.38 0.41 0.38 0.38
Worsta
0.98 (1.01)/0.80 0.98 (1.01)/0.80 0.96 (0.99)/0.60 0.96 (0.99)/0.61 0.96 (0.98)/0.61 0.95 (0.97)/0.63 0.95 (0.98)/0.61 0.95 (0.97)/0.63 0.95 (0.98)/0.61 0.95 (0.97)/0.63 0.95 (0.97)/0.62 0.91 0.9 0.9 0.89 0.89 0.87 0.89 0.86 0.89 0.86 0.87
0°
0.98 (0.69)/0.82 0.98 (0.70)/0.83 0.96 (0.40)/0.64 0.96 (0.40)/0.65 0.96 (0.42)/0.65 0.95 (0.44)/0.66 0.95 (0.42)/0.65 0.95 (0.45)/0.66 0.95 (0.42)/0.65 0.95 (0.45)/0.66 0.95 (0.44)/0.66 0.8 0.79 0.79 0.78 0.78 0.77 0.78 0.77 0.78 0.77 0.77
Excluded Beamb 0.63 (0.69)/0.81 0.64 (0.70)/0.82 0.34 (0.40)/0.63 0.35 (0.40)/0.63 0.36 (0.42)/0.64 0.39 (0.44)/0.65 0.37 (0.42)/0.64 0.39 (0.45)/0.65 0.37 (0.42)/0.64 0.40 (0.45)/0.65 0.39 (0.44)/0.65 0.66 0.66 0.66 0.65 0.65 0.65 0.65 0.65 0.65 0.65 0.65
Between Glazingsc
0.50
45°
0.71 (0.66)/0.77 0.71 (0.67)/0.78 0.47 (0.36)/0.55 0.47 (0.37)/0.56 0.48 (0.38)/0.56 0.50 (0.41)/0.58 0.49 (0.39)/0.57 0.51 (0.42)/0.58 0.49 (0.39)/0.57 0.51 (0.42)/0.58 0.50 (0.41)/0.58 0.69 0.69 0.7 0.7 0.69 0.69 0.69 0.68 0.69 0.68 0.69
Closed
0.61 (0.64)/0.68 0.63 (0.65)/0.69 0.33 (0.35)/0.41 0.34 (0.36)/0.41 0.35 (0.37)/0.42 0.38 (0.40)/0.45 0.35 (0.37)/0.43 0.38 (0.40)/0.45 0.35 (0.37)/0.43 0.38 (0.40)/0.45 0.38 (0.40)/0.45 0.65 0.65 0.64 0.64 0.64 0.64 0.64 0.64 0.64 0.64 0.64
Worsta
0.97 (1.03)/0.77 0.97 (1.03)/0.78 0.97 (1.06)/0.63 0.97 (1.06)/0.63 0.96 (1.05)/0.64 0.95 (1.03)/0.65 0.96 (1.04)/0.64 0.95 (1.03)/0.65 0.96 (1.04)/0.64 0.95 (1.03)/0.65 0.95 (1.03)/0.65 0.92 0.9 0.91 0.89 0.9 0.87 0.89 0.87 0.89 0.87 0.87
0°
0.97 (0.65)/0.81 0.97 (0.66)/0.81 0.97 (0.42)/0.68 0.97 (0.42)/0.68 0.96 (0.44)/0.69 0.95 (0.46)/0.69 0.96 (0.44)/0.69 0.95 (0.46)/0.70 0.96 (0.44)/0.69 0.95 (0.47)/0.70 0.95 (0.46)/0.69 0.83 0.82 0.83 0.82 0.82 0.8 0.81 0.8 0.81 0.79 0.8
Excluded Beamb 0.54 (0.65)/0.80 0.55 (0.66)/0.80 0.32 (0.42)/0.66 0.32 (0.42)/0.67 0.34 (0.44)/0.67 0.37 (0.46)/0.68 0.34 (0.44)/0.67 0.37 (0.46)/0.68 0.34 (0.44)/0.67 0.38 (0.47)/0.68 0.37 (0.46)/0.68 0.7 0.7 0.72 0.71 0.71 0.7 0.71 0.7 0.71 0.7 0.7
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Indoor Side
45°
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Between Glazingsc
0.80
21a
21b
21c
21d
21e
21f
21g
21h
21i
21j
21k
45°
0.64 (0.54)/0.72 0.65 (0.55)/0.73 0.47 (0.31)/0.56 0.47 (0.32)/0.57 0.48 (0.34)/0.58 0.50 (0.37)/0.59 0.49 (0.34)/0.58 0.51 (0.37)/0.59 0.49 (0.34)/0.58 0.51 (0.38)/0.59 0.50 (0.37)/0.59 0.73 0.72 0.75 0.74 0.74 0.72 0.73 0.72 0.73 0.72 0.72
Closed
0.47 (0.47)/0.56 0.48 (0.48)/0.57 0.26 (0.25)/0.36 0.27 (0.26)/0.37 0.28 (0.28)/0.38 0.32 (0.32)/0.41 0.29 (0.29)/0.39 0.32 (0.32)/0.41 0.29 (0.29)/0.39 0.33 (0.32)/0.42 0.32 (0.32)/0.41 0.66 0.66 0.67 0.66 0.66 0.65 0.66 0.65 0.66 0.65 0.66
Worsta
0.97 (1.04)/0.74 0.97 (1.04)/0.75 0.97 (1.12)/0.67 0.97 (1.12)/0.68 0.96 (1.11)/0.68 0.96 (1.09)/0.69 0.96 (1.11)/0.68 0.96 (1.09)/0.69 0.96 (1.11)/0.68 0.95 (1.09)/0.69 0.96 (1.09)/0.69 0.93 0.92 0.91 0.9 0.9 0.88 0.9 0.88 0.9 0.88 0.88
0°
0.97 (0.60)/0.79 0.97 (0.60)/0.80 0.97 (0.46)/0.73 0.97 (0.46)/0.73 0.96 (0.47)/0.73 0.96 (0.50)/0.74 0.96 (0.48)/0.73 0.96 (0.50)/0.74 0.96 (0.48)/0.73 0.95 (0.50)/0.74 0.96 (0.50)/0.74 0.88 0.86 0.87 0.86 0.86 0.84 0.86 0.83 0.86 0.83 0.84
Fenestration
Table 14D IAC Values for Louvered Shades: Coated Double Glazings with 0.1 Low-e (Continued) Glazing ID:
Excluded Beamb 0.45 (0.60)/0.78 0.46 (0.60)/0.78 0.33 (0.46)/0.71 0.33 (0.46)/0.71 0.35 (0.47)/0.71 0.38 (0.50)/0.72 0.35 (0.48)/0.72 0.38 (0.50)/0.72 0.35 (0.48)/0.72 0.39 (0.50)/0.72 0.38 (0.50)/0.72 0.8 0.79 0.82 0.81 0.8 0.77 0.79 0.77 0.79 0.77 0.77
Licensed for single user. © 2017 ASHRAE, Inc.
Outdoor Side
0.15
45°
0.59 (0.43)/0.68 0.59 (0.44)/0.68 0.49 (0.29)/0.60 0.50 (0.30)/0.60 0.51 (0.31)/0.61 0.53 (0.35)/0.62 0.51 (0.32)/0.61 0.53 (0.35)/0.63 0.51 (0.32)/0.61 0.53 (0.36)/0.63 0.53 (0.35)/0.62 0.81 0.8 0.82 0.81 0.8 0.78 0.8 0.77 0.8 0.77 0.78
Closed
0.33 (0.30)/0.44 0.34 (0.31)/0.45 0.22 (0.17)/0.34 0.23 (0.18)/0.35 0.25 (0.21)/0.37 0.28 (0.25)/0.40 0.25 (0.22)/0.37 0.29 (0.26)/0.40 0.25 (0.22)/0.37 0.29 (0.26)/0.41 0.28 (0.25)/0.40 0.73 0.72 0.75 0.74 0.72 0.7 0.72 0.7 0.72 0.7 0.7
Worsta
0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.36 0.93 (0.89)/0.35 0.93 (0.89)/0.36 0.93 (0.89)/0.35 0.94 0.93 0.92 0.91 0.91 0.88 0.9 0.88 0.9 0.88 0.88
0°
0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.05)/0.41 0.9 0.88 0.88 0.87 0.87 0.84 0.86 0.84 0.86 0.83 0.84
Excluded Beamb 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.03 (0.05)/0.39 0.02 (0.04)/0.39 0.03 (0.05)/0.39 0.02 (0.04)/0.39 0.03 (0.05)/0.39 0.03 (0.05)/0.39 0.8 0.79 0.79 0.78 0.77 0.74 0.77 0.74 0.77 0.74 0.74
Outdoor Side
0.50
45°
0.19 (0.02)/0.29 0.19 (0.02)/0.29 0.19 (0.02)/0.29 0.19 (0.02)/0.29 0.19 (0.02)/0.29 0.20 (0.03)/0.29 0.19 (0.03)/0.29 0.20 (0.03)/0.29 0.19 (0.03)/0.29 0.20 (0.03)/0.29 0.20 (0.03)/0.29 0.84 0.83 0.83 0.82 0.81 0.79 0.81 0.78 0.81 0.78 0.79
Closed
0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.10 0.02 (0.02)/0.09 0.02 (0.03)/0.10 0.02 (0.02)/0.09 0.02 (0.03)/0.10 0.02 (0.02)/0.10 0.67 0.66 0.67 0.66 0.66 0.65 0.66 0.65 0.66 0.65 0.65
Worsta
0.94 (1.02)/0.44 0.94 (1.01)/0.44 0.94 (1.01)/0.44 0.94 (1.01)/0.44 0.94 (0.98)/0.44 0.94 (0.96)/0.44 0.94 (0.98)/0.44 0.94 (0.96)/0.44 0.94 (0.98)/0.44 0.94 (0.96)/0.44 0.94 (0.97)/0.44 0.94 0.93 0.92 0.91 0.91 0.88 0.91 0.88 0.91 0.88 0.88
0°
0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 0.92 0.92 0.9 0.9 0.87 0.9 0.87 0.9 0.87 0.88
Excluded Beamb 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.08 (0.14)/0.48 0.07 (0.14)/0.48 0.08 (0.14)/0.48 0.07 (0.14)/0.48 0.08 (0.14)/0.48 0.08 (0.14)/0.48 0.91 0.9 0.89 0.88 0.88 0.85 0.87 0.84 0.87 0.84 0.85
Outdoor Side
0.80
45°
0.25 (0.06)/0.37 0.25 (0.06)/0.37 0.25 (0.06)/0.37 0.25 (0.06)/0.37 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.25 (0.06)/0.36 0.92 0.9 0.9 0.89 0.88 0.85 0.88 0.85 0.88 0.85 0.85
Closed
0.04 (0.02)/0.14 0.04 (0.02)/0.14 0.04 (0.02)/0.14 0.04 (0.02)/0.14 0.04 (0.02)/0.14 0.04 (0.03)/0.14 0.04 (0.02)/0.14 0.04 (0.03)/0.14 0.04 (0.02)/0.14 0.04 (0.03)/0.14 0.04 (0.03)/0.14 0.84 0.82 0.83 0.81 0.8 0.77 0.8 0.77 0.8 0.77 0.77
Worsta
0.95 (1.14)/0.56 0.95 (1.12)/0.56 0.95 (1.13)/0.56 0.95 (1.12)/0.56 0.95 (1.07)/0.55 0.95 (1.04)/0.55 0.95 (1.07)/0.55 0.95 (1.04)/0.54 0.95 (1.07)/0.55 0.95 (1.03)/0.54 0.95 (1.05)/0.55 0.94 0.93 0.92 0.91 0.91 0.88 0.91 0.88 0.91 0.88 0.88
0°
0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 (0.28)/0.61 0.95 0.93 0.92 0.91 0.91 0.88 0.91 0.88 0.91 0.88 0.88
Excluded Beamb 0.17 (0.29)/0.60 0.17 (0.29)/0.60 0.17 (0.29)/0.60 0.17 (0.29)/0.60 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.16 (0.28)/0.59 0.94 0.92 0.91 0.9 0.9 0.87 0.9 0.87 0.9 0.87 0.87
aLouvers bLouvers
0.35 (0.13)/0.49 0.35 (0.13)/0.48 0.35 (0.13)/0.48 0.35 (0.13)/0.48 0.34 (0.12)/0.47 0.34 (0.12)/0.46 0.34 (0.12)/0.47 0.34 (0.12)/0.46 0.34 (0.12)/0.47 0.34 (0.12)/0.46 0.34 (0.12)/0.47 0.94 0.92 0.92 0.9 0.9 0.88 0.9 0.87 0.9 0.87 0.88
Closed
0.08 (0.04)/0.22 0.08 (0.04)/0.22 0.08 (0.04)/0.22 0.08 (0.04)/0.22 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.92 0.9 0.9 0.89 0.88 0.85 0.88 0.85 0.88 0.85 0.85
track so that profile angle equals negative slat angle and maximum direct beam is admitted. track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
cGlazing dF R
cavity width equals original cavity width plus slat width. is radiant fraction; ratio of radiative heat transfer to total heat transfer, on room side of glazing system.
15.45
Notes:
45°
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Louver Location Louver Reflection Indoor Side
Indoor Side
Between Glazingsc
Between Glazingsc
0.50
0.80
0.15
0.50
25b
25c
26d
25e
25f
IAC 0 (IAC 60)/IACdiff , FR d
Worsta
0.99 (0.99)/0.95 0.84
0.99 (0.98)/0.95 0.81
0.99 (0.98)/0.96 0.74
0.99 (0.98)/0.96 0.76
0.99 (0.98)/0.96 0.72
0.99 (0.98)/0.96 0.76
0°
0.99 (0.93)/0.96 0.65
0.99 (0.93)/0.96 0.64
0.99 (0.94)/0.96 0.6
0.99 (0.94)/0.96 0.61
0.99 (0.95)/0.96 0.59
0.99 (0.94)/0.96 0.61
Excluded Beamb
0.90 (0.93)/0.96 0.41
0.91 (0.93)/0.96 0.4
0.92 (0.94)/0.96 0.39
0.92 (0.94)/0.96 0.4
0.92 (0.95)/0.96 0.39
0.92 (0.94)/0.96 0.4
45°
0.93 (0.91)/0.94 0.43
0.93 (0.91)/0.94 0.43
0.94 (0.93)/0.95 0.41
0.94 (0.92)/0.95 0.42
0.94 (0.93)/0.95 0.41
0.94 (0.92)/0.95 0.42
Closed
0.88 (0.88)/0.90 0.39
0.89 (0.89)/0.90 0.39
0.90 (0.90)/0.91 0.36
0.90 (0.90)/0.91 0.37
0.90 (0.91)/0.92 0.36
0.90 (0.90)/0.91 0.37
Worsta
0.99 (0.98)/0.91 0.86
0.99 (0.98)/0.92 0.82
0.99 (0.98)/0.93 0.75
0.99 (0.98)/0.93 0.77
0.99 (0.98)/0.93 0.73
0.99 (0.98)/0.93 0.77
0°
0.99 (0.87)/0.93 0.69
0.99 (0.89)/0.94 0.67
0.99 (0.91)/0.95 0.63
0.99 (0.90)/0.94 0.64
0.99 (0.91)/0.95 0.62
0.99 (0.90)/0.94 0.64
Excluded Beamb
0.80 (0.87)/0.92 0.47
0.82 (0.89)/0.93 0.46
0.85 (0.91)/0.94 0.44
0.85 (0.90)/0.94 0.44
0.86 (0.91)/0.94 0.43
0.84 (0.90)/0.94 0.45
45°
0.85 (0.80)/0.88 0.49
0.87 (0.83)/0.90 0.48
0.89 (0.85)/0.91 0.45
0.89 (0.85)/0.91 0.46
0.89 (0.86)/0.91 0.45
0.89 (0.85)/0.91 0.46
Closed
0.74 (0.74)/0.79 0.4
0.77 (0.77)/0.81 0.39
0.81 (0.81)/0.84 0.37
0.80 (0.80)/0.83 0.37
0.82 (0.82)/0.85 0.36
0.80 (0.80)/0.83 0.38
Worsta
0.98 (0.98)/0.86 0.87
0.98 (0.98)/0.88 0.84
0.98 (0.97)/0.90 0.76
0.98 (0.98)/0.90 0.78
0.98 (0.97)/0.90 0.74
0.98 (0.98)/0.89 0.78
0°
0.98 (0.80)/0.89 0.76
0.98 (0.83)/0.91 0.73
0.98 (0.86)/0.92 0.68
0.98 (0.85)/0.92 0.69
0.98 (0.86)/0.92 0.66
0.98 (0.85)/0.92 0.69
Excluded Beamb
0.69 (0.80)/0.88 0.59
0.73 (0.83)/0.90 0.56
0.78 (0.86)/0.91 0.52
0.77 (0.85)/0.91 0.53
0.78 (0.86)/0.92 0.51
0.77 (0.85)/0.91 0.53
45°
0.78 (0.69)/0.82 0.59
0.81 (0.73)/0.84 0.57
0.84 (0.78)/0.87 0.52
0.83 (0.77)/0.86 0.53
0.84 (0.79)/0.87 0.51
0.83 (0.77)/0.86 0.54
Closed
0.60 (0.59)/0.67 0.45
0.65 (0.65)/0.71 0.42
0.71 (0.71)/0.76 0.39
0.70 (0.70)/0.75 0.4
0.72 (0.72)/0.77 0.38
0.70 (0.69)/0.75 0.4
Worsta
0.97 (1.00)/0.80 0.91
0.97 (0.99)/0.81 0.89
0.95 (0.96)/0.80 0.86
0.95 (0.96)/0.80 0.86
0.94 (0.95)/0.80 0.85
0.96 (0.96)/0.80 0.86
0°
0.97 (0.71)/0.82 0.79
0.97 (0.72)/0.82 0.78
0.95 (0.73)/0.82 0.76
0.95 (0.73)/0.82 0.77
0.94 (0.73)/0.82 0.76
0.96 (0.73)/0.82 0.77
Excluded Beamb
0.66 (0.71)/0.82 0.65
0.67 (0.72)/0.82 0.65
0.69 (0.73)/0.81 0.65
0.69 (0.73)/0.81 0.65
0.70 (0.73)/0.81 0.64
0.69 (0.73)/0.81 0.65
45°
0.73 (0.69)/0.78 0.68
0.73 (0.69)/0.78 0.68
0.75 (0.71)/0.78 0.67
0.75 (0.71)/0.78 0.67
0.75 (0.72)/0.79 0.67
0.75 (0.71)/0.78 0.67
Closed
0.65 (0.67)/0.70 0.65
0.66 (0.68)/0.71 0.65
0.68 (0.70)/0.72 0.64
0.68 (0.70)/0.72 0.64
0.69 (0.71)/0.73 0.64
0.68 (0.70)/0.72 0.64
Worsta
0.97 (1.01)/0.79 0.92
0.97 (1.00)/0.79 0.9
0.95 (0.97)/0.80 0.86
0.95 (0.98)/0.80 0.87
0.94 (0.96)/0.80 0.85
0.95 (0.98)/0.80 0.87
0°
0.97 (0.68)/0.82 0.82
0.97 (0.69)/0.82 0.81
0.95 (0.71)/0.82 0.79
0.95 (0.71)/0.82 0.79
0.94 (0.72)/0.82 0.78
0.95 (0.71)/0.82 0.79
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Indoor Side
0.15
25a
15.46
Table 14E IAC Values for Louvered Shades: Double Glazings with 0.05 Low-e Glazing ID:
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Between Glazingsc
Licensed for single user. © 2017 ASHRAE, Inc.
Outdoor Side
Outdoor Side
Outdoor Side
aLouvers bLouvers
0.15
0.50
0.80
25b
25c
26d
25e
25f
0.58 (0.68)/0.81 0.7
0.60 (0.69)/0.81 0.7
0.64 (0.71)/0.81 0.68
0.63 (0.71)/0.81 0.69
0.65 (0.72)/0.81 0.68
0.63 (0.71)/0.81 0.69
45°
0.68 (0.59)/0.74 0.73
0.69 (0.61)/0.75 0.72
0.71 (0.65)/0.76 0.7
0.71 (0.64)/0.76 0.71
0.72 (0.66)/0.77 0.7
0.71 (0.64)/0.76 0.71
Closed
0.52 (0.52)/0.60 0.66
0.54 (0.54)/0.61 0.66
0.59 (0.59)/0.65 0.65
0.58 (0.58)/0.64 0.66
0.60 (0.60)/0.65 0.65
0.58 (0.58)/0.64 0.66
Worsta
0.97 (1.02)/0.77 0.93
0.97 (1.01)/0.78 0.91
0.95 (0.98)/0.79 0.87
0.95 (0.98)/0.79 0.88
0.94 (0.97)/0.79 0.86
0.95 (0.99)/0.79 0.88
0°
0.97 (0.65)/0.81 0.87
0.97 (0.67)/0.82 0.86
0.95 (0.70)/0.82 0.83
0.95 (0.69)/0.82 0.83
0.94 (0.71)/0.82 0.82
0.95 (0.69)/0.82 0.84
Excluded Beamb
0.51 (0.65)/0.80 0.8
0.53 (0.67)/0.81 0.79
0.59 (0.70)/0.81 0.76
0.58 (0.69)/0.81 0.76
0.60 (0.71)/0.81 0.75
0.58 (0.69)/0.81 0.76
45°
0.64 (0.50)/0.71 0.81
0.65 (0.53)/0.73 0.8
0.69 (0.59)/0.74 0.76
0.68 (0.58)/0.74 0.77
0.69 (0.60)/0.75 0.76
0.68 (0.58)/0.74 0.77
Closed
0.39 (0.36)/0.49 0.73
0.42 (0.39)/0.51 0.73
0.49 (0.47)/0.57 0.7
0.48 (0.46)/0.56 0.71
0.51 (0.49)/0.58 0.7
0.48 (0.46)/0.56 0.71
Worsta
0.93 (0.92)/0.36 0.94
0.93 (0.92)/0.36 0.92
0.93 (0.90)/0.36 0.87
0.93 (0.89)/0.36 0.88
0.93 (0.90)/0.36 0.86
0.93 (0.89)/0.36 0.89
0°
0.93 (0.05)/0.41 0.89
0.93 (0.05)/0.41 0.87
0.93 (0.06)/0.41 0.82
0.93 (0.05)/0.41 0.83
0.93 (0.06)/0.42 0.81
0.93 (0.05)/0.41 0.84
Excluded Beamb
0.03 (0.05)/0.39 0.78
0.03 (0.05)/0.39 0.77
0.04 (0.06)/0.40 0.72
0.03 (0.05)/0.39 0.73
0.04 (0.06)/0.40 0.71
0.03 (0.05)/0.39 0.74
45°
0.20 (0.03)/0.29 0.83
0.20 (0.03)/0.29 0.82
0.20 (0.04)/0.30 0.77
0.20 (0.03)/0.29 0.78
0.20 (0.04)/0.30 0.76
0.20 (0.03)/0.29 0.78
Closed
0.02 (0.02)/0.10 0.66
0.02 (0.02)/0.10 0.66
0.03 (0.03)/0.11 0.64
0.03 (0.03)/0.10 0.65
0.03 (0.04)/0.11 0.64
0.03 (0.03)/0.10 0.65
Worsta
0.94 (1.08)/0.45 0.94
0.94 (1.06)/0.45 0.92
0.94 (0.99)/0.44 0.87
0.94 (0.96)/0.44 0.88
0.94 (0.98)/0.44 0.86
0.94 (0.97)/0.44 0.89
0°
0.94 (0.15)/0.51 0.93
0.94 (0.15)/0.51 0.91
0.94 (0.15)/0.50 0.86
0.94 (0.15)/0.50 0.87
0.94 (0.15)/0.50 0.85
0.94 (0.15)/0.50 0.88
Excluded Beamb
0.08 (0.15)/0.49 0.9
0.08 (0.15)/0.49 0.88
0.08 (0.15)/0.48 0.83
0.08 (0.15)/0.48 0.84
0.08 (0.15)/0.48 0.82
0.08 (0.15)/0.48 0.85
45°
0.26 (0.06)/0.38 0.91
0.26 (0.06)/0.38 0.89
0.26 (0.07)/0.37 0.84
0.25 (0.06)/0.36 0.85
0.26 (0.07)/0.37 0.82
0.25 (0.06)/0.36 0.85
Closed
0.04 (0.03)/0.15 0.82
0.04 (0.03)/0.15 0.81
0.05 (0.03)/0.14 0.75
0.04 (0.03)/0.14 0.76
0.05 (0.03)/0.14 0.74
0.04 (0.03)/0.14 0.76
Worsta
0.95 (1.25)/0.59 0.94
0.95 (1.21)/0.58 0.92
0.95 (1.08)/0.56 0.87
0.95 (1.04)/0.55 0.88
0.95 (1.06)/0.55 0.86
0.95 (1.05)/0.55 0.89
0°
0.95 (0.30)/0.64 0.94
0.95 (0.30)/0.64 0.92
0.95 (0.29)/0.62 0.88
0.95 (0.28)/0.61 0.88
0.95 (0.29)/0.61 0.86
0.95 (0.28)/0.61 0.89
Excluded Beamb
0.18 (0.30)/0.62 0.93
0.18 (0.30)/0.62 0.91
0.17 (0.29)/0.60 0.86
0.16 (0.28)/0.59 0.87
0.17 (0.29)/0.59 0.85
0.17 (0.28)/0.59 0.88
45°
0.37 (0.13)/0.51 0.93
0.37 (0.13)/0.50 0.91
0.35 (0.13)/0.48 0.86
0.34 (0.12)/0.47 0.87
0.35 (0.13)/0.47 0.85
0.34 (0.12)/0.47 0.88
Closed
0.09 (0.04)/0.24 0.91
0.09 (0.04)/0.24 0.89
0.09 (0.04)/0.22 0.83
0.08 (0.04)/0.21 0.85
0.09 (0.04)/0.21 0.82
0.08 (0.04)/0.21 0.85
track so that profile angle equals negative slat angle and maximum direct beam is admitted. track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
cGlazing dF R
cavity width equals original cavity width plus slat width. is radiant fraction; ratio of radiative heat transfer to total heat transfer, on room side of glazing system.
15.47
Notes:
0.80
25a
Excluded Beamb
Fenestration
Table 14E IAC Values for Louvered Shades: Double Glazings with 0.05 Low-e (Continued) Glazing ID:
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Louver Location Louver Reflection Indoor Side
0.15
29a
29b
32a
32b
32c
32d
40a
40b
40c
40d
IAC 0 (IAC 60)/IACdiff , FR d
Worsta
0.99 (0.98)/0.94 0.99 (0.98)/0.95 0.99 (0.98)/0.95 0.99 (0.98)/0.96 1.00 (1.00)/0.97 1.00 (1.00)/0.97 1.00 (1.00)/0.98 1.00 (1.00)/0.98 0.99 (0.99)/0.96 0.99 (0.99)/0.96 0.82 0.78 0.8 0.75 0.76 0.71 0.73 0.67 0.78 0.73
0°
0.99 (0.92)/0.95 0.99 (0.93)/0.95 0.99 (0.94)/0.96 0.99 (0.94)/0.96 1.00 (0.96)/0.98 1.00 (0.96)/0.98 1.00 (0.96)/0.98 1.00 (0.97)/0.99 0.99 (0.95)/0.97 0.99 (0.95)/0.97 0.65 0.62 0.64 0.61 0.6 0.56 0.58 0.54 0.63 0.6
15.48
Table 14F IAC Values for Louvered Shades: Triple Glazing Glazing ID:
Excluded Beamb 0.88 (0.92)/0.95 0.89 (0.93)/0.95 0.90 (0.94)/0.96 0.91 (0.94)/0.96 0.93 (0.96)/0.97 0.93 (0.96)/0.98 0.94 (0.96)/0.98 0.95 (0.97)/0.98 0.92 (0.95)/0.97 0.93 (0.95)/0.97 0.41 0.4 0.41 0.4 0.39 0.36 0.37 0.35 0.4 0.4
Indoor Side
0.50
45°
0.91 (0.89)/0.93 0.92 (0.90)/0.93 0.93 (0.91)/0.94 0.93 (0.92)/0.95 0.95 (0.93)/0.96 0.95 (0.94)/0.97 0.95 (0.94)/0.97 0.96 (0.95)/0.97 0.94 (0.92)/0.95 0.94 (0.93)/0.96 0.43 0.43 0.43 0.42 0.41 0.38 0.39 0.37 0.42 0.41
Closed
0.86 (0.87)/0.88 0.87 (0.88)/0.89 0.88 (0.89)/0.90 0.89 (0.90)/0.91 0.91 (0.91)/0.93 0.92 (0.92)/0.94 0.92 (0.93)/0.94 0.93 (0.94)/0.95 0.90 (0.90)/0.92 0.91 (0.91)/0.92 0.39 0.38 0.39 0.37 0.37 0.35 0.36 0.33 0.38 0.37
Worsta
0.98 (0.98)/0.90 0.98 (0.98)/0.91 0.99 (0.98)/0.91 0.99 (0.98)/0.92 0.99 (0.99)/0.92 1.00 (1.00)/0.93 0.99 (1.00)/0.93 1.00 (1.00)/0.95 0.99 (0.98)/0.92 0.99 (0.98)/0.93 0.83 0.79 0.81 0.76 0.77 0.72 0.74 0.68 0.79 0.74
0°
0.98 (0.86)/0.92 0.98 (0.87)/0.93 0.99 (0.88)/0.93 0.99 (0.90)/0.94 0.99 (0.89)/0.94 1.00 (0.91)/0.95 0.99 (0.91)/0.95 1.00 (0.92)/0.96 0.99 (0.89)/0.94 0.99 (0.91)/0.95 0.68 0.65 0.67 0.64 0.62 0.58 0.6 0.56 0.66 0.63
Excluded Beamb 0.77 (0.86)/0.91 0.80 (0.87)/0.92 0.80 (0.88)/0.93 0.83 (0.90)/0.94 0.81 (0.89)/0.94 0.83 (0.91)/0.95 0.83 (0.91)/0.95 0.86 (0.92)/0.96 0.82 (0.89)/0.93 0.84 (0.91)/0.94 0.46 0.45 0.45 0.44 0.42 0.39 0.4 0.37 0.45 0.44
0.80
0.83 (0.78)/0.87 0.85 (0.81)/0.88 0.85 (0.81)/0.89 0.87 (0.84)/0.90 0.86 (0.82)/0.90 0.88 (0.84)/0.91 0.88 (0.84)/0.91 0.90 (0.87)/0.93 0.87 (0.83)/0.90 0.89 (0.85)/0.91 0.48 0.46 0.47 0.45 0.43 0.4 0.41 0.38 0.46 0.45
Closed
0.71 (0.72)/0.77 0.74 (0.75)/0.79 0.75 (0.76)/0.80 0.78 (0.79)/0.82 0.76 (0.76)/0.81 0.79 (0.80)/0.83 0.78 (0.80)/0.83 0.82 (0.83)/0.86 0.77 (0.78)/0.81 0.80 (0.81)/0.84 0.39 0.38 0.39 0.37 0.35 0.33 0.34 0.31 0.38 0.36
Worsta
0.98 (0.97)/0.85 0.98 (0.97)/0.87 0.98 (0.97)/0.86 0.98 (0.97)/0.88 0.99 (0.99)/0.87 0.99 (0.99)/0.89 0.99 (0.99)/0.88 0.99 (0.99)/0.90 0.98 (0.98)/0.87 0.98 (0.98)/0.89 0.84 0.8 0.82 0.77 0.79 0.73 0.75 0.69 0.8 0.75
0°
0.98 (0.78)/0.88 0.98 (0.81)/0.89 0.98 (0.81)/0.89 0.98 (0.84)/0.91 0.99 (0.81)/0.90 0.99 (0.83)/0.91 0.99 (0.83)/0.91 0.99 (0.86)/0.93 0.98 (0.82)/0.90 0.98 (0.85)/0.92 0.74 0.7 0.72 0.68 0.67 0.62 0.64 0.59 0.71 0.67
Excluded Beamb 0.66 (0.78)/0.87 0.70 (0.81)/0.89 0.69 (0.81)/0.88 0.74 (0.84)/0.90 0.69 (0.81)/0.89 0.73 (0.83)/0.90 0.72 (0.83)/0.90 0.77 (0.86)/0.92 0.71 (0.82)/0.89 0.76 (0.85)/0.91 0.56 0.53 0.55 0.52 0.49 0.45 0.46 0.42 0.54 0.51
Between Glazingsc
0.15
45°
0.75 (0.67)/0.80 0.78 (0.71)/0.83 0.77 (0.71)/0.82 0.81 (0.75)/0.85 0.77 (0.70)/0.83 0.80 (0.74)/0.85 0.80 (0.74)/0.85 0.83 (0.78)/0.87 0.79 (0.73)/0.84 0.82 (0.77)/0.86 0.56 0.53 0.55 0.52 0.49 0.44 0.46 0.42 0.54 0.51
Closed
0.57 (0.58)/0.65 0.62 (0.64)/0.70 0.61 (0.63)/0.69 0.67 (0.68)/0.74 0.60 (0.62)/0.69 0.66 (0.67)/0.73 0.65 (0.66)/0.72 0.71 (0.72)/0.77 0.64 (0.65)/0.71 0.69 (0.71)/0.76 0.42 0.4 0.41 0.38 0.35 0.31 0.33 0.29 0.4 0.38
Worsta
0.97 (1.01)/0.63 0.97 (1.02)/0.64 0.98 (1.05)/0.76 0.98 (1.05)/0.77 0.97 (1.01)/0.60 0.98 (1.02)/0.62 0.98 (1.05)/0.73 0.99 (1.05)/0.75 0.98 (1.04)/0.71 0.99 (1.04)/0.73 0.9 0.87 0.88 0.85 0.79 0.74 0.75 0.7 0.87 0.84
0°
0.97 (0.44)/0.66 0.97 (0.46)/0.68 0.98 (0.62)/0.78 0.98 (0.63)/0.80 0.97 (0.40)/0.64 0.98 (0.42)/0.66 0.98 (0.57)/0.76 0.99 (0.60)/0.78 0.98 (0.55)/0.74 0.99 (0.58)/0.76 0.79 0.78 0.78 0.77 0.59 0.56 0.57 0.54 0.78 0.76
Excluded Beamb 0.36 (0.44)/0.65 0.38 (0.46)/0.67 0.52 (0.62)/0.77 0.54 (0.63)/0.79 0.32 (0.40)/0.63 0.35 (0.42)/0.65 0.48 (0.57)/0.75 0.50 (0.60)/0.77 0.46 (0.55)/0.73 0.49 (0.58)/0.75 0.66 0.65 0.66 0.65 0.34 0.33 0.33 0.33 0.65 0.65
Between Glazingsc
0.50
45°
0.48 (0.41)/0.59 0.50 (0.43)/0.60 0.62 (0.58)/0.72 0.64 (0.60)/0.73 0.45 (0.37)/0.56 0.47 (0.40)/0.58 0.58 (0.54)/0.69 0.60 (0.57)/0.71 0.57 (0.52)/0.68 0.59 (0.55)/0.69 0.7 0.69 0.69 0.68 0.43 0.41 0.4 0.39 0.69 0.68
Closed
0.36 (0.41)/0.46 0.38 (0.43)/0.48 0.51 (0.57)/0.61 0.53 (0.58)/0.63 0.32 (0.37)/0.43 0.34 (0.40)/0.45 0.46 (0.53)/0.57 0.49 (0.55)/0.60 0.45 (0.51)/0.56 0.48 (0.53)/0.58 0.64 0.64 0.65 0.65 0.31 0.31 0.32 0.32 0.64 0.64
Worsta
0.97 (1.06)/0.65 0.98 (1.06)/0.66 0.98 (1.08)/0.74 0.98 (1.08)/0.75 0.97 (1.07)/0.63 0.98 (1.07)/0.65 0.98 (1.09)/0.72 0.99 (1.09)/0.74 0.98 (1.08)/0.71 0.99 (1.08)/0.72 0.9 0.88 0.89 0.86 0.8 0.75 0.77 0.71 0.88 0.85
0°
0.97 (0.45)/0.69 0.98 (0.47)/0.71 0.98 (0.58)/0.78 0.98 (0.60)/0.79 0.97 (0.42)/0.68 0.98 (0.44)/0.69 0.98 (0.55)/0.77 0.99 (0.57)/0.78 0.98 (0.54)/0.75 0.99 (0.56)/0.77 0.83 0.81 0.81 0.79 0.66 0.62 0.62 0.59 0.81 0.79
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Indoor Side
45°
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
29a
29b
32a
32b
32c
32d
40a
40b
40c
40d
Excluded Beamb 0.33 (0.45)/0.68 0.34 (0.47)/0.69 0.44 (0.58)/0.77 0.46 (0.60)/0.78 0.30 (0.42)/0.66 0.32 (0.44)/0.68 0.41 (0.55)/0.75 0.43 (0.57)/0.77 0.40 (0.54)/0.74 0.42 (0.56)/0.75 0.71 0.7 0.7 0.69 0.46 0.43 0.42 0.4 0.7 0.69
Between Glazingsc
0.80
45°
0.47 (0.35)/0.59 0.49 (0.37)/0.60 0.57 (0.47)/0.68 0.59 (0.48)/0.69 0.45 (0.32)/0.57 0.47 (0.35)/0.59 0.55 (0.44)/0.66 0.56 (0.46)/0.68 0.54 (0.43)/0.65 0.56 (0.45)/0.67 0.74 0.73 0.72 0.71 0.51 0.48 0.47 0.44 0.72 0.71
Closed
0.28 (0.29)/0.39 0.29 (0.31)/0.41 0.38 (0.40)/0.50 0.40 (0.41)/0.51 0.25 (0.27)/0.37 0.27 (0.29)/0.39 0.35 (0.37)/0.48 0.37 (0.39)/0.49 0.34 (0.36)/0.46 0.36 (0.38)/0.48 0.66 0.66 0.66 0.66 0.36 0.35 0.35 0.34 0.66 0.65
Worsta
0.97 (1.11)/0.68 0.98 (1.11)/0.69 0.97 (1.11)/0.73 0.98 (1.11)/0.73 0.98 (1.12)/0.67 0.98 (1.12)/0.69 0.98 (1.12)/0.72 0.99 (1.12)/0.73 0.98 (1.11)/0.71 0.99 (1.11)/0.72 0.91 0.89 0.9 0.87 0.81 0.76 0.79 0.73 0.89 0.86
0°
0.97 (0.48)/0.74 0.98 (0.49)/0.75 0.97 (0.55)/0.78 0.98 (0.56)/0.79 0.98 (0.47)/0.73 0.98 (0.48)/0.74 0.98 (0.54)/0.78 0.99 (0.55)/0.79 0.98 (0.53)/0.77 0.99 (0.54)/0.78 0.87 0.85 0.85 0.83 0.75 0.7 0.71 0.66 0.85 0.82
Fenestration
Table 14F IAC Values for Louvered Shades: Triple Glazing (Continued) Glazing ID:
Excluded Beamb 0.32 (0.48)/0.72 0.33 (0.49)/0.73 0.38 (0.55)/0.76 0.39 (0.56)/0.77 0.31 (0.47)/0.71 0.32 (0.48)/0.72 0.37 (0.54)/0.76 0.38 (0.55)/0.77 0.36 (0.53)/0.75 0.38 (0.54)/0.76 0.81 0.79 0.78 0.76 0.63 0.59 0.58 0.54 0.78 0.76
Licensed for single user. © 2017 ASHRAE, Inc.
Outdoor Side
0.15
45°
0.48 (0.31)/0.61 0.49 (0.32)/0.62 0.54 (0.37)/0.65 0.55 (0.38)/0.66 0.47 (0.30)/0.60 0.49 (0.32)/0.62 0.52 (0.36)/0.65 0.54 (0.37)/0.66 0.52 (0.35)/0.64 0.53 (0.36)/0.65 0.82 0.79 0.79 0.77 0.65 0.6 0.59 0.55 0.79 0.77
Closed
0.22 (0.19)/0.36 0.23 (0.21)/0.37 0.27 (0.24)/0.41 0.29 (0.26)/0.42 0.21 (0.18)/0.35 0.22 (0.20)/0.37 0.26 (0.23)/0.40 0.28 (0.25)/0.41 0.25 (0.23)/0.39 0.27 (0.24)/0.40 0.73 0.72 0.71 0.7 0.49 0.46 0.45 0.43 0.72 0.7
Worsta
0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.89)/0.35 0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.90)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.93 (0.89)/0.35 0.92 0.9 0.91 0.88 0.83 0.78 0.81 0.75 0.9 0.87
0°
0.93 (0.04)/0.41 0.93 (0.05)/0.41 0.93 (0.04)/0.40 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.41 0.93 (0.04)/0.40 0.93 (0.04)/0.41 0.93 (0.04)/0.40 0.93 (0.04)/0.40 0.88 0.86 0.88 0.85 0.77 0.72 0.76 0.7 0.87 0.84
Excluded Beamb 0.02 (0.04)/0.39 0.02 (0.05)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.02 (0.04)/0.39 0.78 0.76 0.79 0.77 0.59 0.55 0.59 0.55 0.79 0.77
Outdoor Side
0.50
45°
0.19 (0.02)/0.29 0.19 (0.03)/0.29 0.19 (0.02)/0.28 0.19 (0.02)/0.28 0.19 (0.02)/0.29 0.19 (0.02)/0.29 0.19 (0.02)/0.28 0.19 (0.02)/0.28 0.19 (0.02)/0.28 0.19 (0.02)/0.28 0.83 0.81 0.83 0.81 0.67 0.63 0.66 0.62 0.83 0.8
Closed
0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.01 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.02 (0.02)/0.09 0.01 (0.01)/0.09 0.01 (0.02)/0.09 0.01 (0.01)/0.09 0.01 (0.02)/0.09 0.66 0.66 0.67 0.66 0.36 0.35 0.37 0.35 0.67 0.66
Worsta
0.94 (1.00)/0.44 0.94 (0.99)/0.44 0.94 (0.99)/0.44 0.94 (0.99)/0.44 0.94 (1.00)/0.44 0.94 (1.00)/0.44 0.94 (1.00)/0.44 0.94 (0.99)/0.44 0.94 (0.99)/0.44 0.94 (0.99)/0.43 0.92 0.9 0.91 0.88 0.84 0.79 0.81 0.75 0.9 0.87
0°
0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.94 (0.14)/0.50 0.92 0.89 0.91 0.88 0.83 0.78 0.81 0.75 0.9 0.87
Excluded Beamb 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.48 0.07 (0.14)/0.47 0.89 0.86 0.88 0.85 0.78 0.73 0.76 0.7 0.88 0.85
Outdoor Side
0.80
45°
0.24 (0.06)/0.36 0.25 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.24 (0.06)/0.36 0.9 0.87 0.89 0.86 0.8 0.74 0.78 0.72 0.88 0.85
Closed
0.04 (0.02)/0.14 0.04 (0.02)/0.14 0.03 (0.02)/0.13 0.04 (0.02)/0.13 0.03 (0.02)/0.14 0.04 (0.02)/0.14 0.03 (0.02)/0.13 0.03 (0.02)/0.13 0.03 (0.02)/0.13 0.03 (0.02)/0.13 0.82 0.79 0.82 0.79 0.65 0.61 0.65 0.6 0.82 0.79
Worsta
0.95 (1.11)/0.56 0.95 (1.10)/0.56 0.95 (1.10)/0.56 0.95 (1.08)/0.55 0.95 (1.11)/0.56 0.95 (1.10)/0.56 0.95 (1.10)/0.56 0.95 (1.09)/0.55 0.95 (1.10)/0.55 0.95 (1.08)/0.55 0.92 0.9 0.91 0.88 0.84 0.79 0.81 0.75 0.9 0.87
0°
0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.95 (0.29)/0.62 0.93 0.9 0.92 0.89 0.84 0.79 0.82 0.76 0.9 0.87
Excluded Beamb 0.16 (0.29)/0.60 0.16 (0.29)/0.60 0.16 (0.29)/0.60 0.16 (0.29)/0.59 0.15 (0.29)/0.60 0.15 (0.29)/0.60 0.15 (0.29)/0.60 0.15 (0.29)/0.60 0.15 (0.29)/0.60 0.15 (0.29)/0.59 0.91 0.88 0.9 0.87 0.82 0.76 0.79 0.73 0.89 0.86
aLouvers bLouvers
0.33 (0.13)/0.48 0.33 (0.13)/0.48 0.33 (0.12)/0.48 0.33 (0.12)/0.47 0.33 (0.13)/0.48 0.33 (0.13)/0.48 0.33 (0.12)/0.48 0.33 (0.12)/0.47 0.33 (0.12)/0.48 0.33 (0.12)/0.47 0.92 0.89 0.91 0.88 0.83 0.77 0.8 0.74 0.9 0.87
Closed
0.08 (0.04)/0.22 0.08 (0.04)/0.22 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.22 0.08 (0.04)/0.22 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.08 (0.04)/0.21 0.89 0.87 0.89 0.86 0.79 0.74 0.77 0.71 0.88 0.85
track so that profile angle equals negative slat angle and maximum direct beam is admitted. track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
c
Glazing cavity width equals original cavity width plus slat width. is radiant fraction; ratio of radiative heat transfer to total heat transfer, on room side of glazing system.
dF R
15.49
Notes:
45°
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
15.50
Table 14G IAC Values for Draperies, Roller Shades, and Insect Screens Drapery Glazing ID: Shade
1b
1c
1d
1e
1f
1g
1h
1i
0.74, 0.44 0.66, 0.46 0.56, 0.50 0.78, 0.49 0.71, 0.54 0.65, 0.64 0.83, 0.56 0.76, 0.63 0.71, 0.72 0.78, 0.76
0.72, 0.47 0.63, 0.49 0.51, 0.54 0.76, 0.52 0.68, 0.58 0.61, 0.70 0.82, 0.59 0.74, 0.68 0.68, 0.78 0.75, 0.82
0.74, 0.44 0.66, 0.45 0.56, 0.49 0.78, 0.49 0.71, 0.54 0.65, 0.63 0.83, 0.56 0.76, 0.63 0.72, 0.71 0.78, 0.75
0.74, 0.45 0.65, 0.46 0.55, 0.51 0.78, 0.49 0.70, 0.55 0.64, 0.65 0.82, 0.57 0.76, 0.64 0.71, 0.73 0.77, 0.77
IAC, FRd
Dark Closed Weave Medium Closed Weave Light Closed Weave Dark Semiopen Weave Medium Semiopen Weave Light Semiopen Weave Dark Open Weave Medium Open Weave Light Open Weave Sheer
IIID IIIM IIIL IID IIM IIL ID IM IL
100% 100% 100% 100% 100% 100% 100% 100% 100% 100%
5a
5b
5c
5d
5e
5f
5g
5h
5i
Dark Closed Weave
IIID
100%
0.81, 0.46
0.82, 0.45
0.82, 0.44
0.83, 0.42
0.82, 0.44
0.83, 0.42
0.82, 0.44
0.84, 0.42
0.83, 0.42
Medium Closed Weave Light Closed Weave Dark Semiopen Weave Medium Semiopen Weave Light Semiopen Weave Dark Open Weave Medium Open Weave Light Open Weave Sheer
IIIM IIIL IID IIM IIL ID IM IL
100% 100% 100% 100% 100% 100% 100% 100% 100%
0.70, 0.48 0.57, 0.54 0.84, 0.51 0.75, 0.57 0.65, 0.70 0.88, 0.59 0.79, 0.68 0.72, 0.79 0.78, 0.83
0.72, 0.46 0.60, 0.52 0.85, 0.50 0.76, 0.55 0.68, 0.67 0.88, 0.57 0.80, 0.65 0.74, 0.76 0.8, 0.8
0.72, 0.46 0.59, 0.52 0.85, 0.49 0.76, 0.55 0.67, 0.67 0.88, 0.57 0.80, 0.65 0.73, 0.76 0.8, 0.8
0.75, 0.43 0.64, 0.47 0.86, 0.47 0.79, 0.51 0.71, 0.61 0.89, 0.54 0.82, 0.61 0.77, 0.69 0.82, 0.73
0.72, 0.46 0.60, 0.51 0.85, 0.49 0.76, 0.55 0.68, 0.66 0.88, 0.57 0.80, 0.65 0.74, 0.75 0.8, 0.79
0.75, 0.43 0.65, 0.47 0.86, 0.47 0.79, 0.51 0.72, 0.60 0.89, 0.54 0.82, 0.60 0.77, 0.69 0.82, 0.73
0.72, 0.46 0.60, 0.51 0.85, 0.49 0.76, 0.55 0.68, 0.66 0.88, 0.57 0.80, 0.65 0.74, 0.75 0.8, 0.79
0.75, 0.43 0.65, 0.47 0.86, 0.46 0.79, 0.51 0.72, 0.60 0.89, 0.54 0.82, 0.60 0.77, 0.68 0.82, 0.73
0.75, 0.43 0.64, 0.47 0.86, 0.47 0.79, 0.52 0.72, 0.61 0.89, 0.54 0.82, 0.61 0.77, 0.69 0.82, 0.74
17a
17b
17c
17d
17e
17f
17g
17h
17i
17j
17k
Dark Closed Weave
IIID
100%
0.85, 0.45
0.86, 0.43
0.86, 0.44
0.87, 0.43
0.86, 0.43
0.88, 0.41
0.87, 0.42
0.88, 0.40
0.87, 0.42
0.88, 0.40
0.88, 0.41
Medium Closed Weave Light Closed Weave Dark Semiopen Weave Medium Semiopen Weave Light Semiopen Weave Dark Open Weave Medium Open Weave Light Open Weave Sheer
IIIM IIIL IID IIM IIL ID IM IL
100% 100% 100% 100% 100% 100% 100% 100% 100%
0.74, 0.47 0.60, 0.52 0.88, 0.50 0.78, 0.56 0.68, 0.68 0.90, 0.58 0.81, 0.66 0.74, 0.77 0.8, 0.81
0.76, 0.45 0.64, 0.50 0.89, 0.48 0.8, 0.54 0.71, 0.64 0.91, 0.56 0.83, 0.64 0.76, 0.73 0.82, 0.77
0.76, 0.45 0.64, 0.50 0.88, 0.49 0.8, 0.54 0.71, 0.64 0.91, 0.56 0.83, 0.64 0.76, 0.73 0.82, 0.78
0.79, 0.44 0.68, 0.47 0.89, 0.47 0.82, 0.52 0.74, 0.61 0.91, 0.55 0.85, 0.61 0.79, 0.69 0.84, 0.74
0.77, 0.44 0.66, 0.48 0.89, 0.48 0.81, 0.52 0.72, 0.62 0.91, 0.55 0.84, 0.62 0.77, 0.71 0.83, 0.75
0.80, 0.42 0.71, 0.44 0.90, 0.45 0.83, 0.49 0.76, 0.57 0.92, 0.52 0.86, 0.58 0.81, 0.65 0.85, 0.7
0.78, 0.44 0.66, 0.48 0.89, 0.47 0.81, 0.52 0.73, 0.62 0.91, 0.55 0.84, 0.62 0.78, 0.71 0.83, 0.75
0.80, 0.41 0.71, 0.44 0.90, 0.45 0.83, 0.49 0.77, 0.57 0.92, 0.52 0.86, 0.58 0.81, 0.65 0.85, 0.69
0.78, 0.44 0.66, 0.48 0.89, 0.47 0.81, 0.52 0.73, 0.62 0.91, 0.55 0.84, 0.62 0.78, 0.71 0.83, 0.75
0.80, 0.41 0.71, 0.44 0.90, 0.45 0.83, 0.49 0.77, 0.57 0.92, 0.52 0.86, 0.58 0.81, 0.65 0.85, 0.69
0.80, 0.42 0.71, 0.45 0.90, 0.45 0.83, 0.49 0.76, 0.57 0.92, 0.52 0.86, 0.58 0.81, 0.65 0.85, 0.7
21a
21b
21c
21d
21e
21f
21g
21h
21i
21j
21k
Dark Closed Weave
IIID
100%
0.87, 0.44
0.88, 0.43
0.88, 0.43
0.88, 0.42
0.88, 0.42
0.89, 0.40
0.88, 0.42
0.89, 0.40
0.88, 0.42
0.89, 0.40
0.89, 0.40
Medium Closed Weave Light Closed Weave Dark Semiopen Weave Medium Semiopen Weave Light Semiopen Weave Dark Open Weave Medium Open Weave Light Open Weave Sheer
IIIM IIIL IID IIM IIL ID IM IL
100% 100% 100% 100% 100% 100% 100% 100% 100%
0.77, 0.46 0.64, 0.52 0.89, 0.49 0.8, 0.55 0.71, 0.67 0.92, 0.57 0.83, 0.65 0.76, 0.77 0.82, 0.81
0.79, 0.45 0.67, 0.49 0.90, 0.48 0.82, 0.53 0.74, 0.64 0.92, 0.56 0.85, 0.63 0.78, 0.73 0.83, 0.77
0.79, 0.45 0.68, 0.49 0.90, 0.48 0.82, 0.53 0.74, 0.63 0.92, 0.55 0.85, 0.63 0.79, 0.72 0.84, 0.76
0.81, 0.43 0.70, 0.47 0.90, 0.47 0.83, 0.51 0.76, 0.61 0.92, 0.54 0.86, 0.61 0.81, 0.69 0.85, 0.74
0.80, 0.43 0.69, 0.47 0.90, 0.47 0.83, 0.52 0.75, 0.61 0.92, 0.54 0.86, 0.61 0.80, 0.70 0.85, 0.74
0.82, 0.41 0.73, 0.45 0.91, 0.45 0.85, 0.49 0.78, 0.57 0.93, 0.52 0.87, 0.58 0.82, 0.65 0.86, 0.7
0.80, 0.43 0.70, 0.47 0.90, 0.47 0.83, 0.51 0.76, 0.61 0.92, 0.54 0.86, 0.61 0.80, 0.69 0.85, 0.74
0.82, 0.41 0.73, 0.44 0.91, 0.45 0.85, 0.49 0.78, 0.57 0.93, 0.52 0.87, 0.58 0.82, 0.65 0.86, 0.69
0.80, 0.43 0.70, 0.47 0.90, 0.47 0.83, 0.51 0.76, 0.61 0.92, 0.54 0.86, 0.61 0.80, 0.69 0.85, 0.74
0.82, 0.41 0.73, 0.44 0.91, 0.45 0.85, 0.49 0.78, 0.57 0.93, 0.52 0.87, 0.58 0.82, 0.65 0.86, 0.69
0.82, 0.41 0.73, 0.45 0.91, 0.45 0.85, 0.49 0.78, 0.57 0.93, 0.52 0.87, 0.58 0.82, 0.66 0.86, 0.7
Glazing ID:
Glazing ID:
0.71, 0.50 0.59, 0.53 0.45, 0.62 0.75, 0.55 0.65, 0.63 0.56, 0.79 0.80, 0.63 0.71, 0.73 0.64, 0.87 0.73, 0.89
0.71, 0.49 0.60, 0.52 0.46, 0.60 0.75, 0.54 0.66, 0.62 0.57, 0.77 0.80, 0.62 0.72, 0.72 0.65, 0.85 0.73, 0.88
0.72, 0.47 0.62, 0.49 0.50, 0.56 0.76, 0.52 0.68, 0.59 0.60, 0.71 0.81, 0.60 0.73, 0.69 0.68, 0.80 0.75, 0.83
0.74, 0.45 0.65, 0.46 0.55, 0.50 0.78, 0.49 0.70, 0.55 0.64, 0.65 0.82, 0.57 0.76, 0.64 0.71, 0.73 0.77, 0.77
0.72, 0.47 0.63, 0.49 0.51, 0.54 0.76, 0.52 0.68, 0.58 0.61, 0.70 0.82, 0.59 0.74, 0.68 0.68, 0.78 0.75, 0.82
2017 ASHRAE Handbook—Fundamentals (SI)
Fullness
Glazing ID:
Licensed for single user. © 2017 ASHRAE, Inc.
1a
Fabric Designator
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Glazing ID:
25b
25c
26d
25e
25f
0.89, 0.42
0.90, 0.40
0.90, 0.40
0.91, 0.39
0.90, 0.40
0.85, 0.41
0.84, 0.41
0.85, 0.40
0.76, 0.44
0.76, 0.45
0.77, 3
0.92, 0.44
0.92, 0.45
0.92, 0.43
0.87, 0.48
0.86, 0.49
0.87, 0.47
IIID
100%
Medium Closed Weave
IIIM
100%
0.80, 0.45
0.82, 0.44
Light Closed Weave
IIIL
100%
0.68, 0.51
0.72, 0.48
Dark Semiopen Weave
IID
100%
0.91, 0.48
0.91, 0.47
Medium Semiopen Weave
IIM
100%
0.83, 0.54
0.85, 0.52
29a
29b
Dark Closed Weave
IIID
Glazing ID: 100%
0.86, 0.44
0.87, 0.42
0.84, 0.42
Medium Closed Weave
IIIM
100%
0.77, 0.45
0.80, 0.43
0.76, 0.45
Light Closed Weave
IIIL
100%
0.65, 0.50
0.69, 0.47
0.92, 0.45
Dark Semiopen Weave
IID
100%
0.89, 0.48
0.90, 0.47
0.86, 0.49
Medium Semiopen Weave
IIM
100%
0.81, 0.54
0.83, 0.51
Light Semiopen Weave
IIL
100%
0.75, 0.66
0.78, 0.63
0.81, 0.57
0.81, 0.58
0.82, 0.55
0.81, 0.58
Light Semiopen Weave
IIL
100%
0.72, 0.65
0.76, 0.60
Dark Open Weave
ID
100%
0.93, 0.56
0.93, 0.55
0.94, 0.51
0.94, 0.52
0.94, 0.50
0.94, 0.52
Dark Open Weave
ID
100%
0.91, 0.56
0.92, 0.54
Medium Open Weave
IM
100%
0.86, 0.65
0.87, 0.62
0.89, 0.57
0.89, 0.58
0.89, 0.56
0.89, 0.59
Medium Open Weave
IM
100%
0.84, 0.64
0.85, 0.60
Light Open Weave
IL
100%
0.79, 0.75
0.82, 0.72
0.85, 0.65
0.84, 0.66
0.85, 0.63
0.84, 0.67
Light Open Weave
IL
100%
0.77, 0.73
0.80, 0.69
100%
0.84, 0.8
0.86, 0.76
0.88, 0.69
0.88, 0.7
0.89, 0.67
0.88, 0.71
Sheer
0.83, 0.78
0.85, 0.73
Sheer
Glazing ID:
Licensed for single user. © 2017 ASHRAE, Inc.
25a 0.88, 0.43
Dark Closed Weave
100%
100%
32a
32b
32c
32d
40a
40b
40c
40d
0.89, 0.43
0.90, 0.41
0.91, 0.42
0.92, 0.39
0.93, 0.41
0.94, 0.37
0.90, 0.42
0.91, 0.40
Dark Closed Weave
IIID
Medium Closed Weave
IIIM
100%
0.80, 0.44
0.83, 0.42
0.82, 0.42
0.84, 0.39
0.84, 0.40
0.87, 0.37
0.82, 0.43
0.85, 0.41
Light Closed Weave
IIIL
100%
0.69, 0.48
0.73, 0.45
0.69, 0.44
0.73, 0.40
0.73, 0.42
0.77, 0.38
0.71, 0.47
0.76, 0.44
Dark Semiopen Weave
IID
100%
0.91, 0.47
0.92, 0.46
0.93, 0.46
0.94, 0.43
0.94, 0.45
0.95, 0.41
0.92, 0.46
0.93, 0.45
Medium Semiopen Weave
IIM
100%
0.83, 0.52
0.85, 0.50
0.84, 0.50
0.86, 0.46
0.86, 0.48
0.89, 0.44
0.85, 0.51
0.87, 0.48
Light Semiopen Weave
IIL
100%
0.75, 0.62
0.79, 0.58
0.75, 0.58
0.79, 0.52
0.78, 0.55
0.82, 0.49
0.77, 0.60
0.80, 0.56
Dark Open Weave
ID
100%
0.93, 0.55
0.93, 0.53
0.95, 0.53
0.95, 0.50
0.96, 0.51
0.96, 0.47
0.94, 0.54
0.94, 0.52
Medium Open Weave
IM
100%
0.86, 0.62
0.87, 0.59
0.87, 0.59
0.88, 0.54
0.88, 0.56
0.90, 0.51
0.87, 0.60
0.89, 0.57
Light Open Weave
IL
100%
0.80, 0.71
0.83, 0.66
0.80, 0.66
0.82, 0.60
0.82, 0.63
0.85, 0.57
0.81, 0.69
0.84, 0.64
100%
0.84, 0.75
0.87, 0.7
0.85, 0.71
0.87, 0.65
0.86, 0.68
0.89, 0.61
0.86, 0.73
0.88, 0.68
Sheer
Fenestration
Table 14G IAC Values for Draperies, Roller Shades, and Insect Screens (Continued)
Roller Shades and Insect Screens Shade/Screen
Openness
Refl./Trans. Glazing ID:
1a
1b
1c
1d
1e
1f
1g
1h
1i
Light Translucent
0.14
0.60/0.25
0.44, 0.74
0.45, 0.72
0.49, 0.66
0.55, 0.59
0.51, 0.65
0.56, 0.59
0.51, 0.65
0.57, 0.58
0.55, 0.6
White Opaque
0.00
0.65/0.00
0.34, 0.45
0.35, 0.44
0.4, 0.41
0.47, 0.38
0.42, 0.4
0.48, 0.38
0.42, 0.4
0.49, 0.38
0.47, 0.38
Dark Opaque
0.00
0.20/0.00
0.64, 0.48
0.65, 0.47
0.67, 0.45
0.69, 0.43
0.67, 0.45
0.7, 0.42
0.67, 0.45
0.7, 0.42
0.69, 0.43
Light Gray Translucent
0.10
0.31/0.15
0.61, 0.57
0.62, 0.57
0.64, 0.54
0.68, 0.51
0.65, 0.53
0.68, 0.5
0.65, 0.53
0.69, 0.5
0.68, 0.51
Dark Gray Translucent
0.14
0.17/0.19
0.71, 0.58
0.72, 0.58
0.73, 0.55
0.76, 0.52
0.74, 0.55
0.76, 0.52
0.74, 0.55
0.76, 0.52
0.76, 0.52
0.00
0.84/0.00
0.3, 0.71
0.32, 0.68
0.38, 0.6
0.45, 0.53
0.39, 0.58
0.46, 0.52
0.39, 0.58
0.47, 0.52
0.45, 0.53
Reflective White Translucent
0.07
0.75/0.16
0.23, 0.42
0.25, 0.41
0.31, 0.38
0.39, 0.36
0.33, 0.37
0.4, 0.35
0.33, 0.37
0.41, 0.35
0.39, 0.36
Outdoor Insect Screen
0.64, 0.98
0.64, 0.98
0.64, 0.95
0.64, 0.92
0.64, 0.95
0.64, 0.91
0.64, 0.95
0.64, 0.91
0.64, 0.92
Indoor Insect Screen
0.88, 0.81
0.88, 0.8
0.89, 0.78
0.9, 0.75
0.89, 0.78
0.9, 0.75
0.89, 0.78
0.9, 0.75
0.9, 0.76
5a 0.55, 0.65 0.48, 0.4 0.76, 0.44 0.72, 0.53 0.81, 0.54 0.43, 0.6 0.37, 0.38 0.64, 0.96 0.92, 0.78
5b 0.58, 0.62 0.52, 0.39 0.77, 0.43 0.74, 0.51 0.82, 0.53 0.47, 0.55 0.42, 0.36 0.64, 0.94 0.93, 0.77
5c 0.58, 0.62 0.51, 0.39 0.77, 0.43 0.74, 0.51 0.82, 0.53 0.46, 0.56 0.41, 0.36 0.64, 0.94 0.93, 0.76
5d 0.63, 0.56 0.57, 0.37 0.8, 0.41 0.77, 0.48 0.84, 0.5 0.54, 0.5 0.49, 0.34 0.64, 0.91 0.93, 0.74
5e 0.58, 0.61 0.52, 0.39 0.78, 0.43 0.74, 0.51 0.82, 0.52 0.47, 0.55 0.42, 0.36 0.64, 0.94 0.93, 0.76
5f 0.64, 0.56 0.58, 0.37 0.8, 0.41 0.77, 0.48 0.84, 0.5 0.55, 0.49 0.5, 0.34 0.64, 0.9 0.93, 0.74
5g 0.58, 0.61 0.52, 0.39 0.78, 0.43 0.74, 0.51 0.82, 0.52 0.47, 0.55 0.42, 0.36 0.64, 0.94 0.93, 0.76
5h 0.64, 0.56 0.58, 0.36 0.8, 0.4 0.77, 0.48 0.84, 0.5 0.55, 0.49 0.5, 0.34 0.64, 0.9 0.93, 0.73
5i 0.63, 0.56 0.57, 0.37 0.8, 0.41 0.77, 0.48 0.84, 0.5 0.54, 0.5 0.49, 0.34 0.64, 0.91 0.93, 0.74
Light Translucent White Opaque Dark Opaque Light Gray Translucent Dark Gray Translucent Reflective White Opaque Reflective White Translucent Outdoor Insect Screen Indoor Insect Screen
0.14 0.00 0.00 0.10 0.14 0.00 0.07
Glazing ID: 0.60/0.25 0.65/0.00 0.20/0.00 0.31/0.15 0.17/0.19 0.84/0.00 0.75/0.16
15.51
Reflective White Opaque
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
17a
17b
17c
17d
17e
17f
17g
17h
17i
17j
17k
Light Translucent
0.14
Glazing ID: 0.60/0.25
0.58, 0.63
0.62, 0.6
0.62, 0.6
0.67, 0.56
0.64, 0.58
0.69, 0.53
0.64, 0.57
0.7, 0.53
0.64, 0.57
0.7, 0.53
0.7, 0.53
White Opaque
0.00
0.65/0.00
0.52, 0.39
0.56, 0.38
0.57, 0.38
0.61, 0.37
0.59, 0.37
0.65, 0.36
0.59, 0.37
0.65, 0.36
0.59, 0.37
0.65, 0.35
0.65, 0.36
Dark Opaque
0.00
0.20/0.00
0.8, 0.43
0.82, 0.42
0.82, 0.42
0.83, 0.41
0.82, 0.41
0.84, 0.39
0.82, 0.41
0.84, 0.39
0.82, 0.41
0.84, 0.39
0.84, 0.39
Light Gray Translucent
0.10
0.31/0.15
0.76, 0.52
0.78, 0.5
0.78, 0.5
0.8, 0.48
0.78, 0.49
0.81, 0.46
0.79, 0.49
0.82, 0.46
0.79, 0.49
0.82, 0.46
0.81, 0.46
Dark Gray Translucent
0.14
0.17/0.19
0.84, 0.53
0.85, 0.52
0.85, 0.52
0.86, 0.5
0.86, 0.51
0.87, 0.48
0.86, 0.51
0.88, 0.48
0.86, 0.51
0.88, 0.48
0.87, 0.48
Reflective White Opaque
0.00
0.84/0.00
0.46, 0.57
0.52, 0.53
0.52, 0.53
0.58, 0.5
0.54, 0.51
0.61, 0.47
0.55, 0.51
0.62, 0.46
0.55, 0.51
0.62, 0.46
0.61, 0.47
Reflective White Translucent
0.07
0.75/0.16
0.41, 0.37
0.47, 0.36
0.47, 0.36
0.53, 0.35
0.49, 0.35
0.57, 0.34
0.5, 0.35
0.58, 0.34
0.5, 0.35
0.58, 0.33
0.57, 0.34
Outdoor Insect Screen
0.64, 0.95
0.64, 0.93
0.64, 0.94
0.64, 0.91
0.64, 0.92
0.64, 0.89
0.64, 0.92
0.64, 0.89
0.64, 0.92
0.64, 0.89
0.64, 0.89
Indoor Insect Screen
0.94, 0.77
0.94, 0.76
0.94, 0.76
0.94, 0.74
0.94, 0.75
0.95, 0.72
0.94, 0.75
0.95, 0.72
0.94, 0.75
0.95, 0.72
0.95, 0.72
21a
21b
21c
21d
21e
21f
21g
21h
21i
21j
21k
0.65, 0.59
0.66, 0.59
0.69, 0.56
0.68, 0.57
0.72, 0.53
0.68, 0.57
0.72, 0.53
0.68, 0.57
0.72, 0.53
0.72, 0.53
Glazing ID: Light Translucent
0.14
0.60/0.25
0.61, 0.63
White Opaque
0.00
0.65/0.00
0.55, 0.39
0.6, 0.38
0.61, 0.38
0.64, 0.37
0.63, 0.37
0.67, 0.36
0.63, 0.37
0.67, 0.36
0.63, 0.37
0.68, 0.35
0.67, 0.36
Dark Opaque
0.00
0.20/0.00
0.82, 0.43
0.84, 0.42
0.84, 0.42
0.85, 0.41
0.85, 0.41
0.86, 0.39
0.85, 0.41
0.86, 0.39
0.85, 0.41
0.86, 0.39
0.86, 0.39
Light Gray Translucent
0.10
0.31/0.15
0.78, 0.51
0.8, 0.5
0.8, 0.5
0.82, 0.48
0.81, 0.48
0.83, 0.46
0.81, 0.48
0.83, 0.46
0.81, 0.48
0.83, 0.46
0.83, 0.46
Dark Gray Translucent
0.14
0.17/0.19
0.86, 0.53
0.87, 0.51
0.87, 0.51
0.88, 0.5
0.88, 0.5
0.89, 0.48
0.88, 0.5
0.89, 0.48
0.88, 0.5
0.89, 0.48
0.89, 0.48
Reflective White Opaque
0.00
0.84/0.00
0.5, 0.56
0.55, 0.53
0.56, 0.52
0.6, 0.5
0.58, 0.51
0.63, 0.47
0.59, 0.5
0.64, 0.47
0.59, 0.5
0.64, 0.47
0.64, 0.47
Reflective White Translucent
0.07
0.75/0.16
0.44, 0.36
0.5, 0.35
0.51, 0.35
0.56, 0.35
0.54, 0.35
0.6, 0.34
0.54, 0.35
0.6, 0.34
0.54, 0.35
0.6, 0.33
0.6, 0.34
0.64, 0.95
0.64, 0.93
0.64, 0.93
0.64, 0.91
0.64, 0.92
0.64, 0.89
0.64, 0.91
0.64, 0.89
0.64, 0.91
0.64, 0.89
0.64, 0.89
0.94, 0.77
0.95, 0.72
0.95, 0.74
0.95, 0.72
0.95, 0.74
0.95, 0.72
0.95, 0.72
Outdoor Insect Screen Indoor Insect Screen
0.95, 0.75
0.95, 0.75
25a
25b
25c
26d
0.95, 0.74 25e
0.95, 0.74 25f
0.66, 0.62
0.71, 0.58
0.75, 0.53
0.75, 0.54
0.77, 0.52
0.74, 0.55
Light Translucent
0.14
0.60/0.25
White Opaque
0.00
0.65/0.00
0.6, 0.38
0.66, 0.37
0.71, 0.35
0.7, 0.36
0.72, 0.35
Dark Opaque
0.00
0.20/0.00
0.85, 0.42
0.86, 0.41
0.88, 0.39
0.88, 0.39
0.88, 0.38
Light Gray Translucent
0.10
0.31/0.15
0.81, 0.5
0.83, 0.49
0.86, 0.46
0.85, 0.46
0.86, 0.45
Dark Gray Translucent
0.14
0.17/0.19
0.88, 0.52
0.89, 0.51
0.9, 0.47
0.9, 0.48
0.91, 0.47
Reflective White Opaque
0.00
0.84/0.00
0.55, 0.55
0.61, 0.52
0.68, 0.47
0.67, 0.48
Reflective White Translucent
0.07
0.75/0.16
Glazing ID:
29a
29b 0.68, 0.56
Light Translucent
0.14
0.60/0.25
0.64, 0.6
0.7, 0.36
White Opaque
0.00
0.65/0.00
0.58, 0.38
0.63, 0.37
0.87, 0.39
Dark Opaque
0.00
0.20/0.00
0.82, 0.42
0.84, 0.41
0.85, 0.47
Light Gray Translucent
0.10
0.31/0.15
0.78, 0.5
0.81, 0.48
0.9, 0.48
Dark Gray Translucent
0.14
0.17/0.19
0.86, 0.52
0.87, 0.5
0.69, 0.46
0.66, 0.48
Reflective White Opaque
0.00
0.84/0.00
0.53, 0.53
0.59, 0.49
0.07
0.75/0.16
0.48, 0.36
0.55, 0.35
0.5, 0.36
0.57, 0.35
0.64, 0.33
0.62, 0.34
0.65, 0.33
0.62, 0.34
Reflective White Translucent
Outdoor Insect Screen
0.65, 0.95
0.65, 0.93
0.64, 0.88
0.64, 0.89
0.64, 0.87
0.64, 0.89
Outdoor Insect Screen
0.64, 0.94
0.64, 0.91
Indoor Insect Screen
0.95, 0.76
0.96, 0.75
0.96, 0.72
0.96, 0.72
0.96, 0.71
0.96, 0.72
Indoor Insect Screen
0.94, 0.76
0.95, 0.74
32a
32b
32c
32d
40a
40b
40c
40d
Light Translucent
0.14
Glazing ID: 0.60/0.25
0.67, 0.58
0.72, 0.54
0.67, 0.52
0.71, 0.46
0.7, 0.49
0.75, 0.43
0.69, 0.56
0.74, 0.52
White Opaque
0.00
0.65/0.00
0.62, 0.37
0.68, 0.36
0.62, 0.33
0.67, 0.3
0.67, 0.31
0.72, 0.29
0.65, 0.37
0.71, 0.36
Dark Opaque
0.00
0.20/0.00
0.85, 0.41
0.87, 0.4
0.87, 0.4
0.89, 0.38
0.89, 0.39
0.91, 0.36
0.87, 0.41
0.88, 0.39
Light Gray Translucent
0.10
0.31/0.15
0.81, 0.49
0.84, 0.47
0.82, 0.46
0.85, 0.43
0.85, 0.44
0.87, 0.41
0.83, 0.48
0.85, 0.46
Dark Gray Translucent
0.14
0.17/0.19
0.88, 0.51
0.89, 0.49
0.9, 0.49
0.91, 0.45
0.91, 0.47
0.92, 0.43
0.89, 0.5
0.91, 0.48
Reflective White Opaque
0.00
0.84/0.00
0.57, 0.51
0.64, 0.47
0.57, 0.44
0.63, 0.39
0.61, 0.41
0.68, 0.36
0.6, 0.5
0.67, 0.46
Reflective White Translucent
0.07
0.75/0.16
0.53, 0.35
0.61, 0.34
0.52, 0.28
0.59, 0.26
0.58, 0.27
0.65, 0.25
0.56, 0.35
0.64, 0.34
Outdoor Insect Screen
0.64, 0.92
0.64, 0.9
0.64, 0.86
0.64, 0.8
0.64, 0.83
0.64, 0.77
0.64, 0.91
0.64, 0.88
Indoor Insect Screen
0.95, 0.75
0.96, 0.73
0.95, 0.7
0.95, 0.67
0.96, 0.68
0.96, 0.64
0.96, 0.74
0.96, 0.72
Notes:
aLouvers
track so that profile angle equals negative slat angle and maximum direct beam is admitted. bLouvers track to block direct beam radiation. When negative slat angles result, slat defaults to 0°.
cGlazing
cavity width equals original cavity width plus slat width. dF is radiant fraction; ratio of radiative heat transfer to total heat transfer, on room side of glazing system. R
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Glazing ID:
15.52
Table 14G IAC Values for Draperies, Roller Shades, and Insect Screens (Continued)
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Fenestration
15.53 Table 15 Summary of Environmental Control Capabilities of Draperies Designator (Figure 23)
Item
ID
IM
IL
IID
IIM
IIL
IIID
IIIM
IIIL
1. Protection from direct solar radiation and long-wave radiation to or from window areas
Fair
Fair
Fair
Fair
Good
Good
Fair
Good
Good
2. Effectiveness in allowing outward vision through fenestration
Good
Good
Fair
Fair
Fair
Some
None
None
None
None
Poora
Fair
Faira
Goodb
Goodb
Goodb
3. Effectiveness in attaining privacy (limiting inward vision from outdoors)
None
4. Protection against excessive brightness and glare from sunshine and external objects
Mild
Mild
Mildc Poorc
Good
Good
Goodc Poorc
Good
Good
Goodc Poorc
5. Effectiveness in modifying unattractive or distracting view out of window
Little
Little
Some
Some
Good
Good
Blocks
Blocks
Blocks
aGood bTo
Poor
Gooda
when bright illumination is on viewing side. obscure view completely, material must be completely opaque.
Gooda
cPoor
rating applies to white fabric in direct sunlight. Use off-white color to avoid excessive transmitted light.
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compared to 0.02 for glass and 0.03 for plaster. For drapery fabrics at 100% fullness, NRC ranges from 0.10 to 0.65, depending on the tightness of weave. Class III (tightly woven) fabrics have NRC values of 0.35 to 0.65. Figure 26 shows the relationship between NRC and openness factor for fabrics of normal weight.
Double Drapery Double draperies (two sets of drapery covering the same area) have a light, open weave on the fenestration product side for outward vision and daylight when desired and a heavy, closed weave or opaque drapery on the room side to block out sunlight and provide privacy when desired. When properly selected and used, double draperies can provide a reduced U-factor and a lowered IAC. The reduced U-factor results principally from adding a semiclosed air space to the barrier. To most effectively reduce solar heat gain, drapery exposed to sunlight should have high reflectance and low transmittance. The light, open-weave drapery should be opened when the heavy drapery is closed to prevent entry of sunlight. Properly used double draperies give (1) extreme flexibility of vision and light intensity, (2) a lowered U-factor and IAC, and (3) improved comfort, because the room-side drapery is more nearly at room temperature. Table 14 gives characteristics of individual draperies. For large areas, the IAC should be calculated in detail to determine the cooling load.
7. AIR LEAKAGE Fig. 26 Noise Reduction Coefficient Versus Openness Factor for Draperies device can be used to control brightness for the other shading devices and, when used alone, to reduce brightness while still allowing some view of the outdoors. View Modification. When the view is unattractive or distracting, draperies modify the view to some degree, depending on fabric weave and color (summarized in Table 15), but the fenestration product remains as an effective connection to the outdoors. Sound Control. Indoor shading devices, particularly draperies, can absorb some sounds originating in the room but have little or no effect in preventing outdoor sounds from entering. For excessive internally generated sound, the usual remedy is to apply acoustical treatment to the ceiling and other room surfaces. Although these materials can be effective in controlling sound, they are often located on the two horizontal surfaces (ceiling and floor) and leave the opposing vertical surfaces of glazing and bare wall to reflect sound. The noise reduction coefficient (NRC = average absorptance coefficient at four frequencies) for venetian blinds is about 0.10,
Infiltration Through Fenestration Air infiltration through fenestration products affects occupant comfort and energy consumption. Infiltration is the uncontrolled inward leakage of air caused by pressure effects of wind or differences in air density, such as stack effect. Infiltration should not be confused with ventilation. Although fenestration products can be operated to intentionally provide natural ventilation and increase comfort, infiltration should be reasonably minimized to avoid unpleasant accompanying problems. If additional air is required, controlled ventilation is preferable to infiltration. Mechanical ventilation provides air in a comfortable manner and when desired. For infiltration, however, peak supply is more likely to occur as an uncomfortable draft and when least desired, such as during a storm or the coldest weather. ASHRAE/IES Standard 90.1, ASHRAE’s energy standard for all buildings other than low-rise residential buildings, establishes an air leakage maximum of 0.3 L/s per square metre for curtain wall and storefront fenestration and 1.0 L/s per square metre of gross fenestration product area for most other products (5.0 L/s per square metre for glazed swinging entrance doors and revolving
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15.54
2017 ASHRAE Handbook—Fundamentals (SI)
doors, 2.0 L/s per square metre for nonswinging opaque doors, and 1.5 L/s per square metre for unit skylights). This air leakage is as determined in accordance with NFRC Technical Document 400 (NFRC 2014h) and ASTM Standard E283 and allows direct comparison of all fenestration products: operable and fixed, windows and doors. Most manufactured fenestration products achieve these reasonable standards of maximum air infiltration. However, products that do not completely seal, such as jalousie windows or doors, are not likely to do so and are most appropriate for installation in unconditioned spaces. For products achieving this infiltration standard, energy consumption caused by infiltration is likely to be significantly less than energy associated with U-factor and solar heat gain coefficient. Also, although overall air infiltration is a significant component in determining a building’s heating and cooling loads, infiltration through fenestration products meeting the standard is generally likely to be a small portion of that total.
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Indoor Air Movement Because supply air grilles are frequently located directly below fenestration products, air sweeps the indoor glazing surface. Heated supply air should be directed away from the glazing to prevent large temperature differences between the center and edges of the glazing. These thermal effects must be considered, particularly when annealed glass is used and air is forced over the glass surface during the heating season. Direct flow of heated air over the glass surface can increase the heat transfer coefficient and temperature difference, causing a substantial increase in heat loss, as well as leading to thermally induced stress and risk of glass breakage. Systems designed predominantly for cooling lower the glazing temperature and rapidly pick up the cooling load. Both tend to improve comfort conditions. However, the air-conditioned space has an increased net heat gain caused by increases in (1) solar heat gain coefficient (SHGC) caused by delivery of more of the absorbed heat to the indoor space, (2) fenestration U-factor because of the greater convection effect at the indoor surface, and (3) air-to-air temperature difference because supply air rather than room air is in contact with the indoor glazing surface. The principal increase in heat gain with clear glazing is the result of increased U-factor and air-to-air temperature difference.
8. 8.1
DAYLIGHTING DAYLIGHT PREDICTION
Daylighting is the illumination of building interiors with sunlight and sky light and is known to affect visual performance, lighting quality, health, human performance, and energy efficiency. In many European countries with predominantly cloudy skies, codes regulate minimum window size, minimum daylight factor, and window position to provide views to all occupants and to create a minimum indoor brightness level. Daylighting also provides back-up indoor illumination in the event of power outages. Daylighting may have some positive or negative health effects on the skin, eyes, hormone secretion, and mood. Its intensity, spectral content, and diurnal and temporal variation may be used to combat jet lag, sick building syndrome, and other health problems. In terms of energy efficiency, daylighting can provide substantial whole-building energy reductions in nonresidential buildings through the use of electric lighting controls. Daylight admission can displace the need for electric lighting at the perimeter zone with vertical windows (sidelighting) and at the core zone with skylights (toplighting) and special core sunlighting systems designed to bring sunlight to core spaces and distribute it as necessary, with good lighting quality. Lighting and its associated cooling energy use constitute 30 to 40% of a nonresidential building’s energy use. Energy
use reductions can be achieved, perhaps less reliably, in residential buildings with manual or automated switching of electric lights on and off to match space occupancy. For internal-load-dominated buildings, daylight admission must be balanced against solar heat admission to achieve optimum energy efficiency. Because heat gains through the building envelope from solar radiation typically define peak load conditions, daylighting is also a very effective method of decreasing peak demand. Daylighting can not only decrease annual operating costs through energy efficiency, but may also reduce capital cost by mechanical system downsizing. For daylighting designs using direct-beam sunlight entry, take care to avoid overheating and glare. Such problems can be avoided by carefully controlling or eliminating direct beam entry through orientation and shading of daylighting apertures, optical management components, and other architectural features. For conventional sidelit nonresidential buildings, three basic relationships for daylight optimization are given as functions of (1) glazing properties and (2) fenestration area as a percent of the gross wall area: 1. Annual cooling energy use (including fan energy use) increases linearly with solar radiation admission, as indicated by the product of SHGC and WWR, but is affected (nonlinearly) by decreases in electric lighting heat gains. 2. Annual lighting energy use decreases exponentially/asymptotically with daylight admission, as indicated by the product of Tv and WWR. 3. Annual heating energy use (including fan energy use) increases linearly with decreased lighting heat gains. Figure 27 shows the first two relationships for a prototypical nonresidential building. A similar relationship can be demonstrated with skylights. Different shading properties and control would affect electricity use differently. The fenestration design that achieves an optimum balance between daylight admission and solar rejection can be determined by iterative calculations where the glazing area and/or glazing solaroptical properties are varied parametrically (Tzempelikos and Athienitis 2005). For each case, the following general steps should be taken for each hour over a year: 1. Indoor Daylight Illuminance. Determine the building characteristics, configuration, outdoor design conditions, and operating schedules as described in Chapter 18. These include building orientation, outdoor obstructions, ground reflectance, etc. Determine the depth from the window wall for each electric lighting zone. Typical sidelighting windows can effectively daylight the perimeter zone to a depth of 1.5 times the head height of the window. In private offices, one dimming zone is typically cost effective, whereas in open-plan offices, two zones are cost effective. Select a typical task location in each of the lighting zones. Determine indoor daylight illuminance from all window and skylight sources at these locations. Indoor illuminance may be determined using computer simulation tools or physical scale models. Comprehensive explanations of simple and computer-based tools are available (IEA 1999). The majority of these tools can model simple box geometry with noncomplex fenestration systems. Some advanced simulation tools, such as Radiance (Ward 1990) and Adeline (Erhorn and Dirksmöller 2000), can model complex geometry and fenestration systems with adequate bidirectional solar-optical data, but this capability is not routine. 2. Lighting Energy Use. Determine the type of lamps, ballasts, and control system to be used in the perimeter zones. Determine whether the lamp can be dimmed or switched. For example, LED and fluorescent lamps can be dimmed, but metal halides cannot be switched or dimmed. Cold, outdoor applications of some lamps may prevent switching. For electronic dimming ballasts, obtain dimming power and light output characteristics. Obtain control
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Fenestration
15.55 Instantaneous lighting heat gains qel , described in Chapter 18, must be multiplied by the power reduction factor Fdaylight . Mechanical loads and energy use may then be determined as described in Chapter 18. Many studies have investigated the magnitude of change in heating and cooling energy use associated with reductions of lighting energy use in nonresidential buildings, as will be realized with daylighting controls. In a DOE-2.1E simulation study (Sezgen and Koomey 1998), the greatest savings were generated in hospitals, large offices, and large hotels; for every $1.00 saved through lighting energy efficiency, additional savings as a result of reduced HVAC were $0.26, $0.16, and $0.14, respectively. These results emphasize the need to include HVAC effects when assessing the effects of daylighting. Simplified design tools are available to conduct such parametric runs for preliminary analysis. Skylighting tools based on regressions using DOE-2 data or simplified DOE-2 procedures are also available [American Architectural Manufacturers Association (AAMA) 1987; Heschong et al. 1998]. More comprehensive building energy prediction tools combined with daylighting algorithms, such as DOE-2.1E (Winkelmann 1983), implement hour-by-hour calculations using existing weather data and enable evaluation of glare, visual comfort, and quality of light as well (Hitchcock et al. 2008; Lee and Selkowitz 1995; Ochoa and Capeluto 2009; Tzempelikos and Athienitis 2007).
Fig. 27 Window-to-Wall Ratio Versus Annual Electricity Use in kWh/(m2 ·floor·year) specifications to determine how the system will respond to available light; dead-band ranges, response times, and commissioning affect the sensitivity and accuracy of the system. The type of switching (on/off, bilevel, multilevel, and continuous dimming controls) are dictated by both the type of lamp and space use. Determine the task illuminance design set point for each zone. Determine the percentage electric lighting power reduction Fdaylight that will result with automatic daylight controls, and apply to the installed wattage. Simplified methods for calculating lighting power reductions based on task illuminance levels are given in Robbins (1986). More sophisticated programs (Choi and Mistrick 1999) model commercially available photosensor dimming control systems (typically located in the ceiling above the work plane task) more rigorously; the spectral and bidirectional response of the photosensor to incident flux is used to determine voltage output, which is then used by the ballast controller algorithm to determine the lighting power reduction. Response delays and commissioning set points further affect this predicted output. Lights may also be switched manually, but there are very few modeling prediction tools for manual switching (Reinhart 2004). Field tests (Jennings et al. 1999) indicate that with bilevel switching, 45% of the lighting zone-hours were at less than full-power lighting, with 28% at only one-third of full lighting output levels. Manual switching occurred less in public spaces. Occupancy and other types of switching may occur as well and should be accounted for as a confounding effect with any daylighting controls. 3. Mechanical Energy Use. Determine mechanical energy use caused by fenestration loads and reduced electric lighting heat gains. Fenestration heat gains and losses may be computed using the section on Determining Fenestration Energy Flow.
In the United States, a general rule has been that the fenestration area should be at least 20% of the floor area. In Europe, a similar rule was based on a minimum illumination value on the normal work plane from a standard overcast sky condition. In general, it is more energy efficient to use larger fenestration areas to elevate indoor surface brightness as a glare reduction strategy than to increase indoor electric lighting levels. As fenestration area increases, indoor brightness increases while fenestration brightness remains the same. Of course, mitigating considerations include increased cost and heat transfer with larger fenestration. The latter problem can be mitigated with multiple-pane fenestration and special coatings to reduce solar gain without serious loss of light transmission, as discussed in the section on Selecting Fenestration. Orientation and shading can also be effective at mitigating glare and overheating problems. The secondary visual benefit of fenestration is the amount and quality of light it produces in the work environment. One general rule determined the need for auxiliary electric light by assuming that daylight was adequate for a depth of two and one-half times the height of the fenestration product into the room based on a normal sill height. To prevent excessive glare, all fenestration should have sun controls. Variable and removable controls are often more effective in daylighting than fixed controls. For more accurate evaluation of daylight distribution in a space, several prediction tools, such as the Recommended Practice of Daylighting (IES 1999), are available. This practice shows a simple way of calculating the daylight distribution on the work plane from windows and skylights with and without controls. Many other daylight prediction tools calculate illuminance from radiant flux transfer or ray tracing. The Illuminating Engineering Society’s Daylighting Committee, Subcommittee on Daylighting Metrics, identified various instantaneous and annual measures of daylighting performance in building spaces. New techniques and improved computer tools are being developed to help calculate these measures, and various building code bodies and energy incentive programs are expected to specify daylighting performance metrics for compliance with these programs. Any or all of the various daylight prediction tools can be used to compare the relative value of daylight distribution from alternative fenestration systems, but ultimately the designer must evaluate costs and benefits to choose between alternative designs. This may be based on energy use or, more properly, on overall costs and benefits
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15.56
2017 ASHRAE Handbook—Fundamentals (SI)
to the client. Also, the risk of total loss of productivity from an electric brown-out in a space with no natural ventilation or daylight may be as important as the benefits of many energy-saving schemes.
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8.2
LIGHT TRANSMITTANCE AND DAYLIGHT USE
When daylight is to be the primary lighting system, the minimum expected daylight in the building must be calculated for the building performance cycle and integrated into lighting calculations. IES (1999) gives daylight design and calculation procedures. In some glazing applications, such as artists’ studios and showrooms, maximum transmittance may be required for adequate daylighting, and care is needed to maintain good color rendering when glazing tints and coatings are used. Regular clear glass, produced by float, plate, or sheet process, may be the logical choice. When daylight is a supplementary light source, electric lighting can be designed independently of the daylight system. However, adequate switching must be included in the electric distribution to substitute available daylight for electric lighting by automatic or prescribed manual control whenever practical. Photosensitive controls automatically adjust shading devices to provide uniform illumination and reduce energy consumption. Manual control is less effective. Buildings with large areas of glazing usually have glazing units with clear, tinted, or reflective coatings. Tinted and reflecting units reduce the brightness contrast between fenestration products and other room surfaces and provide a relatively glare-free environment for most daylight conditions. Table 10 lists typical approximate solar energy transmittances and daylight transmittances for various glass types. Manufacturers’ literature has more appropriate type-specific values. The color of glazing chosen for a building depends largely on where and how it is used. For commercial building lobbies, showroom fenestration products, and other areas where maximum visibility from outdoors to indoors is required, regular clear glass is generally best. Clear glass with a low-e coating is also suitable for these locations, including for retail storefronts, because it only decreases light transmittance by about 10%. For other glazing areas, tinted or coated glass may best complement the indoor colors. Bronze, gray, and reflective-film glasses also give some privacy to building occupants during daylight hours. Patterned, fritted, etched, or sandblasted glass that diffuses lighting is available. In warm climates, tinted outer glazing in an insulated double-pane system can have solar heat gain rejection benefits, while providing good colorrendering illumination of the interior without apparent color. The primary purpose of a fenestration product is not just to save energy but to provide a view of the outdoors and bring useful daylight indoors. The light-transmitting properties of fenestration systems are therefore of great importance. It is conceivable that one could design a fenestration product with excellent solar heat gain performance for hot climates (meaning a very low solar heat gain coefficient) but very poor view and daylight illumination performance. If this problem is bad enough, it can cause occupants to turn on electric lights indoors during the daytime, which adds to the electric bill and possibly causes problems of thermal discomfort as well. The light-transmitting property of a fenestration product is called the visible transmittance Tv . It is similar to the solar-weighted solar transmittance, except that an additional weighting function is needed, in this case to account for the spectral response of the human eye. In most applications, it is important to have high visible transmittance. In cooler climates, good solar heat gain is also important for offsetting wintertime heating costs. In warmer climates, low solar heat gain is good for offsetting summertime cooling costs. In the latter situation, it is difficult to have both high visible transmittance and a low solar heat gain coefficient. Figures 28 and 29 show plots of visible transmittance versus SHGC for several glazing systems covering a range of spectral selectivities (McCluney 1996). The data
are for normal incidence and a single, ASTM standard solar spectral distribution. For daylighting design, a rule of thumb is to select a glazing unit having a visible transmittance 1.5 to 2.0 times greater than its solar heat gain coefficient. For maximum light with minimum solar gain, there are fenestration products available having a visible transmittance as high as 3.0 times the SHGC. For maximizing passive solar gain (e.g., on a south orientation in a cold climate in the northern hemisphere), select a glazing unit with a low-e coating whose visible transmittance is greater than its solar heat gain coefficient. Three different zones are delineated in Figure 29. In the neutral zone, it is possible to have colorless glazing systems (i.e., glazings with approximately uniform transmittance over the visible spectrum). Glazings in this zone can have some color, but this is not necessary. In the color zone, the only way to achieve higher visible transmittance for a given level of solar heat gain coefficient is by stripping off some of the red and blue wavelengths at the edges of
Fig. 28 Visible Transmittance Versus SHGC for Several Glazings with Different Spectral Selectivities
Fig. 29 Visible Transmittance Versus SHGC at Various Spectral Selectivities (McCluney 1996)
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Fenestration
15.57
Table 16 Spectral Selectivity of Several Glazings Glazing
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Reflective blue-green Film on clear glass Green tinted, medium Green low-e Sun-control low-e + green Super low-e + clear Super low-e + green
9.1
Tv
SHGC
LSG
0.33 0.19 0.75 0.71 0.36 0.71 0.60
0.38 0.22 0.69 0.49 0.23 0.40 0.30
0.87 0.86 1.09 1.45 1.56 1.77 2.00
the human spectral response function with a spectrally selective glazing transmittance, imparting color to the transmitted radiation (or by otherwise altering the spectral transmittance and hence the color over the visible portion of the spectrum). In the forbidden zone, no combination of visible transmittance and solar heat gain coefficient is possible for normal incidence and for the solar spectral distribution used. (Changing the solar spectral distribution used to calculate Tv and SHGC shifts the transition curves somewhat. A low solar altitude angle, direct-beam spectrum will move the curves to the left on the plot in Figure 29.) Glazings that transmit more solar radiant heat than light cluster on the lower portion of the plot. The Tv versus SHGC chart can be a useful tool for illustrating the degree of spectral selectivity attained by a glazing system. These concepts lead to an index of spectral selectivity that can be useful. It is called the light-to-solar-gain ratio (LSG) (McCluney and Jindra 2001), defined as Tv LSG = ---------------SHGC
(41)
Some characteristic values for Tv , SHGC, and LSG are given in Table 16 for several different glazings, using the ASTM standard spectral distribution at normal incidence to calculate the values. The LSG can be useful in spotting errors in calculating the SHGC. Values of SHGC that lie outside reasonable ranges can be spotted fairly quickly and used to identify possible problems in calculations or measurements. In general, it is very difficult and therefore unlikely to have a useful glazing system for buildings with an LSG value greater than 3.0. Values below 0.3 should be particularly suspect, because they indicate a glazing that transmits considerably more heat than light and would be unlikely candidates for general use. Generally, a high value of LSG is desired for residential buildings in hot climates, to maximize daylight admission with minimal solar heat gain. This is also true for internal-load-dominated nonresidential buildings in many climates, because solar gain rejection is often desired for such buildings, even in cool or cold climates. Provided that the fenestration product has a low-e coating, an LSG value somewhat below 1.0 is appropriate in cold climates for residential buildings and nonresidential buildings without strong internal cooling loads.
9.
SELECTING FENESTRATION
Because fenestration systems provide so many functions, and because environmental conditions and user needs vary widely, it is difficult to make a completely optimal selection of a fenestration system. Considering aesthetics and cost, visual and comfort performance, annual energy costs, peak-load consequences, and acoustic characteristics, the choice is seldom optimal. The HVAC system designer, fortunately, has a more restricted range of interests, mainly dealing with the energy consequences of a particular fenestration selection. This section therefore focuses on fenestration energy performance determination.
ANNUAL ENERGY PERFORMANCE
Instantaneous energy performance indices (e.g., U-factor, solar heat gain coefficient, visible transmittance, air leakage) are typically used to compare fenestration systems under a fixed set of conditions. However, the absolute and relative effect of these indices on a building’s heating and cooling load can fluctuate as environmental conditions change. Further, internal loads from electric lighting fluctuate as automatic controls dim or turn off electric lighting in response to available daylight. As a result, these indices alone are not good indicators of the annual energy performance attributable to the fenestration. Such energy performance is difficult to quantify in and of itself because of numerous dynamic responses between the fenestration system and the total environment in which it is installed. The four basic mechanisms of fenestration energy performance (thermal transfer, solar heat gains, air leakage, and daylighting) should all be taken into account but are not independent of many other parameters that influence performance. As a result, the annual energy performance of fenestration systems can be accurately determined only when many variables are considered. Building type and orientation, climate (weather, temperature, wind speed), microclimate (shading from adjacent buildings, trees, terrain), occupant usage patterns, and certain HVAC parameters can significantly affect the annual energy effects of fenestration systems. For these reasons, the most effective means of establishing fenestration annual energy performance is through detailed, dynamic, hourly computer simulations for the specific building and climate of interest. Because instantaneous performance of fenestration often varies by differing magnitudes as climatic conditions change, the most accurate simulation results are obtained when these variances are accounted for in a building energy simulation computer program. After constructing the simulation model following the procedures defined in Chapters 17 and 18 (for residential and commercial construction, respectively), specific changes to the fenestration system can be modeled, and the annual energy performance changes attributable to fenestration can be quantified. These analytical techniques do not consider issues of performance durability for the various instantaneous indices and should only be used as an initial annual energy performance indicator (Mathis and Garries 1995).
Simplified Techniques for Rough Estimates of Fenestration Annual Energy Performance Although dynamic hourly modeling is certainly the most accurate technique for determining fenestration annual energy performance, and software for hourly modeling now runs on most desktop and laptop computers, it is not always available to decision makers and end users of fenestration products, simply because users may not want to make the necessary inputs to run the analysis. Under these circumstances, it may be useful to assess the relative importance of, or balance the trade-off between, the known instantaneous performance indices of U-factor, SHGC, Tv , and air leakage, for any given fenestration system when considering heating, cooling, and lighting loads for many different building types and climates. Huang et al. (1999) and Mitchell et al. (1999, 2012) describe personal computer programs to run this simplified analysis for residential windows. Broad generalizations can be made for some classifications of building types and climates. For instance, with large commercial buildings, which require substantial cooling energy use during daytime occupied hours because of high internal loads, significant thermal mass, or high orientation dependency, low SHGC is important to reduce the cooling load. Also, an evaluation of commercial fenestration annual energy use can take into account the tradeoff between electric lighting and the natural daylighting benefits associated with a particular fenestration system. However, low U-factor is also important because commercial buildings have bimodal operation: they can have significant heating energy
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15.58 consumption during morning warm-up, which occurs during unoccupied predawn hours, before people arrive and lights and equipment are turned on, and before any passive solar gain. In low-rise, detached residential buildings, electric lighting loads are typically very small in comparison to the heating and cooling loads because of high envelope-dependent energy use, egress requirements, and occupant usage patterns. The slight energy benefits of daylighting may be neglected altogether when using light-emitting diode (LED) lighting, which reduces residential light loads substantially. Despite these generalizations, the problem still exists of balancing and assessing the effect of each of the remaining parameters to establish seasonal or annual energy performance for cases in which detailed computer modeling is not performed. Furthermore, in all building types, the fact that energy savings afforded by daylighting are small does not negate the desirability of well-controlled natural light: daylighting without glare or excessive contrast is an important amenity. Developing simplified annual energy performance indices for fenestration typically involves using instantaneous fenestration performance indices to quantify building- and climate-independent scalars of annual or seasonal energy performance for rating purposes. Many of these performance indices can be relatively independent of building type, climate, distribution of products, orientation, and other items needed for hourly dynamic building energy analyses. These normalized, scalar-based approaches are also limited in accuracy for the same reasons. A further limitation of the simplified techniques is that they do not have broad applicability to varied building types (e.g., commercial versus residential buildings). The usefulness of these scalar-based approaches can be increased when limiting the comparison to a single building type. Currently, simplified techniques for characterizing fenestration annual energy performance are applicable only to fenestration systems for detached residential buildings and are not appropriate for use with multifamily residential or commercial building fenestration systems. Where a building function requires controlling outdoor air through a door opening, evaluation of the fenestration annual energy performance involving the resulting air exchange cannot be based on instantaneous indices, but must be based on an annualized process. Evaluate an automated door product designed to limit dooropening, door-open, and door-closing times for its air exchange, air leakage, and thermal transmittance properties, and compare its performance to that of a conventionally operating automated insulated sectional door complying with U-factor and air leakage requirements. Because a door product limiting air exchange is specified in terms of the number of opening and closing cycles per day averaged over a one-year period, not only can its annual energy performance be assessed, but its heating and cooling savings can also be estimated with respect to the sectional door, assuming the same number of average cycles per day and a similar door-open time.
2017 ASHRAE Handbook—Fundamentals (SI) In Australia, the Window Energy Rating Scheme (WERS; www .wers.net) publishes simplified, consumer-friendly AEP ratings for heating and cooling performance of residential windows. It displays regression-based AEP rankings as stars, on a scale of 0 to 10; one rating is for cooling performance, and another for heating. Ten stars for either heating and cooling corresponds to the “perfect” window for that purpose. As inputs, WERS uses U-factor and SHGC, determined using procedures licensed from the NFRC and the Australian Fenestration Rating Council (AFRC; www.afrc.org.au).
9.2
CONDENSATION RESISTANCE
Water vapor condenses in a film on fenestration surfaces that are at temperatures below the dew-point temperature of the indoor air. If the surface temperature is below freezing, frost forms. Sometimes, condensation occurs first, and ice from the condensed water forms when temperatures drop below freezing. Condensation frequently occurs on single glazing and on aluminum frames without a thermal break. The edge seal creates a thermal bridge at the perimeter of the glazing unit. Circulation of fill gas caused by temperature differences in the glazing unit cavity contributes to the condensation problem at the bottom of the indoor glazing (Curcija and Goss 1994, 1995; Wright 1996b; Wright and Sullivan 1995a, 1995b). In winter, fill gas near the indoor glazing is warmed and flows up, while gas near the outdoor glazing is cooled and flows down. The descending gas becomes progressively colder until it reaches the bottom of the cavity. There, the gas turns and flows to the indoor glazing, resulting in higher heat transfer rates at the bottom. Thus, the bottom edge of the indoor glazing is cooled both by edge-seal conduction and by fillgas convection. The combined effect of these two heat transfer mechanisms is shown in Figure 30. The surface isotherms show a wider band of cold glazing at the bottom of the window. Typical condensation patterns match these isotherms. The vertical indoor surface temperature profile also shows the effect of edge-seal conduction and that the minimum indoor surface temperature is near the bottom edge of the glazing.
Simplified Residential Annual Energy Performance Ratings Annual energy performance (AEP) ratings can provide a simple means of product comparisons for consumers. These ratings have been derived with many assumptions, usually to suit local climatic conditions. The Canadian Standards Association (CSA Standard A440.2) developed a simplified energy rating applicable to residential heating in the Canadian climate, which was adopted in the 1995 National Energy Code for Houses. The standard also provides for specific energy ratings to compare products by orientation and climate. In the United States, where heating and cooling are both significant, NFRC is developing a rating system that includes both effects (Arasteh et al. 2000; Crooks et al. 1995). NFRC’s (2014i) Technical Document 901 provides guidance on how fenestration affects heating and cooling energy consumption in single-family residences.
Fig. 30 Temperature Distribution on Indoor Surfaces of Glazing Unit
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Fenestration Condensation on fenestration and surrounding structures can cause extensive structural, aesthetic, and health problems. Specific examples include peeling of paint, rotting of wood, saturation of insulation, and mold growth. Ice can render doors and windows inoperable and prevent egress during an emergency. Energy-efficient housing has been accompanied by reduced ventilation. The resulting increase in indoor humidity has contributed to the condensation problem. However, the solution does not lie in the reduction of humidity levels to a minimum. Relative humidity below 20% and above 70% can increase health risks and reduce comfort. Generally, a minimum of 30% rh should be maintained, and 40% to 50% is more desirable (Sterling et al. 1985). Consequently, a better solution is to improve the fenestration product so interior surface temperatures are warmer and condensation does not occur or is minimized. Minimum indoor surface temperatures can be quantified in a variety of ways. De Abreu et al. (1996), Elmahdy (1996), Griffith et al. (1996), Sullivan et al. (1996), and Zhao et al. (1996) demonstrated good agreement between detailed two-dimensional numerical simulation and surface temperature measurements using thermographs. Curcija et al. (1996) and Wright and Sullivan (1995c) developed simplified simulation models to predict condensation resistance. Center-glass and bottom-edge surface temperatures that can be expected for two different glazing systems exposed to a range of outdoor temperature are shown in Figure 31. Both glazing systems include insulating foam edge seals. High-performance glazing systems (e.g., low-e/argon and insulated spacers) allow significantly higher indoor humidity levels. Current measures of condensation resistance of a fenestration system are the condensation resistance (CR) as defined by NFRC
15.59 500 (2014j) and its user guide (2014h), the condensation resistance factor (CRF) as defined by AAMA (1988), or the temperature index (I), as defined in CSA Standards A440 and A440.1. Note that the temperature index method in CSA Standard A440 stipulates that the test is performed on the fenestration with all the cracks not sealed. This represents a major difference between the CSA Standard A440 method and the AAMA and NFRC methods. There are some merits of leaving cracks unsealed during testing for condensation resistance. In particular, any inherent deficiencies in fenestration design may result in uncontrolled air leakage through the fenestration. This air leakage could not be detected or dealt with in the simulation models, and can only be seen in the results of the determined temperature index. On the other hand, there is some financial benefit to the window manufacturer in testing the fenestration for condensation resistance with cracks sealed, because one test can determine R-value and condensation resistance. Research shows that air leakage does affect the temperature index (measure of condensation resistance as determined by CSA Standard A440). Elmahdy (2001, 2003) showed that sealing cracks during testing artificially improves the temperature index, compared to the results of the same window tested with cracks unsealed. Condensation resistance is a measure of condensation potential, based on both area and temperature weighting and expressed as a minimum of center-of-glazing, edge-of-glazing, and frame CRs. The novelty of this index is that it is determined using computer simulation tools unless the overall thermal performance cannot be validated with testing. If thermal performance cannot be validated, a testing option for determining CR is used. The other two standards define the values by a single dimensionless number as t – tc CRF or I = -------------th – tc
Fig. 31
Minimum Indoor Surface Temperatures Before Condensation Occurs
(42)
where th and tc are the warm- and cold-side temperatures, respectively. Figure 32 can be used to determine the acceptable range of CRF/I for a specific climatic zone. The two standards differ in the methods used to determine temperature. The CSA test procedure is based on thermocouple measurements at the coldest location on the frame plus three locations on the glazing, each 10 mm above the bottom sightline. The AAMA procedure specifies two separate factors: one for the frame (CRFF), which uses weighted frame temperature obtained from surface temperature measurements at predetermined and roving locations on the frame, and one for the glazing unit (CRFG), which uses the average of six temperatures measured at predetermined locations near the top, middle, and bottom of the glazed area. Indoor details can significantly alter the potential for condensation on fenestration surfaces. Items such as venetian blinds, roll blinds, insect screens, and drapes increase thermal resistance between the indoor space and fenestration and lower the temperature of the fenestration surfaces. These fenestration treatments do not prevent migration of moisture, so they can cause increased condensation. Figure 33 shows different situations that affect the potential for condensation. Note that window reveal plays an important role. If the window is placed near the outside of the wall, the increase in the outdoor film coefficient and decrease in the indoor film coefficient cause colder window surfaces. This effect is more pronounced near the corners of the recess where the indoor film coefficient is locally suppressed because air movement is restricted. Also, blinds should be placed at least 100 mm from the plane of the wall to allow some natural convection between the window and the blind. Air leakage, especially in operable sections of fenestration, is another important cause of low surface temperature. Leakage near
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edge-of-glass sections can further increase the potential for condensation. However, the drier outdoor air decreases relative humidity near leakage sites and, in some cases, offsets the undesirable effect of lower surface temperatures. The net effect of air leakage cannot readily be determined experimentally or with simulation.
9.3
OCCUPANT COMFORT AND ACCEPTANCE
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Human thermal comfort is an immediate sensation that reflects building occupants’ perceived response to many physical factors. Unlike much building design that is based primarily on long-term energy and economic considerations, comfort-related design focuses on, and must take heed of, short-term responses of the body’s physiology to its surroundings.
Fenestration influences thermal comfort through a combination of three mechanisms: long-wave radiation exchange, absorption of solar radiation, and convective draft effects (Figure 34). An understanding of these phenomena is important to help designers evaluate the benefits of improved fenestration and create comfortable buildings. Although it is well understood that high-performance fenestration can reduce building energy consumption, a better understanding of their effect on comfort might lead to further savings. For example, Hawthorne and Reilly (2000) suggest that significant energy consumption is caused by the standard practice of using perimeter duct distribution in houses to mitigate potential discomfort caused by fenestration. They found that perimeter heating is often not necessary when high-performance fenestration is installed and that heating energy savings of 10 to 15% could result from installing a simpler, less expensive duct system. Better windows can allow thermostat settings to be lowered with no loss
Fig. 32 Minimum Condensation Resistance Requirements (th = 20°C)
Fig. 33 Location of Fenestration Product Reveals and Blinds/ Drapes and Their Effect on Condensation Resistance
Fig. 34 Fenestration Effects on Thermal Comfort: Long-Wave Radiation, Solar Radiation, Convective Draft
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of comfort. Another simulation study (Lyons et al. 2000) examined the relative magnitudes of a residential window’s physical influences under a wide variety of winter and summer climates, glazing parameters, and clothing levels. They found that
Table 17 Sound Transmittance Loss for Various Types of Glass
• Long-wave, thermal radiation influences of the window dominate unless direct sun strikes the occupant • Direct solar load has a major influence on perceptions of comfort • For most residential-size windows, draft effects are generally small
3 mm double-strength sheet glass 6 mm plate or float glass 13 mm plate glass 19 mm plate glass 25 mm plate glass 6 mm laminated glass (11 mm plastic interlayer) 25 mm double glass 13 mm laminated glass (11 mm plastic interlayer) Multiple-glazing unit, 150 mm air space, 6 mm plate or float glass
With all but highly insulating fenestration, the indoor surface temperature of the fenestration is heavily influenced by outdoor conditions, and this temperature can significantly affect radiant heat exchange between an occupant and the environment. If this heat exchange moves outside the acceptable range, discomfort results. Mean radiant temperature (MRT) is commonly used to simplify the characterization of the radiant environment. On a cold day, the indoor surface temperature can easily drop below –9°C for a clear single-pane window and below 4°C for a clear, double-pane window. If the occupant is sitting sufficiently near the window, MRT could drop to 13°C for the single-pane case and 17°C for the doublepane case. Based on ASHRAE Standard 55, even use of the clear double-pane window could result in discomfort. [This example assumes an outdoor air temperature of –18°C, indoor air temperature of 22°C, nonwindow surface temperatures of 22°C, occupant/ window view factor of 0.3, 0.9 clo (standard winter indoor clothing), and activity level of 1 met.] In addition to the MRT effect, a cold indoor glazing surface can induce a downward draft that increases air movement, contributing to further discomfort. If direct solar radiation strikes the glazing or occupant, the situation is much more complex. Conversely, for a double-pane window with a low-e coating and a triple-pane window with two low-e coatings, the net results would be MRTs of 19°C and 21°C, respectively. Both of these would satisfy the ASHRAE Standard 55 criteria for comfort. In winter, the warming effect of sunlight on skin and clothing is often welcome, depending on the compounding effect of other factors such as air temperature. Fenestration also absorbs and transmits a significant amount of solar radiation. Because of such absorption, solar-heated fenestration may improve MRT for a nearby person. The premise of passive solar design is that occupants will welcome, or at least tolerate, solar gain in exchange for savings on heating energy. However, it is desirable that the onset of discomfort be able to be predicted; otherwise, the energy-saving design may be defeated if occupants draw shades to prevent overheating. In summer, solar-heated glazing may become uncomfortably hot and, in commercial premises, actually devalue rented space near windows. The indoor surface of body-tinted, heat-absorbing glass can routinely reach temperatures above 50°C in summer conditions, raising MRT by as much as 8 K. This can be ameliorated by adding a second glazing on the indoors. Transmitted radiation often causes discomfort if it falls directly on the occupant. A person sitting near a window in direct solar radiation can experience heat gain equivalent to a 11 K rise in MRT (Arens et al. 1986). Similarly, in residential applications, the perceived need for solar control is affected both by the contribution of window surfaces to MRT and by overheating from direct solar load. Advances in fenestration technology, especially high-performance glazings, mean that the designer has a choice of potential glazing systems. On the basis of annual energy performance for heating, cooling, and lighting, these alternatives may give similar outcomes. However, because they represent different combinations of U-factor, SHGC, and indoor glazing surface temperature, their comfort outcomes may differ considerably. Research continues to develop tools to help designers evaluate such difficult trade-offs. In the meantime, several general rules of thumb may be followed: • In heating-dominated climates, fenestration with the lowest Ufactor tends to give the best comfort outcomes. However, there is
Type of Glass
Sound Transmittance Loss, dB 24 27 32 35 36 30 32 34 40
likely to be a trade-off between the twin goals of maximizing instantaneous comfort and minimizing annual energy consumption. • In cooling-dominated climates or for orientations where cooling loads are of concern, fenestration with the lowest rise in surface temperature for a given SHGC tends to give the best comfort outcomes.
Sound Reduction Proper acoustical treatment of outdoor walls can decrease noise levels in certain areas. The airtightness of a wall is the primary factor to consider in reducing sound transmission from outdoors. Once walls and fenestration products are tight, the choice of glazing and draperies becomes important. Draperies do not prevent sound from coming through the fenestration; they act as an absorber for sound that does penetrate. Table 17 lists average sound transmission losses for various types of glass. These averages apply for the frequency range of 125 to 4000 Hz and were determined by tests based on ASTM Standard E90.
Strength and Safety In addition to its thermal, visual, and aesthetic functions, glazing for building exteriors must also perform well structurally. Wind loads are specified in most building codes, and these requirements may be adequate for many structures. However, detailed wind tunnel tests should be run for tall or unusually shaped buildings and for buildings where the surroundings create unusual wind patterns. The strength of annealed, heat-strengthened, tempered, laminated, and insulated glass is given in ASTM Standard E1300. Thermal expansion and contraction can break ordinary annealed glass. This expansion and contraction can be caused by solar radiation onto partly shaded glazing, by heat traps from drop ceilings and tight-fitting drapes, or by HVAC ducts incorrectly directed toward the glazing. High-performance tinted and reflective glasses with low-e coatings are usually more vulnerable to thermal stress breakage than clear glass. Heat treating (heat strengthening or fully tempering) the glass resists thermal stress breakage. Heat-strengthened glass, although not a safety glass, is usually preferred to tempered (safety) glass because it typically has less distortion and is much less likely to have spontaneous breakage, which can occur on very rare occasions in tempered glass. Consult the glass manufacturer or fabricator for information on thermal stress performance. Building codes may require glass in certain positions to perform with certain breakage characteristics, which can be satisfied by tempered, laminated, or wired glass. In this case, glass should meet Code of Federal Regulations 16CFR1201 or other appropriate breakage performance requirements.
Life-Cycle Costs Alternative building shells should be compared to ensure satisfactory energy use and total energy budget compliance, if required. ASHRAE Standards 90.1 and 90.2 should be used as a starting point. A life-cycle cost model should be developed for each system
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considered. See Chapter 37 of the 2015 ASHRAE Handbook— HVAC Applications.
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9.4
DURABILITY
Service life and long-term performance of fenestration systems depend on the durability of all the system’s components. Representative samples of glazing units are usually tested (for seal durability) according to test methods to ensure the integrity of the seal. Failure of glazing units is usually indicated by loss of adhesion of sealant to the glazing; as a result, fogging occurs inside the glazing cavity. For argon-filled units, seal failure means a loss of argon and, hence, degradation in the unit’s thermal characteristics. Extensive study of the durability of glazing units filled with argon gas (Elmahdy and Yusuf 1995) indicated that, under normal conditions, argon loss by diffusion through the sealant is very small. However, when cracks or pinholes exist in the sealant, most of the argon gas escapes, which implies that stringent quality control procedures are essential for production of durable glazing units. Degradation of organic materials and other chemical components in glazing units as a result of exposure to ultraviolet radiation is also a factor affecting durability and service life of fenestration systems. Low-e coatings on glass tend to enhance the appearance of chemical deposits on the glass surface. Also, inserting muntin bars in glazing cavities may result in excessive rates of unit failure during ultraviolet volatile (fogging) tests unless strict quality assurance processes are implemented. Current ASTM (United States) and CGSB (Canada) durability standards are being reviewed to reflect the emergence of new technologies in the fenestration industry. A 15-year correlation study of insulating glazing units products by the Insulating Glass Manufacturers Association (IGMA) found that long-term performance and durability of glazing units correlated well with the test level to which the unit’s construction had been manufactured under ASTM Standard E2190’s specification for sealed glazing units. Units showing the highest percentage of resistance to seal failure were those tested in conformance with the ASTM Standard E2190 class CBA standard. Units that did not qualify to the A level showed a definite correlation to a higher percentage of failure. Field correlation studies found that units glazed in compliance with IGMA recommendations perform for longer periods than units not constructed properly, having deficiencies in the glazing system, or not meeting ASTM requirements. Durability of fenestration systems also depends on durability of other system components, such as weatherstripping, gaskets, glazing tapes, air seals, and hardware. Wear of these elements with time and use may result in excessive air and water leakage, which affects overall performance and service life of the system. Excessive water leakage may cause damage to the fenestration product, especially the edge seal, as well as the wall section where the product is mounted. Excessive air leakage may lead to frost build-up and condensation on fenestration surfaces. Studies conducted at the National Research Council of Canada (Elmahdy 1995) and elsewhere (Patenaude 1995) showed that, when windows are tested at high pressure and temperature differentials, they experience air leakage rates exceeding those determined at 75 Pa and zero temperature differential (conditions used in rating window air leakage in U.S. and Canadian standards). In other studies (CANMET 1991, 1993), pressure and motion cycling on windows resulted in excessive degradation in almost all performance factors, particularly condensation resistance, ease of operation, and air and water leakage. To predict long-term performance, unit construction for glazing units should be tested and certified in accordance with ASTM Standard E2190 class CBA level and the requirements of IGMA or equivalent.
Durability may also affect long-term energy performance. Consequently, NFRC now requires third-party certification for sealedglass units.
9.5
SUPPLY AND EXHAUST AIRFLOW WINDOWS
Airflow windows allow air to flow between glass panes of multilayered glazing units, to improve the window assembly’s thermal performance. Exhaust air windows allow indoor air to flow between the inner two panes of a triple-glazed window. In the cooling season, this airflow helps reduce the cooling load by transferring heat to the flowing air and discharging it to the outdoors. During the heating season, heat loss through the outer pane of the window comes mostly from exhaust airflow, which helps reduce thermal transmission loss through the window. In addition, exhaust airflow helps maintain the inner pane surface temperature close to the indoor air temperature, thus improving the thermal comfort of occupants (Haddad and Elmahdy 1998, 1999). The supply air window allows outdoor air to flow between the outer two panes of a triple-glazed window and into the building. The airflow helps reduce the heating load when heat picked up by the flowing air finds its way back into the indoor space. In the cooling season, the supply air window may increase the cooling load when heat is picked up from the outer glass pane and delivered into the indoor space. Haddad and Elmahdy (1998, 1999) provide results of computer models comparing thermal performance of supply and exhaust airflow windows with conventional windows in various locations in North America.
9.6
CODES AND STANDARDS
National Fenestration Rating Council (NFRC) The National Fenestration Rating Council (NFRC) was formed in 1989 to respond to a need for fair, accurate, and credible ratings for fenestration products. NFRC has developed rating procedures for U-factor [NFRC (2014a) Technical Document 100], solar heat gain coefficient and visible transmittance [NFRC (2014d) Technical Document 200], optical properties [NFRC (2014c) Technical Document 300], air leakage [NFRC (2014f) Technical Document 400], and condensation resistance [NFRC (2014j) Technical Document 500]. To provide certified ratings, manufacturers follow the requirements in the NFRC Product Certification Program (PCP), which involves working with laboratories accredited to the NFRC Laboratory Accreditation Program (LAP), and independent certification and inspection agencies accredited through the NFRC Certification Agency Program (CAP). NFRC (2014a) Technical Document 100 was the first NFRC rating procedure approved and thus the first NFRC procedure adopted into energy codes in the United States. It requires using a combination of state-of-the-art computer simulations and improved thermal testing to determine U-factors for the whole product. The next step is product certification. NFRC has a series of checks and balances to ensure that the rating system is accurately and uniformly used. Products and their ratings are authorized for certification by an NFRC-licensed independent certification and inspection agency (IA). Finally, two labels are required: the temporary label, which contains the product ratings, and a permanent label, which allows tracking back to the IA and information in the NFRC Product Directory. In addition to informing the buyer, the temporary label provides the building inspector with information necessary to verify energy code compliance. The permanent label provides access to energy rating information for a future owner, property manager, building inspector, lending agency, or building energy rating organization.
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Fenestration This process has noteworthy features that make it superior to previous fenestration energy rating systems and correct past problems:
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• The procedures provide a means for manufacturers to take credit for all the nuances and refinement in their product design and a common basis for others to compare product claims. • Involvement of independent laboratories and the IA provides architects, engineers, designers, contractors, consumers, building officials, and utility representatives with greater confidence that the information is unbiased. • Requiring simulation and testing provides an automatic check on accuracy. This also remedies a shortcoming of previous energy code requirements that relied on testing alone, which allowed manufacturers to perform several tests and then use the best one for code purposes. • The certification process indicates that the manufacturer is consistently producing the product that was rated. This corrects a past concern that manufacturers were able to make an exceptionally high-quality sample and obtain a good rating in a test but not consistently produce that product. • A readily visible temporary label can be used by the building inspector to quickly verify compliance with the energy code. • A permanent label enables future access to energy rating information. Although the NFRC program is similar for other fenestration characteristics, there are differences worth noting. Solar heat gain coefficient and visible transmittance ratings [NFRC (2014d) Technical Document 200], referenced in several codes, and condensation resistance ratings [NFRC (2014j) Technical Document 500] are based on simulation alone. Optical properties [NFRC (2014c) Technical Document 300] and emissivity [NFRC (2014g) Technical Document 301] are based on measurements by the manufacturer, with independent verification. Air leakage ratings [NFRC (2014h) Technical Document 400] are based on testing alone. For site-assembled fenestration products (e.g., curtain walls, window walls), an NFRC label certificate fulfills the labeling requirements and serves the certification purpose. A separate NFRC label certificate is required for each “individual product” in a particular project.
United States Energy Policy Act (EPAct) In the United States, the 1992 Energy Policy Act (EPAct) required the development of national fenestration energy rating systems and specified NFRC as the preferred developer. (The U.S. Department of Energy was to establish procedures if NFRC did not.) Although this recognition provided an impetus for NFRC to develop the desired procedures and programs, the EPAct sections on energy codes have been a key factor in their implementation. EPAct set baselines for state energy codes. The ICC 2015 International Energy Conservation Code® (IECC®) and ASHRAE/IES Standard 90.1-2016, Energy Standard for Buildings Except LowRise Residential Buildings, are the current successors to the versions cited in the 1992 legislation. The majority of states have adopted the predecessors to the 2015 IECC (including the 2012, 2009, 2006, 2003, 2000, and 1998 IECC and the CABO 1995 Model Energy Code) and to ASHRAE/IES Standard 90.1-2016 (i.e., ASHRAE/IES Standard 90.1-2013/2010/2007/2004/2001/1999/ 1989) into their codes either directly or by reference when adopting a building code published by one of the three national code organizations in the United States. The ICC 2015 International Building Code (the U.S. model building code jointly developed by ICBO, BOCA, and SBCCI) references the 2015 IECC.
ICC’s 2015 International Energy Conservation Code The ICC’s 2015 IECC references NFRC (2014a) Technical Document 100 for U-factor (as did the 2012, 2009, 2006, 2003, 2000, and 1998 IECC and the 1995 Model Energy Code) and NFRC (2014d) Technical Document 200 for solar heat gain coefficient
15.63 (SHGC) (as did the 2012, 2009, 2006, 2003, 2000, and 1998 IECC). Sections C303.1.1 and R303.1.3, which cover all occupancies, require U-factors and SHGCs and visible transmittances (Tvs) of fenestration products (windows, doors, and skylights) to be determined in accordance with NFRC Technical Documents 100 and 200 by an accredited independent laboratory, and labeled and certified by the manufacturer. The language does not specify NFRC accreditation; however, it requires both the use of the NFRC rating procedure by an independent entity, and labeling and certification. Sections C402.4.3 and R402.4.3 cite NFRC (2014h) Technical Document 400, and others for air leakage testing.
ASHRAE/IES Standard 90.1-2016 In 1999, ASHRAE and IES published a comprehensive update to Standard 90.1-1989 that included fenestration rating, labeling, and certification criteria in Sections 5.2.2 and 5.2.3. U-factors were to be determined in accordance with NFRC (2014a) Technical Document 100, solar heat gain coefficient and visible transmittance in accordance with NFRC (2014d) Technical Document 200, and air leakage in accordance with NFRC (2014h) Technical Document 400. In 2001, ASHRAE and IES made nominal modifications to Standard 90.1. The most significant changes for the 2004 version were in the lighting section, with fenestration rating, labeling, and certification criteria found in Sections 5.8.2. The 2007 revision included substantial increases in stringency for the building envelope, including both opaque assemblies and fenestration. The NFRC references remained unchanged. The 2010 update had limited changes to the fenestration U-factor and SHGC criteria, but called for reduced air leakage. The standard began to address fenestration orientation and daylighting: in the northern hemisphere, vertical fenestration must have more area on the south side than on either the east or west side. Spaces with tall ceilings and under a roof must have a minimum skylight area for daylighting purposes. Exceptions are provided to the SHGC requirements for dynamic glazing. The 2013 version made significant changes to fenestration Ufactor and SHGC criteria. The standard expanded the consideration of daylighting by establishing criteria for a minimum VT/SHGC ratio where there is automatic control of lighting in daylight zones. The 2016 update further changed the requirements for fenestration U-factors in most climate zones, and for SHGC in a few climate zones. Limited modifications were made to the fenestration orientation criteria. For further information on U.S. energy codes, the Building Codes Assistance Project (BCAP) publishes a bimonthly summary entitled “Status of State Energy Codes,” which provides information on current codes and pending legislation. For additional information, contact BCAP at www.bcap-energy.org.
ASHRAE/USGBC/IES Standard 189.1-2014 In 2006, ASHRAE, the U.S. Green Building Council (USGBC), and IES embarked on a project to develop a baseline standard for high-performance, green buildings that would apply to all buildings except low-rise residential buildings. Their Standard 189.1-2009 addresses sustainable sites, energy and water efficiency, the building’s effect on the atmosphere, materials and resources, and indoor environmental quality (IEQ). The standard is not a rating system, but it is hoped that it will be integrated into future building rating systems. Energy-efficiency goals for the first version of Standard 189.1 were to achieve a 30% additional energy savings beyond that in ASHRAE/IES Standard 90.1-2007. Standard 189.1 builds on Standard 90.1, but the prescriptive option in Standard 189.1 substitutes more stringent values in the tables and adds other criteria. For example, the prescriptive option requires that, in the northern hemisphere, vertical fenestration on the west, south, and east be shaded
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by a device or devices with a 0.50 projection factor (e.g., an overhang that is one-half as wide as the window is tall). There were subsequent updates in 2011 and 2014.
ICC’s 2015 International Green Construction Code™ The ICC’s 2012 International Green Construction Code™ (IgCC™) is an overlay code to ICC’s 2012 IECC. The fenestration criteria call for a 10% reduction in U-factor and SHGC from the 2012 IECC. The prescriptive option requires that, in the northern hemisphere, vertical fenestration on the west, south, and east be shaded by a device with a 0.25 projection factor (e.g., an overhang that is onequarter as wide as the window is tall). The IgCC also allows ASHRAE/USGBC/IES Standard 189.1 as a compliance option. After publication of the 2015 International Green Construction Code, ASHRAE and USGBC began investigating the possibilities for greater alignment between the IgCC and Standard 189.1.
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Canadian Standards Association (CSA) In Canada, the Canadian Standards Association (CSA) promulgates fenestration energy rating standards. CSA Standard A440.2 addresses most fenestration products, and CSA Standard A453 addresses doors. These are companion standards to NFRC (2014a) Technical Document 100. NFRC and CSA established a Thermal Harmonization Task Force to attempt to harmonize their fenestration energy rating standards.
Building Code of Australia/National Construction Code In Australia, the Australian Building Codes Board maintains the National Construction Code, which includes the Building Code of Australia (BCA). The code is divided into separate volumes for residential and commercial construction. Sections 2 and 3 (residential) and Section J (commercial) set out various compliance paths for building envelope performance. Section J of the National Construction Code covers energy efficiency provisions for nonresidential building stock in Australia. To demonstrate compliance with Section J2 of the National Construction Code, the U-factor and SHGC of fenestration systems are to be rated in accordance to with Australian Fenestration Rating Council (AFRC; www.afrc.org.au), which adopts the same technical procedures as NFRC. Users should note that the energy provisions for fenestration systems are based on an energy-flow approach and do not have an upper limit for fenestration area as a percentage of gross wall area.
Complex Glazings and Window Coverings In North America, the European Union, and Australia, fenestration rating programs are being extended to include shading devices and other nonspecular, “complex” glazing layers. These rating schemes rely on expanded capabilities of existing modeling software and measurements of basic material properties, supported by a set of industry-consensus assumptions about fitting details. However, such calculations can also be performed in “expert” mode to serve the needs of building energy modelers who require exact, rather than generalized, details.
9.7
SYMBOLS
a = absorptance in a layer, considered as an isolated layer A = total projected area of a fenestration product; apparent solar constant A = absorptance in a layer or a collection of layers (system or subsystem) e = hemispherical emissivity Ed = diffuse sky irradiance ED = direct irradiance EDN = direct normal irradiance Er = diffuse ground reflected irradiance Et = total irradiance FR = radiant fraction h = surface heat transfer coefficient
k L n PH PV q Q R
= = = = = = = =
RH = RW = SH SW SHGC t T
= = = = =
U= W=
thermal conductivity glass thickness refractive index horizontal projection depth vertical projection depth instantaneous energy flux instantaneous energy flow reflectance of a layer or collection of layers (system or subsystem) height of opaque surface between fenestration product and horizontal projection width of opaque surface between fenestration product and vertical projection shadow height shadow width solar heat gain coefficient relative temperature absolute temperature; transmittance of layer or collection of layers (system or subsystem) overall coefficient of heat transfer fenestration product width
Greek g
= = = = = = = = = = = = = =
material absorptivity solar altitude angle surface solar azimuth vertical projection profile angle declination incident angle wavelength refractive angle ground reflectivity surface tilt transmissivity solar azimuth horizontal projection profile angle solid angle
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Fenestration Laouadi, A., A.D. Galasiu, H.H. Saber, and C. Arsenault. 2013b. Tubular daylighting devices- part I: development of an optical model (1415-RP). HVAC&R Research (now Science and Technology for the Built Environment) 19(5):536-556. Laouadi, A., C. Arsenault, H.H. Saber, and A.D. Galasiu. 2013c. Tubular daylighting devices- part II: validation of the optical model (1415-RP). HVAC&R Research (now Science and Technology for the Built Environment) 19(5):557-572. Laouadi, A., H.H. Saber, A.D. Galasiu, and C. Arsenault. 2013d. Tubular daylighting devices—Development and validation of a thermal model (1415-RP). HVAC&R Research (now Science and Technology for the Built Environment) 19(5):513-535. LBNL. 2016. WINDOW 7. Windows and Daylighting Group, Lawrence Berkeley Laboratory, Berkeley, CA. windows.lbl.gov/software/window /window.html. Lee, E., and S.E. Selkowitz. 1995. The design and evaluation of integrated envelope and lighting control strategies for commercial buildings. ASHRAE Transactions 101(1):326-342. Lyons, P.R., D.A. Arasteh, and C. Huizenga. 2000. Window performance for human thermal comfort. ASHRAE Transactions 106(1):594-602. Lyons, P., J. Wong, and M. Bhandari. 2010. A comparison of window modeling methods in Energyplus 4.0. SimBuild 2010, New York. Mathis, R.C., and R. Garries. 1995. Instant, annual life: A discussion on the current practice and evolution of fenestration energy performance rating. Proceedings of Windows Innovations Conference ’95, Toronto, ON. McCluney, R. 1990. Awning shading algorithm update. ASHRAE Transactions 96(1):34-38. McCluney, R. 1993. Sensitivity of optical properties and solar gain of spectrally selective glazing systems to changes in solar spectrum. Solar ’93, 22nd American Solar Energy Society Conference, Washington, D.C. McCluney, R. 1994a. Angle of incidence and diffuse radiation influences on glazing system solar gain. Proceedings Solar ’94 Conference, American Solar Energy Society, San Jose, CA. McCluney, R. 1994b. Introduction to radiometry and photometry. Artech House, Boston. McCluney, R. 1996. Sensitivity of fenestration solar gain to source spectrum and angle of incidence. ASHRAE Transactions 102(2):112-122. McCluney, R., and P. Jindra. 2001. Industry guide to selecting the best residential window options for the Florida climate. FSEC-PF-358-01. Florida Solar Energy Center, University of Central Florida, Cocoa. McGowan, A, R. Jutras, G. Riopel, and M. Hanam. 2006. Heat transfer through roll-up doors, revolving doors and opaque non-residential swinging, sliding and rolling doors. ASHRAE Research Project RP-1236, Final Report. Mitchell, R., J. Huang, D. Arasteh, R. Sullivan, and S. Phillip. 1999. RESFEN 3.1: A PC program for calculating the heating and cooling energy use of windows in residential buildings—Program description. LBNL40682 Rev. BS-371. Lawrence Berkeley National Laboratory, Berkeley, CA. Mitchell, R., M. Yazdanian, L. Zhu, D. Arasteh, and J. Huang. 2012. RESFEN6: Program description. LBNL-40682 Rev. Lawrence Berkeley National Laboratory, Berkeley, CA. Moeseke, G., I. Bruyere, and A.D. Herde. 2007. Impact of control rules on the efficiency of shading devices and free cooling for office buildings. Building and Environment 42:784-793. Newsham, G.R. 1994. Manual control of window blinds and electric lighting: implications for comfort and energy consumption. Indoor Environment 3, pp. 135-144. NFRC. 2014a. Procedure for determining fenestration product U-factors. Technical Document 100-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014b. Spectral data library. Available from www.nfrc.org /software.aspx. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014c. Standard test method for determining the solar optical properties of glazing materials and systems. Technical Document 300-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014d. Procedure for determining fenestration product solar heat gain coefficient and visible transmittance at normal incidence. Technical Document 200-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014e. Procedure for interim standard test method for measuring he solar heat gain coefficient of fenestration systems using calorimetry hot box methods. Technical Document 201-2014. National Fenestration Rating Council, Silver Spring, MD.
15.67 NRFC. 2014f. Procedure for determining visible transmittance of tubular daylighting devices. Technical Document 203-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014g. Standard test method for emittance of specular surfaces using spectromteric measurements. Technical Document 301-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014h. Procedure for determining fenestration product air leakage. Technical Document 400-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014i. Guidelines to estimate the effects of fenestration on heating and cooling energy consumption in single family residences. Technical Document 901-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014j. Procedure for determining fenestration product condensation resistance values. Technical Document 500-2014. National Fenestration Rating Council, Silver Spring, MD. NFRC. 2014k. User guide to the procedure for determining fenestration product condensation resistance rating values. Technical Document 5012014. National Fenestration Rating Council, Silver Spring, MD. NRCC. 2011. Tubular daylighting devices: An alternative to skylights. Construction Innovation 16(1). National Research Council Canada, Ottawa. Ochoa, C.E., and I.G. Capeluto. 2009. Advice tool for early design stages of intelligent facades based on energy and visual comfort approach. Energy and Buildings 41(5):480-488. Parmelee, G.V., and R.G. Huebscher. 1947. Forced convection heat transfer from flat surfaces. ASHVE Transactions, pp. 245-284. Patenaude, A. 1995. Air infiltration rate of windows under temperature and pressure differentials. Proceedings of Windows Innovations Conference ’95, Toronto, ON. Pennington, C.W., and G.L. Moore. 1967. Measurement and application of solar properties of drapery shading materials. ASHRAE Transactions 73(1):8.3.1. Rao, S.R., and A. Tzempelikos. 2012. The impact of a combined dynamic shading system on the thermal performance of building perimeter zones. ASHRAE Transactions 118(1):265-273. Reinhart, C.F. 2004. Lightswitch-2002: A model for manual and automated control of electric lighting and blinds. Solar Energy 77(2004):15-28. Robbins, C.L. 1986. Daylighting: Design and analysis. Van Nostrand Reinhold, New York. Rubin, M. 1982a. Solar optical properties of windows. Energy Research 6: 122-133. Rubin, M. 1982b. Calculating heat transfer through windows. Energy Research 6:341-349. Rubin, M. 1985. Optical constants and bulk optical properties of soda lime silica glasses for windows. Solar Energy Materials 12:275-288. Rubin, M., K. von Rottkay, and R. Powles. 1998. Window optics. Solar Energy 62(3):149-161. Rubin, M., R. Powles, and K. von Rottkay. 1999. Models for the angledependent optical properties of coated glazing materials. Solar Energy 66(4):267-276. Schutrum, L.F., and N. Ozisik. 1961. Solar heat gains through domed skylights. ASHRAE Journal, pp. 51-60. Sezgen, O., and J.G. Koomey. 1998. Interactions between lighting and space conditioning energy use in U.S. commercial buildings. LBNL-39795, Lawrence Berkeley National Laboratory, Berkeley, CA. eetd.lbl.gov /sites/all/files/lbnl-39795.pdf. Shen, H., and A. Tzempelikos. 2012. Daylighting and energy analysis of private offices with automated interior roller shades. Solar Energy 86, pp. 681-704. Shewen, E.C. 1986. A Peltier-effect technique for natural convection heat flux measurement applied to the rectangular open cavity. Ph.D. dissertation. Department of Mechanical Engineering, University of Waterloo, ON. Smith, W.A., and C.W. Pennington. 1964. Shading coefficients for glass block panels. ASHRAE Journal 5(12):31. Sodergren, D., and T. Bostrom. 1971. Ventilating with the exhaust air window. ASHRAE Journal 13(4):51. Sterling, E.M., A. Arundel, and T.D. Sterling. 1985. Criteria for human exposure in occupied buildings. ASHRAE Transactions 91(1). Sullivan, H.F., J.L. Wright, and R.A. Fraser. 1996. Overview of a project to determine the surface temperatures of insulated glazing units: Thermographic measurement and two-dimensional simulation. ASHRAE Transactions 102(2):516-522.
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Tzempelikos, A., and A.K. Athienitis. 2005. Integrated daylighting and thermal analysis of office buildings. ASHRAE Transactions 111(1):227-238. Tzempelikos, A., and A.K. Athienitis. 2007. The impact of shading design and control on building cooling and lighting demand. Solar Energy 81(3), 369-382. van Dijk, D. 2003. Comparison of standards on thermal, solar and/or light properties of windows and window components. TNO Building and Construction Research, Delft, Netherlands. April 8. Ward, G.W. 1990. Visualization. Lighting Design + Application 20(6):4-20. Winkelmann, F.C. 1983. Daylighting calculation in DOE-2. LBNL-11353. Lawrence Berkeley National Laboratory, Berkeley, CA. Wong, J.C.S. 2016. Personal communication. Wright, J.L. 1995a. VISION4 glazing system thermal analysis: Reference manual. Advanced Glazing System Laboratory, University of Waterloo. Wright, J.L. 1995b. Summary and comparison of methods to calculate solar heat gain. ASHRAE Transactions 101(1):802-818. Wright, J.L. 1996a. A correlation to quantify convective heat transfer between window glazings. ASHRAE Transactions 102(1):940-946. Wright, J.L. 1996b. A simplified numerical method for assessing the condensation resistance of windows. ASHRAE Transactions 104(1):12221229. Wright, J.L. 2008. Calculating centre-glass performance indices of glazing systems with shading devices. ASHRAE Transactions 114(2):199-209. Wright, J.L., and H.F. Sullivan. 1995a. A two-dimensional numerical model for glazing system thermal analysis. ASHRAE Transactions 101(1):819-831. Wright, J.L., and H.F. Sullivan. 1995b. A two-dimensional numerical model for natural convection in a vertical, rectangular window cavity. ASHRAE Transactions 100(2):1193-1206.
Wright, J.L., and H.F. Sullivan. 1995c. A simplified method for the numerical condensation resistance analysis of windows. Proceedings of Windows Innovations Conference ’95, Toronto, ON. Wright, J.L., N.Y.T. Huang, and M.R. Collins. 2008. Thermal resistance of a window with an enclosed venetian blind: A simplified model. ASHRAE Transactions 114(1):471-482. Wright, J.L., M.R. Collins, and N. Kotey. 2009a. Solar gain through windows with shading devices: Simulations versus measurements. ASHRAE Transactions 115(2):18-30. Wright, J.L., C. Barnaby, M.R. Collins, and N. Kotey. 2009b. Improving load calculations for fenestrations with shading devices. ASHRAE Research Project RP-1311, Final Report. Yazdanian, M., and J.H. Klems. 1993. Measurement of the exterior convective film coefficient for windows in low-rise buildings. ASHRAE Transactions 100(1):1087-1096. Yellott, J.I. 1963. Selective reflectance—A new approach to solar heat control. ASHRAE Transactions 69:418. Zhang, X., and T. Muneer. 2000. Mathematical model for the performance of lightpipes. Lighting Research and Technology 32(3):141-146. Zhao, Y., D. Curcija, and W.P. Goss. 1996. Condensation resistance validation project—Detailed computer simulations using finite element methods. ASHRAE Transactions 102(2):508-515.
BIBLIOGRAPHY Wright, J.L., and N.A. Kotey. 2006. Solar absorption by each element in a glazing/shading layer array. ASHRAE Transactions 112(2):3-12. Zhao, Y., D. Curcija, and W.P. Goss. 1999. Convective heat transfer correlations for fenestration glazing cavities: A review. ASHRAE Transactions 105(2).
Related Commercial Resources
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Related Commercial Resources CHAPTER 16
VENTILATION AND INFILTRATION Basic Concepts and Terminology ............................................ 16.1 Tracer Gas Measurements....................................................... 16.5 Driving Mechanisms for Ventilation and Infiltration ............................................................................ 16.7 Indoor Air Quality ................................................................. 16.11 Thermal Loads ....................................................................... 16.11 Natural Ventilation ................................................................ 16.13 Residential Air Leakage......................................................... 16.15
Residential Ventilation ........................................................... Residential IAQ Control......................................................... Simplified Models of Residential Ventilation and Infiltration .......................................................................... Commercial and Institutional Air Leakage............................ Commercial and Institutional Ventilation.............................. Office Building Example ........................................................ Symbols ..................................................................................
ROVIDING a comfortable and healthy indoor environment for building occupants is the primary concern of HVAC engineers. Comfort and indoor air quality (IAQ) depend on many factors, including thermal regulation; control of internal and external sources of pollutants; supply of acceptable air; removal of unacceptable air; occupants’ activities and preferences; and proper construction, operation, and maintenance of building systems. Proper ventilation and infiltration are only part of achieving acceptable indoor air quality and thermal comfort. HVAC designers, occupants, and building owners must be aware of and address other factors as well. Further information on indoor environmental health may be found in Chapter 10. Changing ventilation and infiltration rates to solve thermal comfort problems and reduce energy consumption can affect indoor air quality and may be against building code or other regulations, so any changes should be approached with care and be under the direction of a registered professional engineer with expertise in HVAC analysis and design. HVAC design engineers and others concerned with building ventilation and indoor air quality should obtain a copy of ASHRAE Standard 62.1 or 62.2, or those for specific applications (e.g., Standard 170 for health care), whichever is most relevant to the project. These standards are reviewed regularly and contain ventilation design and evaluation requirements for commercial and institutional (Standard 62.1) and residential (Standard 62.2) buildings, respectively. When designing a new building or analyzing an existing building, check which version of Standard 62 has been adopted by the local code authority. An existing building may be required to meet the current version of the standard, or allowed to comply with an older version. The last chapter of each year’s ASHRAE Handbook (Chapter 39 of this volume) has a list of current standards. This chapter addresses commercial and institutional buildings, where ventilation concerns usually dominate (though infiltration should not be ignored), and single- and multifamily residences, where infiltration has traditionally been considered most important but ventilation issues have received increased attention in recent years. Basic concepts and terminology for both are presented before more advanced analytical and design techniques are given. Ventilation of industrial buildings is covered in Chapter 31 of the 2015 ASHRAE Handbook—HVAC Applications. However, many of the fundamental ideas and terminology presented in this chapter can also be applied to industrial buildings.
as U.S. Green Building Council’s (USGBC) Leadership in Energy and Environmental Design™ (LEED®) program, place great importance on creating and maintaining acceptable IAQ. In fact, the LEED rating system was first developed to address IAQ concerns, and roughly one-quarter of the available credit points for new commercial buildings are still IAQ related. Preparers of such rating systems, like others, have struggled with how to characterize complex ventilation and infiltration issues. These issues are addressed in detail by many portions of this chapter; separate ASHRAE design guides, manuals, books, and standards; and the references cited; these sources also provide methods to demonstrate the effectiveness of various HVAC systems and techniques in providing good IAQ in residential, commercial, and other buildings. In all designs, care is needed to eliminate excessive ventilation (e.g., beyond that needed for IAQ or by an air-side economizer) to avoid inappropriately increasing energy use. Increasing the ventilation rate above that required by Standard 62.1, for example, does not necessarily increase the acceptability of the indoor air quality.
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Sustainable Building Standards and Rating Systems Good indoor air quality is necessary for maintaining health and high productivity. Consequently, sustainable building standards such as ASHRAE Standard 189.1 and building rating systems, such The preparation of this chapter is assigned to TC 4.3, Ventilation Requirements and Infiltration.
16.1 Copyright © 2017, ASHRAE
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16.18 16.20 16.23 16.26 16.29 16.30 16.33
BASIC CONCEPTS AND TERMINOLOGY
Outdoor air that flows through a building is often used to dilute and remove indoor air contaminants. However, the energy required to condition this outdoor air can be a significant portion of the total space-conditioning load. The magnitude of outdoor airflow into the building must be determined to size the HVAC equipment properly, and to evaluate energy consumption (if required). For buildings without mechanical cooling and dehumidification, proper ventilation and infiltration airflows are important for providing acceptable IAQ and better thermal comfort for occupants. ASHRAE Standard 55 specifies conditions under which 80% or more of the occupants in a space will find it thermally acceptable. Chapter 9 of this volume also addresses thermal comfort. Airflow into buildings and between zones also affects fires, smoke movement, and safe occupant egress. Smoke management is addressed in Chapter 53 of the 2015 ASHRAE Handbook—HVAC Applications.
Ventilation and Infiltration Air exchange of outdoor air with air already in a building can be divided into two broad classifications: ventilation and infiltration. Ventilation is intentional introduction of air from the outdoors into a building; it is further subdivided into natural and mechanical ventilation. Natural ventilation is the flow of air through open windows, doors, grilles, and other planned building envelope penetrations. Mechanical (or forced) ventilation, shown in Figure 1, is the intentional movement of air into and out of a building using fans, ductwork, intake louvers, and exhaust grilles, for example. Infiltration is the flow of outdoor air into a building through cracks and other unintentional openings and through the normal use
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2017 ASHRAE Handbook—Fundamentals (SI)
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Fig. 1 Two-Space Building with Mechanical Ventilation, Infiltration, and Exfiltration of exterior doors for entrance and egress. Infiltration is also known as air leakage into a building. Exfiltration, depicted in Figure 1, is leakage of indoor air out of a building through similar types of openings. Like natural ventilation, infiltration and exfiltration are driven by natural and/or artificial pressure differences. These forces are discussed in detail in the section on Driving Mechanisms for Ventilation and Infiltration. Transfer air is air that moves from one interior space to another, either intentionally or not. Ventilation and infiltration differ significantly in how they affect energy consumption, air quality, and thermal comfort, and can each vary with weather conditions, HVAC system operation, and building use. Although one mode may be expected to dominate in a particular building, both must be considered in the proper design and operation of an HVAC system. Seasonal weather and other transient factors must be considered, as well.
Ventilation Air Ventilation air is air used to provide acceptable indoor air quality. It may be composed of mechanical or natural ventilation, infiltration, suitably treated recirculated air, transfer air, or an appropriate combination, although the allowable means of providing ventilation air varies in standards and guidelines. Modern commercial and institutional buildings normally have mechanical ventilation and are usually intended to be pressurized somewhat to reduce or eliminate infiltration. Mechanical ventilation has the greatest potential for control of air exchange when the system is properly designed, installed, and operated; it should provide acceptable indoor air quality and thermal comfort when ASHRAE Standards 55 and 62.1’s requirements are followed, although issues (e.g., unusually strong pollutant sources) can still result in unacceptable indoor environment conditions. Mechanical ventilation equipment and systems are described in Chapters 1, 4, and 10 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. In commercial and institutional buildings, natural ventilation (e.g., through uncontrolled use of manually operated windows) may not be desirable from the points of view of energy conservation, comfort, security, or control of airborne pollen or other pollutants in some climates and locations. In commercial and institutional buildings with mechanical cooling and ventilation, an automatically controlled air- or water-side economizer may be preferable to operable windows for taking advantage of cool outdoor conditions when interior cooling is required. When moderate outdoor temperatures occur, an air-side economizer control scheme may not only increase the rate of ventilation but also operate the cooling equipment to optimize energy use (hybrid or mixed mode). Infiltration may be significant in commercial and institutional buildings too, especially in tall, leaky, or partially pressurized buildings and in lobby and loading dock areas. The joint between roof decking and outer walls is often particularly leaky in commercial and other large buildings, and should be properly detailed, constructed, and inspected.
Fig. 2 Simple All-Air Air-Handling Unit with Associated Airflows In most of the United States, residential buildings have historically relied on infiltration and natural ventilation to meet their ventilation air needs. Neither is reliable for ventilation air purposes because they depend on weather conditions, building construction, occupants, and maintenance. Natural ventilation, usually through operable windows and screened doors, is more likely to allow occupants to control indoor airborne contaminants and interior air temperature, but it can have a substantial energy cost if used while the residence’s heating or cooling equipment is operating. Opened windows and doors also may lead to security, noise, or other concerns. In place of or in addition to operable windows, small exhaust fans should be provided for localized venting of residential spaces with high pollutant levels or moisture (e.g., kitchens, bathrooms). Not all local building codes require that such exhaust be vented to the outdoors, but it is required by ASHRAE Standard 62.2. Instead, a local code may allow the air to be treated and returned to the space or to be discharged to an attic space. Poor maintenance of these recirculating treatment devices can make nonducted vents ineffective for ventilation purposes. Warm exhaust air can hold much moisture, so condensation in attics should be avoided. If not already required by code, consider venting attached garages and other storage spaces to the outdoors, as well. Increasingly, building codes require general mechanical ventilation in residences. Heat recovery heat exchangers are popular for reducing energy consumption, especially in cold climates. Residential buildings with low rates of infiltration and natural ventilation, including most new buildings, require mechanical ventilation at rates given in ASHRAE Standard 62.2.
Forced-Air Distribution Systems Figure 2 shows a simple air-handling unit (AHU) or air handler that conditions air for a building. Air brought back to the air handler from the conditioned space is return air (RA). The return air either is discharged to the environment [exhaust air (EA)] or is reused [recirculated air (CA)]. Air brought in intentionally from the environment is outdoor air (OA). Because outdoor air may need treatment to be acceptable for use in a building, it should not be called “fresh air.” Outdoor and recirculated air are combined to form mixed air (MA), which is then conditioned and delivered to the spaces served as supply air (SA). Any portion of the mixed air that intentionally or unintentionally circumvents conditioning is bypass air (BA). Because of the wide variety of air-handling systems, the airflows shown in Figure 2 may not all be present in a particular system as defined here. Also, more complex systems may have additional airflows.
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Ventilation and Infiltration
16.3
In HVAC design, volumetric airflow rates Q are normally reported in litres per second (L/s) or cubic metres per second (m3/s). The incorrect term “volume” should not be used to describe airflow rates.
Outdoor Air Fraction The outdoor airflow introduced to a building or zone by an airhandling unit can also be described by the outdoor air fraction Xoa, which is the ratio of the volumetric flow rate Q of outdoor air brought in by the air handler to the total supply airflow rate:
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Q oa Q oa Q oa Xoa = --------- = ---------- = ------------------------Q oa + Q ca Q ma Q sa
(1)
When expressed as a percentage, the outdoor air fraction is called the percent outdoor air. The design outdoor airflow rate Qoa for a building’s or zone’s ventilation system is found by applying the requirements of ASHRAE Standard 62.1 or 62.2 to that specific building, occupancy, and HVAC system. The supply airflow rate Qsa is that required to meet the thermal load. The outdoor air fraction and percent outdoor air then describe the degree of recirculation, where a low value indicates a high rate of recirculation, and a high value shows little recirculation. Conventional all-air air-handling systems for commercial and institutional buildings often have approximately 10 to 40% outdoor air. 100% outdoor air means no recirculation of return air through the air-handling system. Instead, all the supply air is treated outdoor air, also known as makeup air (KA), and all return air is discharged directly to the outdoors as relief air (LA), via separate or centralized exhaust fans or relief dampers and grilles. An air-handling unit that provides exclusively 100% outdoor air to offset air that is exhausted is typically called a makeup air unit (MAU). When outdoor air via mechanical ventilation is used to provide ventilation air, as is common in commercial and institutional buildings and increasingly in residences, this outdoor air is usually delivered to spaces as all or part of the supply air. With a variableair-volume (VAV) system, the outdoor air fraction of the supply air may need to be increased when supply airflow is reduced to meet a particular thermal load. In some HVAC systems, such as a dedicated outdoor air system (DOAS), conditioned outdoor air may be delivered separately from the way the spaces’ loads are handled (Mumma and Shank 2001).
Room Air Movement Air movement within spaces affects the diffusion of ventilation air and, therefore, indoor air quality and comfort. Two distinct flow patterns are commonly used to characterize air movement in rooms: displacement flow and entrainment flow. Displacement flow, shown in Figure 3, is the movement of air within a space in a pistonor plug-type motion. Ideally, no mixing of the room air occurs, which is desirable for removing pollutants generated within a space. Air mixing does occur, however, to various degrees. A laminarflow air distribution system that is intended to sweep air across a space with reduced turbulence and mixing may produce a high degree of displacement flow and thus more effective pollutant removal. The pollutants’ buoyancy and occupants’ thermal comfort are concerns when deciding on the intended direction of airflow. Entrainment flow, shown in Figure 4, is also known as conventional mixing. Systems with ceiling-based supply air diffusers and return air grilles are common examples of air distribution systems that produce entrainment flow. Airborne pollutants are removed by dilution by the ventilation air that is delivered as all or part of the supply air. Entrainment flow with very poor mixing in the room has been called short-circuiting flow because much of the supply air leaves the room without mixing with room air. There is little evidence that properly designed, installed, and operated air distribution
Fig. 3
Displacement Flow Within a Space
Fig. 4 Entrainment Flow Within a Space systems exhibit substantial short circuiting, although poorly designed, installed, or operated systems may short circuit (especially ceiling-based systems in heating mode) to a higher degree (Offermann and Int-Hout 1989). Theoretical perfect mixing occurs when supply air is instantly and evenly distributed throughout a space. Perfect mixing is also known as complete or uniform mixing; the air may be called well stirred or well mixed. This theoretical performance is approached by entrainment flow systems that have good mixing and by displacement flow systems that allow too much mixing (Rock et al. 1995). The outdoor air requirements given in the minimum ventilation rate in breathing zone table of ASHRAE Standard 62.1 assume delivery of ventilation air with perfect mixing within spaces. For more detailed information on space air diffusion, see Chapter 20. Underfloor air distribution (UFAD or UAD), as shown in Figure 5, is a hybrid method of conditioning and ventilating spaces (Bauman and Daly 2003). Air is introduced through a floor plenum, with or without branch ductwork or terminal units, and delivered to a space by floor-mounted diffusers. These diffusers encourage air mixing near the floor to temper the supply air for thermal comfort. The combined air then moves vertically through the space, with reduced mixing, toward returns or exhausts placed in or near the ceiling. This vertical upward movement of the air is in the same direction as the thermal and contaminant plumes created by occupants and common equipment. The ventilating performance of UFAD systems is thus often between floor-to-ceiling displacement flow and uniform mixing. Supply air that enters a space through a diffuser, grille, or nozzle is also known as primary air. An air jet is formed as this primary air leaves the supply air outlet. Secondary air is the room air entrained into the jet. Total air is the combination of primary and
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2017 ASHRAE Handbook—Fundamentals (SI) For a particular space or zone, IS includes recirculated as well as any outdoor air in the supply air, and is used frequently in evaluating supply air outlet performance and space air mixing.
Time Constants Time constants , which have units of time (usually in hours or seconds), are also used to describe ventilation and infiltration. One time constant is the time required for one air change in a building, zone, or space if ideal displacement flow existed. It is the inverse of the air change rate: = 1/I = V/Q
(5)
The nominal time constant compares the interior volume of a building or zone to the volumetric outdoor airflow rate: N = V/Qoa
Fig. 5
Underfloor Air Distribution to Occupied Space Above
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(Rock and Zhu 2002)
secondary air at a specific point in a jet and increases with distance from the outlet as is described further in Chapter 20. The term primary air is also used to describe supply air provided to fan-powered mixing boxes by a central air-handling unit. For evaluation of indoor air quality and thermal comfort, rooms are often divided into two portions: the occupied zone and the remaining volume of the space. Often, this remaining volume is solely the space above the occupants and is referred to as the ceiling zone. The occupied zone is usually defined as the lowest 1.8 m of a room, although layers near the floor and walls are sometimes deducted from it; when these deductions are made, the occupied zone is sometimes renamed the breathing zone. Ceiling and floor plenums are not normally included in the occupied or ceiling zones. Thermal zones are different from these room air zones, and are defined for HVAC subsystems and their controls.
Air Change Rate The air change (or exchange) rate I compares airflow to the space’s volume and is I = Q/V
(2)
where Q = volumetric flow rate of air into space, m3/s V = interior volume of space, m3
The air change rate has units of 1/time, usually h–1. When the time unit is hours, the air change rate is also called air changes per hour (ACH), with units of h–1. The air change rate may be defined for several different situations. For example, the air change rate for an entire building or thermal zone served by an air-handling unit compares the amount of outdoor air brought into the building or zone to the total interior volume. This nominal air change rate IN is IN = Qoa /V
(3)
where Qoa is the outdoor airflow rate including ventilation and infiltration. IN describes the outdoor air ventilation rate entering a building or zone. It does not describe recirculation or the distribution of ventilation air to each space in a building or zone. For a particular space, the space air exchange rate IS compares the supply airflow rate Qsa to the volume of that space: IS = Qsa /V
(4)
(6)
Like the nominal air change rate, N does not describe recirculation of air in a building or zone, or characterize the distribution of the outdoor air to individual spaces in a building or zone. The space time constant compares the interior volume of a particular space to the total supply airflow rate to that space. The space time constant is the inverse of the space air change rate: S = V/Qsa
(7)
The space time constant includes the effect of recirculated air that is part of the supply air as well as that of outdoor air introduced to the space through the supply air. If infiltration is significant in a space, then the infiltration flow rate should be included when determining both the space air change rate and the space time constant.
Averaging Time-Varying Ventilation Rates When assessing time-varying ventilation in terms of controlling indoor air quality, the quantity of interest is often the temporal average rather than the peak. The concept of effective ventilation (Sherman and Wilson 1986; Yuill 1986, 1991) describes the proper ventilation rate averaging process. In this concept, the average (effective) rate is the steady-state rate that yields the same average contaminant concentration over the period of interest in the occupied space as does the actual sequence of time-varying discrete ventilation rates over the same period and in the same space. This effective rate is only equal to the simple arithmetic average rate when the discrete ventilation rates are constant over the period of interest and the contaminant concentration has reached its steadystate value. Simple arithmetic averaging of instantaneous ventilation rates or concentrations cannot generally be used to determine these averages because of the nonlinear response of indoor concentrations to ventilation rate variations. An important constraint in the effective ventilation concept is that the contaminant source strength F must be constant over the period of interest or must be uncorrelated with the ventilation rate. These conditions are satisfied in many residential and commercial buildings because the emission rates of many contaminants that are controlled by whole-building ventilation systems vary slowly. Sherman and Wilson (1986) describe how to deal with pollutants that have stepped but otherwise constant emission rates. Pollutants such as carbon monoxide, radon, and formaldehyde, whose emission rates can be affected by ventilation, cannot be properly characterized with this concept and require more complex analyses. For constant-source-strength pollutants, the relationship between effective air change rate, effective ventilation rate, volumetric flow, source strength, average concentration, and time-averaged effective turnover time is given by Q 1 F I m = ----- = -------- = ---V VC e
(8)
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The time-averaged effective turnover time e in Equation (8) represents the characteristic time for the concentration in the occupied space to approach steady state over the period of interest. It can be determined from a sequence of discrete, instantaneous ventilation air change rates Ii using the following (Sherman and Wilson 1986):
proposed. The specific measure that meets local code requirements must be determined, if any is needed at all. Air change effectiveness measures I are nondimensional gages of ventilation air delivery. One common definition of air change effectiveness is the ratio of a time constant to an age of air:
(9)
The nominal air change effectiveness I,N shows the effectiveness of outdoor air delivery to the entire building, zone, or space:
1 N e = ---- e, i N i=1
1 – exp – I i t for Ii > 0,e, i = ------------------------------------- + e, i – 1 exp(–Ii t) Ii
(10)
for Ii = 0,e, i = t + te, i – 1
(11)
where
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ASHRAE Standard 62.2 provides a set of factors to help calculate the annual effective air exchange rate.
Age of Air The age of air age (Sandberg 1981) is the length of time t that some quantity of outdoor air has been in a building, zone, or space. The “youngest” air is at the point where outdoor air enters the building by mechanical or natural ventilation, or through infiltration (Grieve 1989). The “oldest” air may be at some location in the building or in the exhaust air. When the characteristics of the air distribution system are varied, age of air is inversely correlated with quality of outdoor air delivery. Units are of time, usually in seconds or minutes, so it is not a true efficiency or effectiveness measure. The age of air concept, however, has gained wide acceptance in Europe. The age of air can be evaluated for existing buildings using tracer gas methods. Using either the decay (step-down) or growth (stepup) tracer gas method and assuming perfect mixing, the zone average or nominal age of air age,N can be determined by taking concentration measurements in the exhaust air. The local age of air age,L is evaluated through tracer gas measurements at any desired point in a space, such as at a worker’s desk. When time-dependent data of tracer gas concentration are available, the age of air can be calculated from dt t=0 -------------------C in – C o
C in – C
I, N = N /age, N
(12)
where Cin is the concentration of tracer gas being injected. Because evaluation of the age of air requires integration to infinite time, an exponential tail is usually added to the known concentration data (Farrington et al. 1990).
Air Change Effectiveness Ventilation effectiveness is a description of an air distribution system’s ability to remove internally generated pollutants from a building, zone, or space. Air change effectiveness is a description of an air distribution system’s ability to deliver ventilation air to a building, zone, or space. The HVAC design engineer usually does not have knowledge or control of actual pollutant sources within buildings, so the minimum prescribed ventilation rates of ASHRAE Standard 62.1 define outdoor air requirements for typical, expected building uses. For most projects, therefore, air change effectiveness is of more relevance to HVAC system design than ventilation effectiveness. Various definitions for air change effectiveness have been
(13)
(14)
where the nominal time constant N is usually calculated from measured airflow rates. The local air change effectiveness I,L shows the effectiveness of outdoor air delivery to one specific point in a space: I, L = N /age, L
t = length of each discrete time period e = time-averaged effective turnover time e, i = instantaneous turnover time in period i e, i–1 = instantaneous turnover time in previous period
age =
I = /age
(15)
where N is found either through airflow measurements or from tracer gas concentration data. An I,L value of 1.0 indicates that the air distribution system delivers air equivalent to that of a system with perfectly mixed air in the spaces. A value less than 1.0 shows less than perfect mixing with some degree of stagnation. A value of I,L greater than 1.0 suggests that a degree of plug or displacement flow is present at that point (Rock 1992). An HVAC design engineer often assumes that a properly designed, installed, operated, and maintained air distribution system provides an air change effectiveness of about 1. However, the zone air distribution table of ASHRAE Standard 62.1 provides some estimates of effectiveness for operating in heating or cooling mode, and with various air distribution techniques. These values are then adjusted for commercial and institutional building design when the ventilation rate procedure (VRP) is used. If the IAQ procedure of Standard 62.1 is used, then actual pollutant sources and the air change effectiveness must be known for the successful design of HVAC systems that have fixed ventilation airflow rates. ASHRAE Standard 129 describes a method for measuring air change effectiveness of mechanically vented spaces and buildings with limited air infiltration, exfiltration, and air leakage with surrounding indoor spaces.
2.
TRACER GAS MEASUREMENTS
The only reliable way to determine an existing building’s air change rate is to measure it. Several tracer gas measurement procedures exist (e.g., ASTM Standard E741 test method), all involving an inert or nonreactive gas used to label the indoor air (Charlesworth 1988; Dietz et al. 1986; Fisk et al. 1989; Fortmann et al. 1990; Harrje et al. 1981, 1990; Hunt 1980; Lagus 1989; Lagus and Persily 1985; Persily 1988; Persily and Axley 1990; Sherman 1989a, 1989b, 1990; Sherman et al. 1980). The tracer is released into the building in a specified manner, and the concentration of the tracer in the building is measured through time and related to the building’s air change rate. Various tracer gases and associated concentration detection devices have been used. Desirable qualities of a tracer gas are detectability, nonreactivity, nontoxicity, neutral buoyancy, relatively low concentration in ambient air, and low cost (Hunt 1980). All tracer gas measurement techniques are based on a mass balance of the tracer gas in the building. Assuming the outdoor concentration is zero and the indoor air is well mixed, this total balance takes the following form: dC V ------- = F(t) – Q(t)C(t) dt where V = volume of space being tested, m3
(16)
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C(t) dC/dt F(t) Q(t) t
2017 ASHRAE Handbook—Fundamentals (SI) = = = = =
tracer gas concentration at time t time rate of change of concentration, s1 tracer gas injection rate at time t, m3/s airflow rate out of building at time t, m3/s time, s
Constant Concentration In the constant concentration technique, the tracer gas injection rate is adjusted to maintain a constant concentration within the building. If the concentration is truly constant, then Equation (16) reduces to
In Equation (16), density differences between indoor and outdoor air are generally ignored for moderate climates; therefore, Q also refers to the airflow rate into the building. Although Q is often referred to as the infiltration rate, any measurement includes both mechanical and natural ventilation in addition to infiltration. The ratio of Q to the volume V being tested has units of 1/time, often converted to ACH, and is the air change rate I described previously in this chapter. Equation (16) is based on the assumptions that (1) no unknown tracer gas sources exist, (2) airflow out of the building is the dominant means of removing the tracer gas from the space so the tracer gas does not react chemically in the space and/or is not adsorbed onto or absorbed by interior surfaces or air cleaners, and (3) the tracer gas concentration in the building can be represented by a single value (i.e., the tracer gas is uniformly mixed within the space). In such tracer gas experiments, box-type fans are often operated in rooms to enhance mixing. Three different tracer gas procedures are used to measure air change rates: (1) decay or growth, (2) constant concentration, and (3) constant injection.
Q(t) = F(t)/C
There is less experience with this technique than with the decay procedure, and an increasing number of applications for it exist (Bohac et al. 1985; Collet 1981; Fortmann et al. 1990; Kumar et al. 1979; Walker and Forest 1995; Walker and Wilson 1998; Wilson and Walker 1993). Because tracer gas injection is continuous, no initial mixing period is required. Another advantage is that tracer gas injection into each zone of the building can be separately controlled; thus, the amount of outdoor air flowing into each zone can be determined. This procedure is best suited for longer-term continuous monitoring of fluctuating infiltration rates. One disadvantage is that it requires measurement of absolute tracer concentrations and injection rates. Also, imperfect mixing of the tracer and interior air causes a delay in the response of the concentration to changes in the injection rate.
Constant Injection In the constant-injection procedure, the tracer is injected at a constant rate, and the solution to Equation (16) becomes C(t) = (F/Q)(1 – e–It)
Decay or Growth Decay. The simplest tracer gas measurement technique is the decay method, also known as the step-down method. A small amount of tracer gas is injected into the space and is allowed to mix with the interior air. After the injection, F = 0 and then the solution to Equation (16) is C(t) = Coe–It
(17)
where Co is the concentration of the tracer in the space at t = 0. Equation (17) is generally used to solve for I by measuring the tracer gas concentration periodically during the decay and then fitting the data to the logarithmic form of Equation (17): ln C(t) = ln Co – I t
(18)
Like all tracer gas techniques, the decay method has advantages and disadvantages. One advantage is that, because logarithms of concentration are taken, only relative concentrations are needed, which can simplify calibration of concentration-measuring equipment. Also, the tracer gas injection rate need not be measured, although it must be controlled so that the tracer gas concentrations are within the range of the concentration-measuring device. The concentrationmeasuring equipment can be located on site, or building samples can be collected in suitable containers (e.g., grab bags) and analyzed elsewhere. The most serious problem with the decay technique is imperfect mixing of tracer gas with interior air, both at initial injection and during decay. Equations (16) and (17) assume that the tracer gas concentration within the building is uniform at any particular time. If the tracer is not well mixed, this assumption is not appropriate and the determination of I is subject to errors. It is difficult to estimate the magnitude of errors caused by poor mixing, and there has been little analysis of this problem. Sometimes a two-zone model is applied to a room, and a mixing coefficient selected, to estimate the effect of poor mixing [e.g., Rock (1992)]. Growth. The growth or step-up method is similar to the decay method except that the initial tracer gas concentration is low and the injected tracer gas is increased suddenly during the test.
(19)
(20)
After sufficient time, the transient term reduces to zero, the concentration attains equilibrium, and Equation (20) reduces to Q = F/C
(21)
Equation (21) is valid only when air change rate I and airflow rate Q are constant; thus, this technique is only appropriate for systems at or near equilibrium. It is particularly useful in spaces with mechanical ventilation or with high air change rates. Constant injection requires measurement of absolute concentrations and injection rates. Dietz et al. (1986) used a special case of the constant-injection technique, using permeation tubes as a tracer gas source. The tubes release the tracer at an ideally constant rate into the building being tested, and a sampling tube packed with an adsorbent collects the tracer from the interior air at a constant rate by diffusion. After a sampling period of one week or more, the sampler is removed and analyzed to determine the average tracer gas concentration in the building during the sampling period. Solving Equation (16) for C and taking the time average gives C = F/Q = F1/Q
(22)
where denotes time average. Note that the time average of dC/dt is assumed to equal zero. Equation (22) shows that the average tracer concentration C and injection rate F can be used to calculate the average of the inverse airflow rate. The average of the inverse is less than the inverse of the actual average, with the magnitude of this difference depending on the distribution of airflow rates during the measurement period. Sherman and Wilson (1986) calculated these differences to be about 20% for one-month averaging periods. Differences greater than 30% have been measured when occupant airing of houses caused large changes in air change rate; errors from 5 to 30% were measured when the variation was caused by weather effects (Bohac et al. 1987). Longer averaging periods and large changes in air change rates during the measurement periods generally lead to larger differences between the average inverse change rate and the inverse of the actual average rate.
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Multizone Air Change Measurement Equation (16) assumes a single, well-mixed enclosure, and the techniques described are for single-zone measurements. Multizone measurement techniques address airflow between internal zones and between the exterior and individual internal zones (Fortmann et al. 1990; Harrje et al. 1985, 1990; Sherman and Dickerhoff 1989). These techniques are important when considering the transport of pollutants from one room of a building to another. A theoretical development is provided by Sinden (1978a). Multizone measurements typically use either multiple tracer gases for the different zones or the constant-concentration technique. A proper uncertainty analysis is essential in all multizone flow determination (Charlesworth 1988; D’Ottavio et al. 1988).
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3.
DRIVING MECHANISMS FOR VENTILATION AND INFILTRATION
Natural ventilation and infiltration are driven by pressure differences across the building envelope caused by wind and air density differences. Mechanical air-moving systems also induce pressure differences across the envelope through operation of appliances, such as combustion devices, leaky forced-air thermal distribution systems, and mechanical ventilation systems. The indoor/outdoor pressure difference at a location depends on the magnitude of these driving mechanisms as well as on the characteristics of the openings in the building envelope.
Stack Pressure Stack pressure is the hydrostatic pressure caused by the weight of a column of air located inside or outside a building. It can also occur within a flow element, such as a duct or chimney that has vertical separation between its inlet and outlet. The hydrostatic pressure in the air depends on density and the height of interest above a reference point. Air density is a function of local barometric pressure, temperature, and humidity ratio, as described in Chapter 1. As a result, standard conditions should not be used to calculate the density. For example, a building site at 1500 m has air density that is about 20% less than if the building were at sea level. An air temperature increase from –30 to 20°C causes a similar air density difference. Combined, these elevation and temperature effects can reduce air density about 45%. Moisture effects on density are generally much less but can be significant if the change in elevation is great (e.g., in a natural draft cooling tower). Saturated air at 40°C has a density about 5% less than that of dry air at the same pressure. Assuming the air temperature and humidity ratio are constant over the height of interest, the stack pressure decreases linearly as the distance above the reference point increases. For a single column of air, the stack pressure can be calculated as ps = pr – gH
(23)
where ps pr g H
= = = = =
stack pressure, Pa stack pressure at reference height, Pa gravitational acceleration, 9.81 m/s2 indoor or outdoor air density, kg/m3 height above reference plane, m
For tall buildings or when significant temperature stratification occurs indoors, Equation (23) should be modified to include the density gradient over the height of the building. Temperature, and thus air density differences between indoors and outdoors cause stack pressure differences that drive airflows across the building envelope; the stack effect is this buoyancy phenomenon. Sherman (1991) showed that any single-zone building can be treated as an equivalent box from the point of view of stack
effect; if there is air leakage, follow the power law as described in the section on Residential Air Leakage. The building is then characterized by an effective stack height and neutral pressure level (NPL) or leakage distribution, as described in the section on Neutral Pressure Level. Once calculated, these parameters can be used in physical, single-zone models to estimate infiltration. Neglecting vertical density gradients, the stack pressure difference for a horizontal leak at any vertical location is T i – T o ps = (o – i)g(HNPL –H) = o ---------------- g(HNPL –H) Ti
(24)
where To Ti o i HNPL
= = = = =
absolute outdoor temperature, K absolute indoor temperature, K outdoor air density, kg/m3 indoor air density, kg/m3 height of neutral pressure level above reference plane without any other driving forces, m
Chastain and Colliver (1989) showed that, when there is stratification, the average of the vertical distribution of temperature differences is more appropriate to use in Equation (24) than the localized temperature difference near the opening of interest. By convention, stack pressure differences are positive when the building is pressurized relative to outdoors, which causes flow out of the building. Therefore, absent other driving forces and assuming no stack effect within the flow elements themselves, when indoor air is warmer than outdoors, the base of the building is depressurized and the top is pressurized relative to outdoors; when indoor air is cooler than outdoors, the reverse is true. At some elevation in the building, with such conditions, the pressure indoors is equal to the outdoors: this height is the neutral pressure level. Absent other driving forces, the location of the NPL is influenced by leakage distribution over the building exterior and by interior compartmentation. As a result, the NPL is not necessarily at the mid-height of the building; with effective horizontal barriers in tall buildings, it is also possible to have more than one NPL. NPL location and leakage distribution are described in the Combining Driving Forces and Neutral Pressure Level sections. For a penetration through the building envelope for which (1) there is vertical separation between its inlet and outlet and (2) air inside the flow element is not at the indoor or outdoor temperature (e.g., in a chimney), more complex analyses than Equation (24) are required to determine the stack effect at any location on the building envelope.
Wind Pressure When wind impinges on and flows around and over a building, it creates a distribution of static pressures on the building’s exterior surfaces that depends on the wind direction, wind speed, air density, surface orientation, and surrounding conditions. Wind pressures are generally positive with respect to the static pressure in the undisturbed airstream on the windward side of a building and negative on the leeward sides and roof. However, these pressures depend highly on wind speed, angle, turbulence, the surroundings, and building shape. Static pressures over building surfaces are almost proportional to the velocity pressure of the undisturbed airstream. The wind pressure or velocity pressure is given by the Bernoulli equation, assuming no height change or pressure losses: 2
U pw = Cp -----2 where
pw = wind surface pressure relative to outdoor static pressure in undisturbed flow, Pa
(25)
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= outdoor air density, kg/m3 (about 1.2 at or near sea level) U = wind speed, m/s Cp = wind surface pressure coefficient, dimensionless
Cp is a function of location on the building envelope and wind direction. Chapter 24 provides additional information on values of Cp. Most pressure coefficient data are for winds approaching perpendicularly to upwind building surfaces. Unfortunately, for a real building, this fixed wind direction rarely occurs, and when the wind is not normal to the upwind wall, these pressure coefficients do not apply. Walker and Wilson (1994) developed a harmonic trigonometric function to interpolate between the surface average pressure coefficients on a wall that were measured with the wind normal to each of the four building surfaces. This function was developed for low-rise buildings three stories or less in height. For each wall of the building, Cp is given by Cp() = 1/2{[Cp(1) + Cp(2)](cos2 )1/4 + [Cp(1) – Cp(2)](cos)3/4 + [Cp(3) + Cp(4)](sin2 )2 + [Cp(3) – Cp(4)]sin}
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= = = = =
1 s = --- s 1 + s 2 cos2 + s 1 – s 2 cos 2
(27)
+ s 3 + s 4 sin2 + s 3 – s 4 sin where
(26)
where Cp (1) Cp (2) Cp (3) Cp (4)
increase wind velocity. Thus, meteorological wind speed data must be adjusted carefully when applied to specific buildings and their locations. Infiltration rates measured by Wilson and Walker (1991) for a row of houses showed reductions in airflow rates of up to a factor of three when the wind changed direction from perpendicular to parallel to the row. They recommended estimating wind shelter for winds perpendicular to each side of the building and then using the interpolation function in Equation (27) to find the wind shelter for intermediate wind angles:
pressure coefficient when wind is at 0° pressure coefficient when wind is at 180° pressure coefficient when wind is at 90° pressure coefficient when wind is at 270° wind angle measured clockwise from the normal to wall 1
Because the cosine term in Equation (26) can be negative, its sign must be tracked. When cos() is negative, subtract the value of the absolute of cos() to the 3/4 power. The measured data used to develop the harmonic function from Akins et al. (1979) and Wiren (1985) show that typical values for the pressure coefficients are Cp(1) = 0.6, Cp(2) = –0.3, and Cp(3) = Cp(4) = –0.65. Because of geometry effects on flow around a building, application of this interpolation function is limited to low-rise buildings of rectangular plan on flat, featureless sites, with the longest wall less than three times the length of the shortest wall. For less regular buildings or sites, simple correlations are inadequate and building-specific pressure coefficients are required; computational fluid dynamic models are often used. Chapter 24 discusses wind pressures for complex building shapes and for high-rise buildings in more detail. The wind speed most commonly available for infiltration calculations is that measured at the local weather station, typically the nearest airport. This wind speed needs to be corrected for reductions caused by the difference between the height where the wind speed is measured and the height of the building, and reductions caused by shelter effects. The reference wind speed used to determine pressure coefficients is usually the wind speed at the eave height for a low-rise building and the building height for a high-rise building. However, meteorological wind speed measurements are made at a different height, typically 10 m for official weather stations, and at a different location than for the buildings of interest. The difference in terrain between the measurement station and the building under study must also be addressed. Chapter 24 shows how to calculate the effective wind speed UH from the reference wind speed Umet using boundary layer theory and estimates of terrain effects. In addition to the reduction in wind pressures caused by reduced wind speed, the effects of local shelter also act to reduce wind pressures. The shielding effects of trees, shrubbery, and other buildings within several building heights, horizontally, of a particular building produce large-scale turbulence eddies that not only reduce effective wind speed but also alter wind direction. Local geological features or gaps between neighboring large buildings can, at times, greatly
s = shelter factor for the particular wind direction s(i) = shelter factor when wind is normal to wall i (i = 1 to 4, for four sides of a building)
Although this method gives a realistic variation of wind shelter effects with wind direction, estimates for numerical values of wind shelter factor s for each of the four cardinal directions must be provided. Table 8 in the section on Residential Calculation Examples lists typical shelter factors. The wind speed used in Equation (25) is then given by U = sUH
(28)
The magnitude of pressure differences found on the surfaces of buildings varies rapidly with time because of turbulent fluctuations in the wind (Etheridge and Nolan 1979; Grimsrud et al. 1979). However, using average wind pressures to calculate pressure differences is usually sufficient to calculate average infiltration values.
Mechanical Systems Operation of mechanical equipment, such as supply or exhaust systems and vented combustion devices, affects pressure differences across the building envelope and thus air change rates. Interior static pressure adjusts such that the sum of all airflows through openings in the building envelope plus equipment-induced airflows balances to zero. To predict these changes in pressure differences and airflow rates caused by mechanical equipment, the location of each opening in the envelope and relationship between pressure difference and airflow rate for each opening must be known. The interaction between mechanical ventilation system operation and envelope airtightness has been discussed for low-rise buildings (Nylund 1980) and for office buildings (Persily and Grot 1985a; Tamura and Wilson 1966, 1967a). Air exhausted from a building by a whole-building exhaust system must be balanced by increasing airflow into the building through other openings or the air-handling systems. As pressures vary, air leakage at some locations changes from outflow to inflow. When using makeup air and no dedicated exhaust, the situation is reversed and envelope inflows may become outflows. Thus, the effects of a mechanical system on a building must be considered. Depressurization caused by an improperly designed exhaust system can increase the rate of radon entry into a building and can interfere with proper operation of combustion device venting or other exhaust systems. Pollutant entry can be increased from garages and other attached storage spaces. Depressurization can also force moist outdoor air through the building envelope; for example, during the cooling season in hot, humid climates, moisture may condense in the building envelope and cause rust, rot, or mold. A similar phe-
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nomenon, but in reverse, can occur during the heating, and potentially humidifying, season in cold climates if the building is pressurized. Active pressure control is often recommended, as is proper use of moisture retarders, drainage, and drying of in situ building materials. The interaction between mechanical systems and the building envelope also pertains to systems serving zones of buildings. Performance of zone-specific exhaust or pressurization systems is affected by leakage in partitions between zones as well as through exterior walls. Mechanical systems can also create infiltration-driving forces in single-zone buildings. Specifically, some single-family houses with central forced-air duct systems have many distributed supply registers, yet only one central return grille. When insufficiently undercut internal doors are closed in these houses, large positive indoor-to-outdoor pressure differentials are created for rooms with only supply registers, whereas the room or hallway with the return grille tends to depressurize relative to the outdoors. This is caused by the resistance of the internal door undercuts, often partially blocked by carpeting, to flow from the supply register to the return; the magnitudes of the indoor/outdoor pressure differentials created average 3 to 6 Pa (Modera et al. 1991). Balanced airflow systems with ducted air return and distributed grilles or adequately sized transfer grilles (where still allowed by fire code) reduce this effect significantly. Building envelope airtightness and interzonal airflow resistance can also affect performance of mechanical systems. The actual airflow rate delivered by these systems, particularly ventilation systems, depends on the pressure differences they work against. This effect is the same as the interaction of a fan with its associated ductwork, which is discussed in Chapter 21 of this volume and Chapter 21 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. The building envelope and its leakage should be considered part of the ductwork in determining the pressure drop of the system. Duct leakage can cause similar problems. Supply leaks to the outdoors tend to depressurize the building; return leaks from the outdoors tend to pressurize it. Keeping these ducts within the conditioned buildings, and sealing all ducts well with durable materials and high-quality construction methods, significantly reduces this problem.
Combining Driving Forces Pressure differences caused by wind, stack effect, and mechanical systems are considered in combination by adding them together and then determining the resulting airflow rate through each building envelope. The airflows must be determined in this manner, as opposed to adding the airflow rates caused by the separate driving forces, because the airflow rate through each opening is not linearly related to pressure difference. For uniform indoor air temperatures, the total pressure difference across each leak can be written in terms of a reference wind parameter PU and stack effect parameter PT common to all leaks: 2
UH PU = o -------2
(29)
PT = go[(Ti – To)/Ti]
(30)
where T is absolute air temperature in K. The pressure difference across each leak, with positive pressures for flow into the building, is then given by p = s2Cp PU + HPT + pI
(31)
where pI is the pressure that acts to balance inflows and outflows, including mechanical system flows. Equation (31) can then be applied to every leak for the building with appropriate values of Cp,
s, and H. Thus, each leak is defined by its pressure coefficient, shelter, and height. Where indoor pressures are not uniform, more complex and often numerical analyses are required.
Neutral Pressure Level The neutral pressure level (NPL) varies and is that height or heights in the building envelope where, at that particular instant, there is no indoor-to-outdoor pressure difference. Internal partitions, stairwells, elevator shafts, utility ducts, chimneys, vents, operable windows, and mechanical supply and exhaust systems complicate the prediction of NPL location. An opening with a large area relative to the total building leakage causes the NPL to shift toward the opening. In particular, chimneys and openings at or above roof height raise the NPL in small buildings. Exhaust systems increase the height of the NPL; outdoor air supply systems lower it. Figure 6 qualitatively shows the addition of driving forces for a building with uniform openings above and below mid-height and without significant internal resistance to airflow. The slopes of the pressure lines are a function of the densities of the indoor and outdoor air. In Figure 6A, with indoor air warmer than outdoor and pressure differences caused solely by thermal forces, the NPL is at mid-height, with inflow through lower openings and outflow through higher openings. For the low air velocities typical in and around buildings, the direction of flow is always from the higher to the lower-pressure region. Figure 6B presents qualitative uniform pressure differences caused by wind alone, with opposing effects on the windward and leeward sides. When temperature difference and wind effects both exist, the pressures caused by each are added together to determine the total pressure difference across the building envelope. In Figure 6B, there is no NPL because no locations on the building envelope have zero pressure difference. Figure 6C shows the combination, where the wind force of Figure 6B has just balanced the thermal force of Figure 6A, causing no pressure difference at the top windward or bottom leeward side. The relative importance of wind and stack pressures in a building depends on building height, internal resistance to vertical airflow, location and flow resistance characteristics of envelope openings, local terrain, and the immediate shielding of the building. The taller the building and the smaller its internal resistances to airflow, the stronger the stack effect. The stack effect can be reduced by effectively sealing the building internally between floors, typically by gasketing elevator and stairway doors, and sealing pipe, duct, and electrical penetrations; these measures, when done by code-approved means, also typically reduce undesired smoke migration during fire events. Gasketing interior doors, especially those from exterior spaces or to elevator lobbies, in tall buildings can also help restrict air leakage paths. The effect of mechanical ventilation on envelope pressure differences is more complex and depends on both the direction of ventilation flow (exhaust or supply) and the differences in these ventilation flows among the zones of the building. If mechanically supplied outdoor air is provided uniformly to each story, the change in the exterior wall pressure difference pattern is uniform. With a nonuniform supply of outdoor air (e.g., to one story only), the extent of pressurization varies from story to story and depends on internal airflow resistance. Pressurizing all levels uniformly has little effect on pressure differences across floors and vertical shaft enclosures, but pressurizing individual stories increases the pressure drop across these internal separations. Pressurizing the ground level is often used in tall buildings in winter to reduce negative air pressures across entries; vestibules and revolving doors are also used to limit air leakage. Vestibules may also be used for elevator lobbies and stair towers to reduce air and smoke movement vertically through tall buildings.
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Fig. 6 Distribution of Indoor and Outdoor Pressures over Height of Building Available data on the NPL in various kinds of buildings are limited. In tall buildings studied by Tamura and Wilson (1966, 1967b), the NPL varied from 0.3 to 0.7 of total building height. For houses, especially those with chimneys, the NPL is usually above midheight. Operating a combustion heat source that vents to the outdoors raises the NPL further, sometimes above the ceiling (Shaw and Brown 1982).
Thermal Draft Coefficient Compartmentation of a building also affects the NPL location. Equation (24) provides a maximum stack pressure difference, given no internal airflow resistance. The sum of pressure differences across the exterior wall at the bottom and top of the building, as calculated by these equations, equals the total theoretical draft for the building. The sum of actual top and bottom pressure differences, divided by the total theoretical draft pressure difference, equals the thermal draft coefficient. The value of the thermal draft coefficient depends on the airflow resistance of exterior walls relative to the airflow resistance between floors. For a building without internal partitions, the total theoretical draft is achieved across the exterior walls (Figure 7A), and the thermal draft coefficient equals 1. In a building with airtight separations between each floor, each story acts independently, its own stack effect being unaffected by that of any other floor (Figure 7B). The theoretical draft is minimized in this case, and each story has its own NPL. Real multistory buildings are neither open inside, nor airtight between stories. Vertical air passages, stairwells, elevators, and other service shafts allow airflow between floors. Figure 7C represents a heated building with uniform openings in the exterior wall, through each floor, and into the vertical shaft at each story. Between floors, the slope of the line representing the indoor pressure is the same as that shown in Figure 7A, and the discontinuity at each floor (Figure 7B) represents the pressure difference across it. Some of the pressure difference maintains flow through openings in the floors and vertical shafts. As a result, the pressure difference across the exterior wall at any level is less than it would be with no internal flow resistance. Maintaining airtightness between floors and from floors to vertical shafts is a way to control indoor/outdoor pressure differences because of the stack effect and, therefore, infiltration and exfiltration. Good separation is also conducive to proper operation of mechanical ventilation and smoke management systems. However, care is needed to avoid creating pressure differences that could prevent
Fig. 7 Compartmentation Effect in Buildings
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Ventilation and Infiltration
16.11
egress doors from opening in an emergency. Tamura and Wilson (1967a) showed that when vertical shaft leakage is at least two times envelope leakage, the thermal draft coefficient is almost one and the effect of compartmentation is negligible. Measurements of pressure differences in three tall office buildings by Tamura and Wilson (1967b) indicated that the thermal draft coefficient ranged from 0.8 to 0.9 with ventilation systems off. Modern internal sealing techniques should result in much less vertical leakage.
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4.
INDOOR AIR QUALITY
Outdoor air requirements for creating and maintaining acceptable indoor air quality (IAQ) have long been debated, and different rationales have produced radically different ventilation standards (Grimsrud and Teichman 1989; Janssen 1989; Klauss et al. 1970; Yaglou et al. 1936; Yaglou and Witheridge 1937). Historically, the major considerations have included the rate of outdoor air admission required to control airborne moisture, carbon dioxide (CO2), odors, and tobacco smoke generated by occupants. These considerations have led to prescriptions of minimum rates of outdoor air supply per occupant or by unit floor area, or both. More recently, a major concern has been maintaining acceptable indoor concentrations of various additional pollutants that are not generated primarily by occupants. Engineering experience and field studies indicate that an outdoor air supply of about 10 L/s per person is very likely to provide acceptable perceived indoor air quality in office spaces, whereas lower rates may lead to increased sick building syndrome symptoms (Apte et al. 2000; Mendell 1993; Seppanen et al. 1999; Sundell et al. 2011). Information on contaminants can be found in Chapter 11, and odors are addressed in Chapter 12. Indoor pollutant concentrations depend on the strength of pollutant sources and the total rate of pollutant removal. Pollutant sources include outdoor air; indoor sources such as occupants, furnishings, and appliances; lack of cleanliness as well as use of cleaning and other products; soil adjacent to the building; and building materials themselves, especially when new. Pollutant removal processes include dilution with cleaner outdoor air, local exhaust ventilation, deposition on surfaces, chemical reactions, and air-cleaning processes. If (1) general building ventilation is the only significant pollutant removal process, (2) indoor air is thoroughly mixed, and (3) pollutant source strength and ventilation rate have been stable for a sufficient period, then the steady-state indoor concentration of a specific pollutant is given by Ci = Co + 106S/Qoa
(32)
where Ci Co S Qoa
= = = =
steady-state indoor concentration, ppm outdoor concentration, ppm total pollutant source strength, m3/s ventilation rate, m3/s
Variation in pollutant source strengths, rather than variation in ventilation rate, is considered the largest cause of building-tobuilding variation in concentrations of pollutants that are not generated by occupants. Turk et al. (1989) found that a lack of correlation between average indoor respirable particle concentrations and wholebuilding outdoor ventilation rate indicated that source strength, high outdoor concentrations, building volume, and removal processes are important. Because pollutant source strengths are highly variable, maintaining minimum ventilation rates does not ensure acceptable indoor air quality in all situations. The lack of health-based concentration standards for many indoor air pollutants, primarily because of the lack of health data, makes the specification of minimum ventilation rates difficult. In cases of high contaminant source strengths, such as with indoor sanding, spray painting, or smoking, impractically high rates of dilution ventilation are required to control contaminant levels,
and other methods of control are more effective. Removal or reduction of contaminant sources is the most effective means of control. Controlling a localized source by means of local exhaust, such as paint booths, laboratory fume hoods, range hoods, or bathroom exhaust fans, as well as filtration and absorption, may also be effective [e.g., Rock (2006)]. Some particles can be removed with various types of air filters. Gaseous contaminants with higher relative molecular mass can often be controlled with activated carbon or alumina pellets impregnated with a substance such as potassium permanganate. Chapter 29 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment has information on air cleaning.
Protection from Extraordinary Events The design, operation and maintenance of a building’s ventilation system and envelope, as well as other factors, can significantly affect the building’s potential vulnerability to extraordinary threats, which range from intentional releases of chemical or biological agents inside or outside a building, to releases of chemicals in industrial or transportation accidents, to natural disasters. ASHRAE (2003) addresses several key steps to manage risk from extraordinary incidents, including • • • •
Evaluate the risk to a facility of an extraordinary incident Assess the building’s vulnerability Determine the degree of acceptable vulnerability Consider protective measures or options in relation to new, renovated, and existing buildings
Persily (2004) details how ventilation affects buildings’ vulnerability to airborne chemical and biological releases, as well as some strategies for using ventilation, particularly involving airtightness and pressurizing the building interior to protect against outdoor releases, to increase the level of building protection against such incidents. Persily et al. (2007) evaluate retrofit options for protecting buildings from airborne threats; approaches considered include enhanced particle filtration, sorbent-based gaseous air cleaning, ventilation system recommissioning, building envelope airtightening, building pressurization, relocation of outdoor air intakes, shelter-in-place (SIP), isolation of vulnerable spaces such as lobbies, and system shutdown and purge cycles. The filtration and air cleaning options have the advantage of always being operational as long as the systems are properly designed, installed, and maintained. However, the lack of standard test methods is a critical issue in application of some air-cleaning technologies. Building envelope air sealing and pressurization can be quite effective in protecting against outdoor releases as long as effective filtration against the contaminant of concern is also in place. Protection provided by operational changes such as system shutdown and purging depends heavily on timing; if timing is inappropriate, occupant exposure may be increased. Isolating vulnerable zones and other systemrelated modifications depend on building layout and system design, and careful implementation is necessary for effectiveness under the range of conditions that exist in buildings. Finally, many retrofits can also increase energy efficiency and improve indoor air quality, which should be included in a life-cycle cost comparison of different options to the degree possible. Chapter 59 of the 2015 ASHRAE Handbook—HVAC Applications addresses extraordinary events further.
5.
THERMAL LOADS
Outdoor air introduced into a building constitutes a large part of the total space-conditioning load, which is one reason to limit air change rates to the minimum required. Air exchange typically represents 20 to 50% of a modern, ventilation-code-compliant building’s thermal load in nontemperate climates. The effect on heating
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2017 ASHRAE Handbook—Fundamentals (SI)
loads tends to be much larger than on cooling loads (McDowell et al. 2003). Chapters 17 and 18 address thermal loads in more detail. Air exchange increases a building’s thermal load in several ways. First, incoming air must be heated or cooled from the outdoor air temperature to the indoor or supply air temperature. The rate of energy consumption by this sensible heating or cooling is qs = Qcp T
(33)
where qs Q cp T
= = = = =
sensible heat load, W airflow rate, m3/s air density, kg/m3 (about 1.2 at or near sea level) specific heat of air, J/(kg·K) (about 1000) temperature difference between indoors and outdoors, K
and at or near sea-level air density, with an adjustment for typical room air humidity, this equation is commonly presented for design use as
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qs = 1230Qt
(34)
Equations (33) and (34) are known as the sensible heat equation. HVAC designers typically assume sea-level air pressure for locations with altitudes of 610 m or lower. A method to adjust for elevation is provided in Chapter 18. Air exchange also modifies the moisture content of the air in a building. The rate of energy consumption associated with these latent loads, to add or remove water from the air and neglecting the energy associated with any condensate, is ql = QW(2501 + 1.805T)
(35)
where ql = latent heat load, W W = humidity ratio difference between indoors and outdoors, kg water/kg dry air, kg/kg T = average of indoor and outdoor temperatures, °C
Equation (35) is known as the latent heat equation. When at or near sea level, and for common comfort air temperatures, the righthand side of Equation (35) is approximately 3.01 106 Q W. Example 1. A makeup air unit (MAU) is to condition 2360 L/s of outdoor air in the winter for a building in Atlanta, Georgia. If the air is to be delivered directly to the occupied spaces at 24°C and 30% rh, how much sensible and latent heat must be added to this ventilation air at winter design conditions? Solution: From the weather data tables provided on the CD included with this volume, Atlanta is at an elevation of about 300 m. Because this is below the rule-of-thumb cutoff of 600 m for assuming sea-level conditions, air density is assumed to be 1.20 kg/m3. Also from the Atlanta data table, the winter 99% design dry-bulb (db) temperature is –3.1°C, but a mean coincident wet bulb is not provided. However, for humidification design, a dew-point temperature of –12.7°C is given along with its –0.1°C mean coincident dry bulb (MCDB). Using these data, a –17.2°C dew point is assumed as the 99% mean coincident dew point. From ASHRAE’s sea-level psychrometric chart and the winter design conditions, the desired humidity ratio W of the 24°C, 30% rh makeup air is about 0.0056 kgw /kgda. For the very dry outdoor air, with a dew point of –17.2°C, a problem occurs: the standard sea-level psychrometric chart does not extend below 0°C. Designers often assume that air below this temperature has W = 0, and this assumption gives conservative results. However, both high- and low-temperature psychrometric charts are available from ASHRAE, as is a table of moist air properties at standard conditions in Chapter 1. From this table, saturated air at –17.2°C, which is also its dew point, has a humidity ratio of 0.0008298 kgw /kgda. With 2360 L/s of outdoor air to be conditioned, and using the sensible and latent heat equations for sea level, the energy needed to condition this outdoor air is qs = 1.23QT = 1.23 × 2360 L/s[24 – (–3.1°C)] = 78 666 W 78.7 kW
and ql = 3010QW = 3010 × 2360 L/s(0.0056 – 0.0008298 kgw / kg da) = 33 886 W 33.9 kW Thus, the MAU’s heating coil and humidifier, neglecting fan heat, need to be sized to provide at least a net 78.7 kW of sensible heat, and 33.9 kW of latent heat. Humidification can be provided by cold water, warm water, or steam, so a more precise psychrometric analysis is needed to size the heating coil correctly after the humidification method is selected and it is decided whether the humidifier will be placed before or after the heating coil.
As Example 1 shows, ventilation loads are substantial. They are often 50% or more of the total space conditioning loads in modern, well-insulated, high-occupancy commercial buildings in less temperate climates. When cooling outdoor air, substantial moisture usually must be removed from the ventilation air; reheat or regenerative heat recovery may be required in all but dry climates.
Effect on Envelope Insulation Air exchange also can affect a building’s thermal load by altering performance of the envelope insulation system. Airflow through insulation can decrease the thermal load through heat exchange between infiltrating or exfiltrating air and the insulation. Conversely, air moving in and out of the insulation from the outdoors can increase the thermal load. Experimental and numerical studies demonstrate that significant thermal coupling can occur between air leakage and insulation layers, thereby modifying the heat transmission in building envelopes. In particular, research (Bankvall 1987; Berlad et al. 1978; Lecompte 1987; Wolf 1966) shows that convective airflow through air-permeable insulation in an envelope assembly may degrade its effective thermal resistance. This R-value degradation occurs when outdoor air moves through and/or around the insulation within the wall cavity and returns to the outdoors without reaching the conditioned space. A literature review by Powell et al. (1989) summarized the findings about air movement effects on the effective thermal resistance of porous insulation under various conditions. The effect of such airflow on insulation system performance is difficult to quantify, but should be considered. Airflow within the insulation system can also decrease the system’s performance because of moisture condensation in and on the insulation. Even if air flows only through cracks instead of through the insulation, the actual heating/cooling load from the combined effect of conduction and airflow heat transfer can be lower than the heating/cooling load calculated by Equation (33). This reduction in total heating/cooling load is a consequence of the thermal coupling between conduction and convection heat transfer and is called infiltration heat recovery (IHR). Using a computer simulation, Kohonen et al. (1987) found that the conduction/infiltration thermal interaction reduced total heating load by 15%. Several experimental studies (e.g., Claridge and Bhattacharyya 1990; Claridge et al. 1988; Liu and Claridge 1992a, 1992b, 1992c, 1995; Timusk et al. 1992), using a test cell under both steady-state and dynamic conditions, found that the actual energy attributed to air infiltration can be 20 to 80% of the values given by Equation (35). Judkoff et al. (1997) measured heat recovery in a mobile home under steady-state conditions, and found that up to 40% heat recovery occurs during exfiltration through the envelope. Buchanan and Sherman (2000) performed two- and three-dimensional computational fluid dynamics (CFD) simulations to study the fundamental physics of the IHR process and developed a simple macro-scale mathematical model based on the steady-state one-dimensional convection-diffusion equation to predict a heat recovery factor. Their results show that the traditional method may overpredict the infiltration energy load. Using physical experiments, ASHRAE research project RP-1169 (Ackerman et al. 2006) showed that thermal resistances are affected
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Ventilation and Infiltration by infiltration and exfiltration, but, on a net basis, the IHR effect can be neglected.
Infiltration Degree-Days Heating and cooling degree-days (HDDs and CDDs) are a simple way to characterize the severity of a particular climate. Heating and cooling degree-day values are based on sensible temperature data, but infiltration loads are both sensible and latent. Infiltration degree days (IDDs) more fully describe a climate and can be used to estimate heat loss or gain from infiltration in residences (Sherman 1986). Total infiltration degree-days is the sum of the heating and cooling infiltration degree-days and is calculated from hour-byhour weather data and base conditions using weather weighted by infiltration rate. The selection of base conditions is an important part of the calculation of the IDDs. ASHRAE Standard 119 lists IDDs for many locations with a particular set of base conditions.
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6.
NATURAL VENTILATION
Natural ventilation is the flow of outdoor air caused by wind and thermal pressures through intentional openings in a building’s shell. Under some circumstances, it can effectively control temperature, contaminants, and possibly airborne moisture in mild climates, but it is not considered practical in hot and humid climates in the summer or in cold climates during the winter. Temperature control by natural ventilation is often the only means of providing some cooling when mechanical air conditioning is not available. The arrangement, location, and control of ventilation openings should combine the driving forces of wind and temperature to achieve a desired ventilation rate and good distribution of ventilation air through the building. However, intentional openings cannot always guarantee adequate temperature and humidity control or indoor air quality because of the dependence on natural wind and stack effects to drive the flow (Wilson and Walker 1992). Using night ventilation and the building’s thermal mass effect may be effective for reducing conventional cooling energy consumption in some buildings and climates if moisture condensation can be controlled. Axley (2001a) and the Chartered Institute of Building Services Engineers (CIBSE 2005) reviewed natural ventilation in commercial buildings, including potential advantages and problems, natural ventilation components and system designs, and recommended design and analysis approaches.
16.13 exhaust fans; their motors need only be energized when the natural exhaust capacity is too low. Gravity ventilator dampers can be manual or controlled by thermostat or wind velocity. A natural-draft roof ventilator should be positioned so that it receives full, unrestricted wind. Turbulence created by surrounding obstructions, including higher adjacent buildings, impairs a ventilator’s ejector action. Inlets can be conical or bell-mouthed to increase their flow coefficients. The opening area at any inlet should be increased if screens, grilles, or other structural members cause flow resistance. Building air inlets, most effectively installed at the lowest levels, should be larger than the combined throat areas of all roof ventilators. Stacks or vertical flues should be located where wind can act on them from any direction. Without wind, stack effect alone removes some air from the rooms with inlets. Much practical experience with natural ventilation from before the adoption of modern air-conditioning appears in publications by ASHVE (a predecessor society to ASHRAE).
Ceiling Heights In buildings that use natural ventilation, floor-to-ceiling heights are often increased well beyond the normal 2.5 to 3.2 m. Higher ceilings, as seen in buildings constructed before air conditioning was available, allow warm air and contaminants to rise above the occupied portions of rooms. Air is then exhausted from the ceiling zones, and outdoor air is introduced near the floors; a degree of floor-to-ceiling displacement airflow is thus desirable when using natural ventilation during the cooling season.
Required Flow for Indoor Temperature Control The ventilation airflow rate required to remove a given amount of heat from a building can be calculated from Equations (33) and (35) if the quantity of heat to be removed and the indoor/outdoor temperature and humidity ratio differences are known.
Airflow Through Large Intentional Openings The relationship describing the airflow through a large intentional opening is based on the Bernoulli equation with steady, incompressible flow. The general form that includes stack, wind, and mechanical ventilation pressures across the opening is Q = CDA 2 p
Natural Ventilation Openings Natural ventilation openings include (1) operable windows, clerestories, doors, and skylights; (2) roof ventilators; (3) stacks; and (4) specially designed inlet or outlet openings such as OA louvers and EA grilles, ventilation penthouses or shafts, chimneys, windcatchers, and roof monitors. Operable windows transmit light, and provide ventilation when opened. They may open by sliding vertically or horizontally; by tilting on horizontal pivots at or near the center; or by swinging on pivots at the top, bottom, or side. The type of pivoting used is important for weather protection and affects airflow rate. Exterior doors, often with insect screens, can also provide a path for natural ventilation; as with operable windows, security concerns must be considered. Roof ventilators provide a weather-resistant air outlet. Capacity is determined by the ventilator’s location on the roof; the resistance to airflow of the ventilator and its ductwork; the ventilator’s ability to use kinetic wind energy to induce flow by centrifugal or ejector action; and the height of the draft. Natural-draft or gravity roof ventilators can be stationary, pivoting, oscillating, or rotating. Selection criteria include appearance, ruggedness, corrosion resistance, stormproofing features, dampers and operating mechanisms, noise, cost, and maintenance. Natural ventilators can be supplemented with power-driven supply or
(36)
where Q CD A p
= = = = =
airflow rate, m3/s discharge coefficient for opening, dimensionless cross-sectional area of opening, m2 air density, kg/m3 pressure difference across opening, Pa
The discharge coefficient CD is a dimensionless number that depends on the geometry of the opening and the Reynolds number of the flow.
Flow Caused by Wind Only Aspects of wind that affect the ventilation rate include average speed, prevailing direction, seasonal and daily variation in speed and direction, terrain, and local obstructions such as nearby buildings, hills, trees, and shrubbery. Liddament (1988) reviewed the relevance of wind pressure as a driving mechanism. A multiflow path simulation model was developed and used to illustrate the effects of wind on air change rate. Average wind speeds may be lower in summer than in winter; directional frequency is also a function of season. Natural ventilation systems are often designed for wind speeds of one-half the seasonal average. Equation (37) shows the rate of air forced through
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2017 ASHRAE Handbook—Fundamentals (SI)
ventilation inlet openings by wind or determines the proper size of openings to produce given airflow rates: Q = Cv AU
(37)
where
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Q = airflow rate, m3/s Cv = effectiveness of openings (Cv is assumed to be 0.5 to 0.6 for perpendicular winds and 0.25 to 0.35 for diagonal winds) A = free area of inlet openings, m2 U = wind speed, m/s
Air intakes should be placed in exterior high-pressure regions, and air reliefs should be placed in exterior low-pressure regions, but because of wind variations, these static locations will, at times, not be optimal. Other considerations include flow control when wind speed is high, and security. Air intakes should face directly into the prevailing wind. If they are not advantageously placed, flow will be less than that predicted by Equation (37); if intakes are unusually well placed, flow will be slightly more. Desirable air relief locations are (1) on the leeward side of the building directly opposite the intake; (2) on the roof, in the low-pressure area caused by flow separation; (3) on a side perpendicular to the windward face, where low-pressure areas occur; (4) in a dormer on the leeward side; (5) in roof ventilators; or (6) by stacks. Chapter 24 gives a general description of the wind pressure distribution on a building, but transient computational modeling will likely be required.
Flow Caused by Thermal Forces Only If building internal resistance is not significant, flow caused by stack effect can be expressed by Q = CDA 2g H NPL T i – T o T i
(38)
Increase in Airflow by Increasing Area of One Opening
When openings are unequal, use the smaller area in Equation (38) and add the increase as determined from Figure 8.
Natural Ventilation Guidelines Several general guidelines should be observed in designing for natural ventilation. Some of these may conflict with other climateresponsive strategies such as using orientation and shading devices to minimize solar gain, with building codes that encourage compartmentalization to restrict fire and smoke movement, or with other design considerations. System selection • In hot, humid climates, use mechanical cooling. If mechanical cooling is not available, air velocities should be maximized in the occupied zones of rooms. • In hot, arid climates, consider evaporative cooling. Airflow throughout the building should be maximized for structural cooling, particularly at night when the outdoor air temperature is low.
where m3/s
Q = airflow rate, CD = discharge coefficient for opening HNPL = height from midpoint of lower opening to NPL, m Ti = absolute indoor temperature, K To = absolute outdoor temperature, K
Equation (38) applies when Ti To. If Ti To, replace Ti in the denominator with To, and replace (Ti – To) in the numerator with (To – Ti). If there is thermal stratification, use an average temperature for Ti. If the building has more than one opening for natural ventilation’s airflow, the relief and intake areas are considered equal. The discharge coefficient CD accounts for all viscous effects such as surface drag and interfacial mixing. Estimating HNPL is difficult for naturally ventilated buildings. If one window or door represents a large fraction (approximately 90%) of the total opening area in the envelope, then the NPL is at the mid-height of that aperture, and HNPL equals one-half the height of the aperture. For this condition, flow through the opening is bidirectional (i.e., air from the warmer side flows through the top of the opening, and air from the colder side flows through the bottom). Interfacial mixing occurs across the counterflow interface, and the orifice coefficient can be calculated by (Kiel and Wilson 1986): CD = 0.40 + 0.0045|Ti – To|
Fig. 8
(39)
If enough other openings are available, airflow through the opening will be unidirectional, and mixing cannot occur. A discharge coefficient of CD = 0.65 should then be used. Additional information on stack-driven airflows for natural ventilation can be found in Foster and Down (1987). The greatest flow per unit area of openings is obtained when the intake and relief areas are equal; Equations (38) and (39) are based on this equality. Increasing the relief area over intake area (or vice versa) increases airflow but not in proportion to the added area.
Building and surroundings characteristics • Topography, landscaping, and surrounding buildings should be used to redirect airflow and give maximum exposure to breezes. Vegetation can funnel breezes and avoid wind dams, which reduce the driving pressure differential around the building. Site objects should not obstruct air intakes’ openings. • The building should be shaped to expose maximum shell openings to breezes. • Architectural elements such as wing walls, parapets, and overhangs should be used to promote airflow into the building interior. • The long façade of the building and the majority of door and window openings should be oriented with respect to prevailing summer breezes. If there is no prevailing direction, openings should be sufficient to provide ventilation regardless of wind direction. Opening locations • Windows should be located in opposing pressure zones. Two openings on opposite sides of a space increase ventilation flow. Openings on perpendicular sides of the building force air to change direction, providing ventilation to a greater area. The benefits of the window arrangement depend on the air relief location relative to the direction of the intake’s airstream. • If a room has only one exterior wall, better airflow is achieved with two widely spaced windows. • If openings are at the same level and near the ceiling, much of the flow may bypass the occupied portion of a room and be less effective in removing contaminants from there.
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Ventilation and Infiltration • Vertical distance between openings is required to take advantage of stack effect; the greater the vertical distance, the greater the ventilation rate. • Openings near the NPL are least effective for thermally induced natural ventilation. If the building has only one large opening, the NPL tends to move to that level, which reduces pressure across the opening.
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Opening characteristics • Greatest airflow per unit area of total opening is obtained by having openings of nearly equal areas. An inlet window smaller than the outlet creates higher inlet velocities. An outlet smaller than the inlet creates lower but more uniform airspeeds through the room. • Openings with areas much larger than calculated are sometimes desirable when anticipating increased occupancy or very hot weather. • In one room, horizontally separated windows are generally better than vertically separated windows. They produce more airflow over a wider range of wind directions and are most beneficial in locations where prevailing wind patterns shift. • Window openings should be accessible to and operable by occupants, unless fully automated. For secondary fire egress, operable windows may be required. • Intake and exhaust openings should not be obstructed by draperies, furniture, or nearby indoor partitions, for example. Partitions can be placed to split and redirect airflow but should not restrict flow between the building’s inlets and outlets. Vertical airshafts or open staircases, where allowable by fire code, can be used to increase and take advantage of stack effects. Enclosed staircases intended for evacuation or safe haven during a fire must not be used for ventilation. Methods and tools have been developed in recent years to move the art of designing natural ventilation systems beyond the application of simple rules of thumb to engineered design (Axley et al. 2002; CIBSE 2005; Emmerich et al. 2012).
Hybrid Ventilation Successful application of purely natural ventilation systems for cooling may be very limited in hot or humid climates, such as in much of the United States, by thermal comfort issues and the need for reliability. However, hybrid (or mixed-mode) ventilation systems or operational strategies offer the possibility of saving energy in a greater number of buildings and climates by combining natural ventilation systems with mechanical equipment (Emmerich 2006). The air-side economizer is one form of hybrid ventilation control scheme, and enjoys wide use in commercial, industrial, and institutional buildings in appropriate climates. The report of the International Energy Agency’s (IEA) Annex 35 describes the principles of hybrid ventilation technologies, control strategies, design and analysis methods, and case studies (Heiselberg 2002). Integrated multizone airflow and thermal modeling is recommended when designing natural and hybrid ventilation systems (Axley 2001a; Dols et al. 2014; Li and Heiselberg 2003).
7.
16.15 pressurization test is relatively quick and inexpensive, and it characterizes building envelope airtightness independent of weather conditions; some jurisdictions now mandate that buildings be tested and rated. In this procedure, a large, variable-flow fan or blower is mounted in a door or window and induces a large, roughly uniform pressure difference across the building shell [ASTM Standards E779 and E1827; Canadian General Standards Board (CGSB) Standard 149.10; ISO Standard 9972]. The airflow required to maintain this pressure difference is then measured. The leakier the building is, the more airflow is necessary to induce a specific indoor/outdoor pressure difference. The airflow rate is generally measured at a series of pressure differences ranging from about 10 Pa to 75 Pa. Depressurization tests are also used. The results of a pressurization test, therefore, consist of several combinations of pressure difference and airflow rate data. An example of typical data is shown in Figure 9. These data points characterize the air leakage of a building and are generally converted to a single value that is reported as the building’s airtightness. There are several different measures of airtightness, most of which involve fitting the data to a curve describing the relationship between the airflow Q through an opening in the building envelope and the pressure difference p across it. This relationship is called the leakage function of the opening. The form of the leakage function depends on the geometry of the opening. Background theoretical material relevant to leakage functions may be found in Chastain et al. (1987), Etheridge (1977), Hopkins and Hansford (1974), Kronvall (1980), and Walker et al. (1997). Openings in a building envelope are usually not uniform in geometry, and wind varies, so generally flow never becomes fully developed. Each opening in the building envelope, however, is often described by Equation (40), commonly called the power law equation: Q = c(p) n
(40)
where Q = airflow through opening, m3/s c = flow coefficient, m3/(s·Pan) n = pressure exponent, dimensionless
RESIDENTIAL AIR LEAKAGE
Most infiltration in U.S. low-rise residential buildings is dominated by unintentional envelope leakage. However, new construction tends toward tighter building envelopes; where mechanical or natural ventilation is not provided, it is possible for a residence to be too “tight,” especially if indoor sources of pollutants are not adequately controlled.
Envelope Leakage Measurement A building’s envelope leakage can be measured with pressurization testing, commonly called a blower-door test. A fan
Fig. 9 Airflow Rate Versus Pressure Difference Data from Whole-House Pressurization Test
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2017 ASHRAE Handbook—Fundamentals (SI)
Sherman (1992a) showed how the power law can be developed analytically by looking at developing laminar flow in short pipes. Equation (40) only approximates the relationship between Q and p. Measurements of single cracks (Honma 1975; Kreith and Eisenstadt 1957) show that n can vary if p changes over a wide range. Additional investigation of pressure and flow data for simple cracks by Chastain et al. (1987) indicated the importance of adequately characterizing the three-dimensional geometry of openings and the entrance and exit effects. Walker et al. (1997) showed that, for the arrays of cracks in a building envelope over the range of pressures acting during infiltration, n is constant. A typical value for n is about 0.65. Values for c and n can be determined for a building by using fan pressurization testing.
Airtightness Ratings In some cases, the predicted airflow rate is converted to an equivalent or effective air leakage area as follows: 2 p r AL = 10 000Qr ------------------------CD
(41)
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where
Conversion Between Ratings Air leakage areas at one reference pressure difference can be converted to air leakage areas at another reference pressure difference by C D ,1 p r , 2 Ar,2 = Ar,1 ----------- ------------- C D , 2 p r ,1
n – 0.5
(42)
where Ar,1 Ar,2 CD,1 CD,2 n
= = = = =
air leakage area at reference pressure difference pr,1, cm2 air leakage area at reference pressure difference pr,2, cm2 discharge coefficient used to calculate Ar,1 discharge coefficient used to calculate Ar,2 pressure exponent from Equation (40)
Air leakage area at one reference pressure difference can be converted to airflow rate at some other reference pressure difference by C D ,1 A r ,1 2 Qr,2 = --------------------- --- (pr,1)0.5 – n(pr,2)n 10 000
(43)
where Qr,2 is airflow rate at reference difference pr,2 , in m3/s.
AL = equivalent or effective air leakage area, cm2 Qr = predicted airflow rate at pr (from curve fit to pressurization test data), m3/s = air density, kg/m3 pr = reference pressure difference, Pa CD = discharge coefficient
Flow coefficient c in Equation (40) may be converted to air leakage area by
All openings in the building shell are combined into an overall opening area and discharge coefficient for the building when the equivalent or effective air leakage area is calculated. Some users of the leakage area approach set CD = 1. Others set CD 0.6, which is the discharge coefficient for a sharp-edged orifice. The air leakage area of a building is, therefore, the area of an orifice with an assumed value of CD that would produce the same amount of leakage as the building envelope at the reference pressure. An airtightness rating, whether based on an air leakage area or a predicted airflow rate, is generally normalized by some factor to account for building size. Normalization factors include floor area, exterior envelope area, and building volume. With the wide variety of possible approaches to normalization and reference pressure difference, and the use of the air leakage area concept, many different airtightness ratings are used. Reference pressure differences include 4, 10, 25, 50, and 75 Pa. Reference pressure differences of 4 and 10 Pa are advocated by researchers because they are closer to the pressure differences that actually induce air exchange and, therefore, better model the opening’s flow characteristics. Although this may be true, they are below the typical range of measured values in the test; therefore, predicted airflow rates at 4 and 10 Pa are subject to significant uncertainty. This uncertainty and its implications for quantifying airtightness are discussed in Chastain (1987), Modera and Wilson (1990), and Persily and Grot (1985b). Round-robin tests by Murphy et al. (1991) to determine the repeatability and reproducibility of fan pressurization devices found that subtle errors in fan calibration or operator technique are greatly exaggerated when extrapolating the pressure versus flow curve down to 4 Pa, with errors as great as ±40%, mainly because of fan and meter calibration errors at low flow. Some common airtightness ratings include the effective air leakage area at 4 Pa assuming CD = 1.0 (Sherman and Grimsrud 1980); the equivalent air leakage area at 10 Pa assuming CD = 0.611 (CGSB Standard 149.10); and the airflow rate at 50 Pa, divided by the building volume to give units of air changes per hour (Blomsterberg and Harrje 1979).
Finally, air leakage area may be converted to flow coefficient c in Equation (40) with
10 000c AL = ------------------- --- pr(n – 0.5 2 CD
CD A L 2 c = ---------------- --- (pr)0.5 – n 10 000
(44)
(45)
Equations (42) to (45) require assumption of a value for n, unless n is reported with the measurement results. When whole-building pressurization test data are fitted to Equation (40), the value of n generally is between 0.6 and 0.7. Therefore, using a value of n in this range is often reasonable.
Building Air Leakage Data Fan pressurization measures a building’s leakage behavior that ideally varies little with time and weather conditions. In reality, unless wind and temperature differences during the measurement period are sufficiently mild, pressure differences induced by weather during the blower door test cause measurement errors. Modera and Wilson (1990) and Persily (1982) studied the effects of wind speed on pressurization test results. Several experimental studies also showed variations of about 20 to 40% over a year in the measured airtightness in the homes studied (Kim and Shaw 1986; Persily 1982; Warren and Webb 1986). Figure 10 summarizes envelope leakage measured in North American housing (Sherman and Dickerhoff 1998) and from several European and Canadian sources (AIVC 1994). This figure shows the large range of measured envelope tightness but also shows typical and extreme values in the housing stock. ASHRAE Standard 62.2 establishes air leakage performance levels for residential buildings. These levels are in terms of effective annual average infiltration rate Qinf , which is based on normalized leakage area NL: ELA H NL = 0.1 ------------- ------ A floor H r
0.4
(46)
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Ventilation and Infiltration
16.17 Air Leakage of Building Components The fan pressurization procedure discussed in the section on Envelope Leakage Measurement allows whole-building air leakage to be measured. The location and size of individual openings in building envelopes are extremely important because they influence the air infiltration rate of a building as well as the envelope’s heat and moisture transfer characteristics. Additional test procedures for pressure-testing individual building components such as windows, walls, and doors are discussed in ASTM Standards E283 and E783 for laboratory and field tests, respectively.
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Leakage Distribution
Fig. 10 Envelope Leakage Measurements where NL = normalized leakage Hr = reference height, 2.5 m H = vertical distance from lowest above-grade floor to highest ceiling, m ELA = effective leakage area, m2, using 4 Pa reference pressure
ELA = (Lpress – Ldepress)/2 where Lpress = leakage area from pressurization, m2 Ldepress = leakage area from depressurization, m2
Dickerhoff et al. (1982) and Harrje and Born (1982) studied air leakage of individual building components and systems. The following points summarize the percentages of whole-building air leakage area that they found associated with various components and systems. Values in parentheses include the range determined for each component and the mean of the range. Walls (18 to 50%; 35%). Both interior and exterior walls contribute to the leakage of the structure. Leakage can occur between the sill plate and the floor or foundation; through cracks below the bottom of the gypsum wallboard, around and through electrical boxes, and through plumbing penetrations; and into the attic at the top plates of walls. Holes drilled through the top plates into the attic for passage of wiring are often unsealed. Ceiling Details (3 to 30%; 18%). Leakage across the top ceiling of the heated space is particularly insidious because it reduces the effectiveness of insulation on the attic floor and contributes to infiltration heat loss. Ceiling leakage also reduces the effectiveness of ceiling insulation in buildings without attics. Recessed lighting, plumbing, and other penetrations leading to the attic are some particular areas of concern, as are intentional openings such as access hatches or for whole-house fans. Forced-Air Heating and/or Cooling Systems (3 to 28%; 18%). The location of the heating or cooling equipment, air handler, or ductwork in conditioned or unconditioned spaces; the venting arrangement of a fuel-burning device; and the existence and location of a combustion air supply all affect air leakage. Modera et al. (1991) and Robison and Lambert (1989), among others, found that the variability of leakage in ducts passing through unconditioned spaces is high, the coefficient of variation being about 50%. Field studies also showed that in situ repairs can eliminate one-quarter to two-thirds of the observed leakage (Cummings and Tooley 1989; Cummings et al. 1990; Jump et al. 1996; Robison and Lambert 1989). The 18% contribution of ducts to total leakage significantly underestimates their effect because, during system operation, pressure differentials across duct leaks are approximately ten times higher than typical pressure differences across envelope leaks (Modera 1989; Modera et al. 1991) and result in large (factors of two to three), changes in ventilation rate (Cummings et al. 1990; Walker 1999; Walker et al. 1999). Windows and Doors (6 to 22%; 15%). More variation in window leakage is seen among window types (e.g., casement versus double-hung) than among new windows of the same type from different manufacturers (Weidt et al. 1979). Windows that seal by compressing the weather strip (casements, awnings) often show significantly lower leakage than windows with sliding seals. Leakage around the frames of windows also can be significant if not properly sealed, typically with controlled-expansion foam, during their installation. The door between a residence and an attached garage, or an enclosed hallway in multiunit housing, also needs gasketing to reduce air leakage as well as smoke and other pollutants’ movement. Fireplaces (0 to 30%; 12%). When a fireplace is not in use, poorly fitting dampers allow indoor air to escape. Glass doors may
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16.18 not seal the fireplace structure more tightly than a closed damper does. Chimney caps or fireplace plugs (with signs that warn they are in place) effectively reduce leakage through a cold fireplace but may not be allowed by fire code. The gap between a metal fireplace insert and the surrounding wall, tiles, or brick is often another leakage path if not properly sealed. Exhaust Vents from Conditioned Spaces (2 to 12%; 5%). Exhaust ducts going from conditioned spaces to the outdoors frequently have either no dampers or dampers that do not close properly. The gap between the duct and the wall or roof penetration often needs sealing, as well. Diffusion Through Walls and Ceilings (1%). Compared to infiltration through holes and other openings in the structure, diffusion through well-constructed walls and ceilings is not an important flow mechanism. At 5 Pa, the permeability of building materials produces an air change rate of less than 0.01 ach by wall diffusion in a typical house. Proper installation of an air-retarding membrane can help minimize this air leakage path. Component Leakage Areas. Individual building component leakage areas vary widely from house to house. Typical variability for an individual component is about a factor of 10, depending on the component’s construction and installation. Testing should establish the installed leakage of a component in critical applications such as energy-efficient manufactured housing.
2017 ASHRAE Handbook—Fundamentals (SI) Existing Buildings. Air leakage paths must first be located before the envelope of an existing building can be tightened. As discussed earlier, air leakage into and out of buildings is caused by not only windows and doors but also a wide range of unexpected and unobvious construction defects. Many important leakage routes can be very difficult to find. A variety of techniques developed to locate leakage paths are described in ASTM Standard E1186 and Charlesworth (1988). Once leakages are located, they can be repaired with materials and techniques appropriate to the size and location of the leak. Diamond et al. (1982), Energy Resource Center (1982), Harrje et al. (1979), and many websites and online videos include information on airtightening or “weatherization” in existing residential buildings with caulking, sealing, weatherstripping, and use of door sweeps, for example. With these procedures, air leakage of residential buildings can be reduced dramatically: anywhere from 5% to more than 50%, depending on the extent of the tightening effort and the experience of those doing the work (Blomsterberg and Harrje 1979; Giesbrecht and Proskiw 1986; Harrje and Mills 1980; Jacobson et al. 1986; Verschoor and Collins 1986). Much less information is available for airtightening large, commercial buildings, but the same general principles apply (Parekh et al. 1991; Persily 1991); the joint between corrugated floor or ceiling decking and the exterior walls is often poorly sealed or not at all, and is often hidden from easy view.
Multifamily Building Leakage Leakage distribution is particularly important in multifamily apartment buildings. These buildings often cannot be treated as single zones because of the internal resistance between apartments. Moreover, leakage between apartments varies widely, from very small for well-constructed buildings with air/moisture retarders and well-sealed wall and floor penetrations between units, to as high as 60% of the total apartment leakage in older, tall apartment buildings of masonry construction (Diamond et al. 1986; Modera et al. 1991). RDH (2013) reviewed the current state of airtightness in multifamily buildings, including testing requirements and techniques, performance targets, and current airtightness levels.
Controlling Air Leakage New Buildings. It is much easier and more cost-effective to build a tight building than to tighten an existing building. Elmroth and Levin (1983), Eyre and Jennings (1983), Marbek Resource Consultants (1984), and Nelson et al. (1985) provide information and construction details on airtight building design for houses. The economic paybacks for basic “weatherization” improvements are often the most attractive of all energy conservation measures in buildings where no actions have been taken previously. A continuous air infiltration retarder, formerly known as an air barrier, can be an effective way to reduce air leakage through walls, around window and door frames, and at joints between major building elements. Take particular care to ensure its continuity at all wall, floor, and ceiling joints; at window and door frames; and at all penetrations of the retarder such as electrical boxes, plumbing connections, and utility service penetrations. Joints in the air/vapor retarder must be lapped and sealed. Plastic vapor retarders installed in the ceiling should be tightly sealed to the vapor retarder in the outer walls and should be continuous over the interior walls. A seal at the top of the interior walls reduces leakage into the attic; the plate on top of the studs generally gives a poor seal. In cold climates, the air infiltration retarder can be installed either on the inside of the wall framing, in which case it usually functions as a vapor retarder as well, or on the outside of the wall framing, in which case it should have a permeance rating high enough to allow diffusion of water vapor from the wall. For a discussion of moisture transfer in building envelopes, see Chapters 25 and 26.
8.
RESIDENTIAL VENTILATION
Typical infiltration rates in housing in North America vary by an order of magnitude, from tightly constructed housing with seasonal average air change rates as low as 0.1 h–1 to loosely constructed housing with air change rates as great as 2.0 h–1 or even higher in a few cases. Figures 11 and 12 show histograms of infiltration rates measured in two different samples of North American housing (Grimsrud et al. 1982; Grot and Clark 1979). Figure 11 shows the average seasonal infiltration of 312 houses located in different areas in North America. The median infiltration value of this sample is 0.5 h–1. Figure 12 represents measurements in 266 houses located in 16 U.S. cities. The median value of this sample is 0.9 h–1. The group of houses in the Figure 11 sample is biased toward then-new, then “energy-efficient” houses, whereas the group in Figure 12 represents older, low-income housing. New houses constructed since these studies likely are somewhat tighter because of greater awareness of leakage paths and energy efficiency efforts, including mandated leakage testing in some locations. A modeling study using a set of 209 dwellings that represent 80% of U.S. housing stock estimated a median value of infiltration of 0.44 h–1 for all single-family houses, whereas the median for those built since 1990 was 0.26 h–1 (Persily et al. 2010). Additional studies have found average values for houses in regional areas. Palmiter and Brown (1989) and Parker et al. (1990) found a heating season average of 0.40 h–1 (range: 0.13 to 1.11 h–1) for 134 houses in Pacific Northwest climates. In a comparison of 292 houses incorporating energy-efficient features, including measures to reduce air infiltration and provide ventilation heat recovery, with 331 control houses, Parker et al. (1990) found an average of about 0.25 h–1 (range: 0.02 to 1.63 h–1) for the energy-efficient houses versus 0.49 h–1 (range: 0.05 to 1.63 h–1) for the control group. Ek et al. (1990) found an average of 0.5 h–1 (range: 0.26 to 1.09 h–1) for 93 double-wide manufactured homes in the Pacific Northwest. Canadian housing stock has been characterized by Riley (1990) and Yuill and Comeau (1989). Although these studies do not represent random samples of North American housing, they indicate the distribution of infiltration rates expected in a group of buildings. The influence of occupants on infiltration has not been measured directly and varies widely. These influences are, for example,
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Ventilation and Infiltration operation of doors, windows, and small exhaust fans. Desrochers and Scott (1985) estimated that they add an average of 0.10 to 0.15 h–1 to the unoccupied values. Kvisgaard and Collet (1990) found that, in 16 Danish dwellings, occupants on average provided 63% of the total air change rate. Ventilation air for houses in the United States has traditionally been provided with the assumption occupants will use windows and exhaust fans when needed, and that the building envelope is leaky enough so that infiltration will suffice. Possible difficulties with this approach include occupant error, low infiltration when natural forces (temperature difference and wind) are weak; unnecessary energy consumption when these forces are strong; drafts in cold climates; lack of control of ventilation rates to meet changing needs; poor humidity control; potential for interstitial condensation from exfiltration in cold climates or infiltration in hot humid climates; and lack of opportunity to recover energy used to condition ventilation air. The solution to these concerns is to have a tight building envelope and a properly designed and operated mechanical ventilation system with automatic control. ASHRAE Standard 62.2 and the National Building Code of Canada (NRCC 2010) encourage tighter envelope construction. Hamlin (1991) found a 30% increase in airtightness of tract-built Canadian houses between 1982 and 1989. Also, 82% of newer houses had natural air change rates below 0.3 h–1 in March. Yuill (1991) derived a
Fig. 11 Histogram of Infiltration Values for Then-New Construction
Fig. 12 Histogram of Infiltration Values for Low-Income Housing
16.19 procedure to show the extent to which infiltration contributes toward meeting ventilation air requirements. As a result, the National Building Code of Canada has requirements for mechanical ventilation capability in all new dwelling units. Canadian Standards Association (CSA) Standard F326 expands the requirements for residential mechanical ventilation systems to include air distribution within the house, thermal comfort, minimum temperatures for equipment and ductwork, system controls, pressurization and depressurization of the dwelling, installation requirements, and verification of compliance. Verification can be by design or by test, but the total rate of outdoor air delivery must be measured. Mechanical ventilation is required by ASHRAE Standard 62.2 and by building code in some U.S. states; some details of these requirements are described in this chapter. The net benefit of using controlled mechanical ventilation has been demonstrated and studied in various energy-efficient and advanced housing programs (Barley 2001; Palmiter et al. 1991; Riley 1990). Systems can be characterized as local or central; exhaust, supply, or balanced; with forced-air or radiant/hydronic heating/cooling systems; with or without heat recovery; and with continuous operation or with control by occupants, pollutant sensing, timers, or humidity. Note that not all combinations are viable. Various options are described by Fisk et al. (1984), Hekmat et al. (1986), Holton et al. (1997), Lubliner et al. (1997), Palmiter et al. (1991), Reardon and Shaw (1997), Sherman and Matson (1997), Sibbitt and Hamlin (1991), and Yuill et al. (1991). The simplest systems use bathroom and kitchen fans to exhaust moisture and pollutants outdoors, and to depressurize to augment infiltration. Noise, installed capacity, durability under high or continuous operation, air movement from all rooms (especially bedrooms), envelope moisture, combustion safety, and energy efficiency issues need to be addressed. Many present bath and kitchen fans are ineffective ventilators because of poor installation and design, and many fail to exhaust outdoors. However, properly specified and installed exhaust fans can form part of good whole-house ventilation systems and are so specified in many building codes. Some systems use their central air-handling unit blower to induce air from the outdoors and distribute it. However, the blower operates intermittently if thermostatically controlled and thus provides little ventilation in mild weather. Continuous blower operation increases energy consumption. If the blower operates continuously when the heat source is off, the combination of lower mixed-air temperature and high air speed can cause cold air drafts when in heating mode. To offset these problems, some systems use variable-speed blowers, which allow operation at lower speeds during mild weather. Others use a timer to cycle the blower when thermostatic demands are inadequate for ventilation purposes (Rudd 1998). Central exhaust systems use leakage paths and, in some cases, intentional, controllable, and filtered openings in the building envelope for admitting makeup air. Such systems may be suitable for retrofit in existing houses. Energy can be recovered from the exhaust airstream to precondition makeup air or to warm potable water using a heat pump, for example. For new houses with tightly constructed envelopes, balanced ventilation with passive heat recovery (e.g., air-to-air heat exchangers, heat recovery ventilators) can be appropriate in some climates. Fan-induced, filtered outdoor and exhaust airflow at nearly equal rates through a heat exchanger, where heat and sometimes moisture is transferred between the airstreams. This typically reduces the energy required to condition ventilation air by 60 to 80% (Cutter 1987). Concerns associated with these systems include airflow balance, leakage between streams, biological contamination of wet surfaces, frosting, and initial and maintenance costs. Air-side economizers, which allow outdoor air to be up to 100% of the supply air at appropriate times, are not typically used in
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16.20 small buildings with low internal heat gains relative to the building envelope. Because of heat transfer through building envelopes, these small buildings quickly require heating or cooling as the outdoor air temperature falls or rises. Consequently, from an energy conservation point of view, small envelope-load-dominated buildings do not benefit as much as internal-load-dominated buildings in cool climates from daytime use of air-side economizers; night ventilation of residences during the cooling season may be very attractive, though, when the outdoor humidity ratio is low. Ventilation rates increase dramatically when air-side economizers are in operation, so the extra moisture introduced or removed must be considered. The type of ventilation system can be selected based on house leakage class as defined in ASHRAE Standard 119. Balanced airto-air systems with heat recovery are optimal for tight houses (leakage classes A to C in the standard). The leakier the house, the larger the contribution from infiltration and the less effective heat recovery ventilation will be. Tightening the envelope beyond the level of ASHRAE Standard 62.2 may be warranted in extremely cold climates to better use the heat recovery effect (Sherman and Matson 1997). In mild climates, these systems can also effectively be used in leakage classes D to F. Central exhaust systems should not be used for leakage classes A to C unless special air intakes are provided; otherwise their operation may depressurize the house enough to cause backdrafting through fossil-fueled appliances without closed-combustion systems. Unbalanced systems (either supply or exhaust) are optimal for leakage classes D to F. Ventilation systems are normally not needed for leakage classes G to J, but when they are needed, an unbalanced system is usually the best choice. More discussion of mechanical systems for residences is available in Russell et al. (2005); some information on practices outside North America can be found in McWilliams and Sherman (2005).
Shelter in Place The most fundamental function of a house is to provide shelter from outdoor conditions. A first response to poor outdoor air quality is to go indoors. Closing windows and other air intakes and turning off exhaust fans reduces air exchange with the outdoors, decreasing the immediate intrusion of outdoor air into the home. However, because no home is perfectly airtight, closing doors and windows does not eliminate intrusion. Because all indoor air ultimately comes from outdoors, indoor air quality eventually comes to dynamic equilibrium with outdoor conditions in an idealized, unoccupied building. The tighter the building, the longer the time needed to come to equilibrium. The delay time (the time it takes to completely change the air in a building) is determined by the ventilation rate. The effectiveness of sheltering within the home thus depends on envelope tightness. For a home with 0.35 air changes per hour, the delay time is roughly 3 h. For a very tight house without mechanical ventilation, the delay time can easily be twice as long. Most houses in the United States are leakier and thus could have a delay time of about one hour (Sherman and Matson 1997). Reactive gases in outdoor air, such as ozone, can be decreased to some degree by the building envelope. For other outdoor contaminants, the building envelope serves to delay, not reduce, their introduction into the indoor environment. Such a delay is not very helpful at reducing exposures to outdoor contaminants that persist over days, but can be an effective strategy for short-duration (e.g., less than a few hours) sources. In houses without indoor sources, ozone levels tend to be higher in houses that do not have air conditioners than in those with air conditioners; ozone levels also are higher when windows are open than when they are closed (Weschler 2000). For outdoor exposure times shorter than the delay time, the house serves as a reservoir of cleaner air of that particular pollutant. After the outdoor contaminant is gone, windows can be opened to
2017 ASHRAE Handbook—Fundamentals (SI) flush out pollutants that entered or were internally generated during the exposure period.
Safe Havens Simply going indoors may not be sufficient for highly unusual but potentially lethal events. Chemical spills or fires, explosions, bioterrorism, or similar toxic air pollutant releases can temporarily create dangerous outdoor conditions that render other air quality issues insignificant. With sufficient warning, occupants should leave the vicinity, but the unexpected nature of these events means that the only viable alternative may be to shelter in place. This strategy may work for short-term releases of airborne toxins. Homes are often too leaky to provide the protection needed for longer-duration events, but individual rooms can be temporarily sealed to become safe havens. A safe haven should be chosen to have as little contact as possible with outer walls, and preferably be on the side of the house farthest downwind from the source. Impermeable tape can be used to seal leaks, cracks, seams, register grilles, and doors, with thick plastic sheeting used to span larger gaps (Sorensen and Vogt 2001). If such a shelter has an air change rate of 0.15 h–1 with the house, it can take 4 to 6 h for contaminated outdoor air to affect significantly the safe haven. With very well-sealed, occupied spaces, the air quality inside degrades because of the occupants and materials within, especially rapidly for smaller spaces with higher occupant density. Suffocation is remotely possible, but thermal stress and odor accumulation are more typical. Where emergencies are somewhat more likely, engineered safe rooms with adequate HVAC systems are needed. In such cases, a safe haven can be designed in advance with thermal conditioning and a highly efficient particle/gas-phase filtration system capable of providing several hours, days, or weeks of protection (Ormerod 1983). A short-term safe haven might be effectively combined with another emergency shelter (e.g., tornado, hurricane, civil defense) to reduce cost.
9.
RESIDENTIAL IAQ CONTROL
ASHRAE Standard 62.2 presents minimum requirements for residential ventilation air and acceptable indoor air quality, and its user’s manual (ASHRAE 2010a) has detailed information for designing and constructing residential buildings in compliance with the standard. Best or good practice may require going beyond the standard’s minima. This section describes good practice; however, it presumes that the minimum requirements of 62.2 are met as well. Traditionally, ventilation air for residences has been provided by natural ventilation controlled by occupants, and also mechanically by increasing infiltration using exhaust fans vented outdoors. Sherman and Matson (1997) showed that many older residential buildings are leaky enough that infiltration alone can meet the minimum requirements of ASHRAE Standard 62.2. Houses built or retrofitted to new standards have substantially tighter envelopes and insufficient infiltration under most weather conditions to meet ventilation standards. Studies show that concerns over safety, noise, comfort, air quality, and energy minimize occupants' use of operable windows (Johnson and Long 2005; Price and Sherman 2006). As a result, these houses require supplemental mechanical ventilation to satisfy current standards. The minimum residential ventilation rate is not always sufficient to adequately dilute all contaminants (e.g., radon), and can introduce undesired substances (e.g., pollen). In these cases, source control, local exhaust, extra ventilation, or more effective air treatment is required to manage contaminant levels. Therefore, especially in single-family dwellings, occupants must be responsible for monitoring and controlling contaminant sources in and around their indoor environments, as well as for operating their dwelling units to meet their individual needs. Increasingly, residences are also used
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Ventilation and Infiltration for business or hobby purposes, which may introduce air contaminants not addressed in Standard 62.2; portions of these residences may require ventilation air as required by Standard 62.1 or industrial guidelines.
Source Control
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When considering how much whole-house ventilation should be supplied, both typical and unusual sources of indoor pollution should be controlled first to limit the required ventilation rate. This can be done either by mitigating the source itself or by using local exhaust to extract contaminants before they can mix into the indoor environment. Typical sources that should be considered include the following. Clothes Dryers and Central Vacuum Systems. Clothes dryer exhaust is heavily laden with moisture and laundry by-products such as flammable lint and various gaseous contaminants. Many moisture problems have been traced to clothes dryers vented indoors or to attics, crawlspaces, garages, or other inappropriate locations. Exhaust from clothes dryers, which is typically about 70 L/s, must be vented directly to the outdoors. Similarly, central vacuum systems must be vented directly outdoors to exhaust the finer particles that pass through their filters. Combustion. Water and carbon dioxide are always produced during combustion of hydrocarbons in air. Dangerous compounds are created, as well. All these products of combustion must be vented directly outdoors, preferably using sealed combustion or direct-vent equipment. Venting must meet or exceed all applicable codes. For buildings with naturally aspirated combustion appliances, excessive depressurization of the building by exhaust systems must be avoided to eliminate backdrafting. Consider a depressurization safety test, such as described in ASTM Standard E1998 or CGSB Standard 51.71. Fireplace combustion products should be isolated from the occupied space using tight-fitting doors and outdoor air intakes, when necessary. Flues and chimneys must be designed and installed to disperse combustion products well away from air intakes and operable windows, for example. Chapter 35 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment has more information on venting systems. Exhaust hoods and fans for ranges and cooktops must vent outdoors. Carbon monoxide is one of the most unwelcome indoor contaminants, and in significant concentration, poses an imminent threat to life. It can come from virtually any source of combustion, including from motorized vehicles and generators (Emmerich et al. 2013). Even combustion appliances that meet manufacturers’ specifications can interact with the building and emit carbon monoxide. At least one carbon monoxide alarm meeting safety standards such as CSA Standard 6.19 should be installed near sleeping areas in each dwelling, including each unit of multifamily residential buildings, that has combustion appliances (e.g., fireplaces, stoves, furnaces, water heaters) within the pressure boundary, or has attached garages or storage sheds. Carbon monoxide alarms also should be considered for nonresidential buildings: poisonings have occurred in many building types, including hotels, motels, stores, restaurants, nursing homes, dormitories, laundromats, and schools. Garages. Garages and storage spaces often contain many sources of contaminants. Doors between them and occupied spaces must be well sealed with gaskets or weatherstripping, including their sills, and possibly be self closing. Depressurized sections of HVAC systems for the regularly occupied portions of residences, such as the systems’ air handlers or return or intake ducts, should not be located in garages. If such sections must pass through garages, they must be well sealed and never have registers or grilles installed in the garages or storage spaces; if such spaces need conditioning, they must use their own isolated systems. Take care to ensure that there is a good pressure barrier between the garage and the occupied
16.21 space, typically using an air/moisture retarder such as heavy polyethylene as well as excellent sealing of the door between. Carbon monoxide sources may be present in garages, so pressure barriers, fire-rated compartmentation, and ventilation of attached residences are life-safety measures. Separate ventilation systems that slightly depressurize attached garages and storage spaces and that exhaust directly outdoors should be considered, especially when these support spaces are tightly constructed or are in cold climates. Several studies (Batterman et al. 2006; Emmerich et al. 2003; Fugler 2004) of contaminant sources and transport in garages found that, in some cases, significant fractions of infiltration air enter houses from attached garages, and that modern residential garages are tighter than older garages to reduce energy consumption and to improve comfort; garages were previously assumed to be leaky enough to avoid many related IAQ problems. Particulates. The ventilation system should be designed such that return and any outdoor air is well filtered before passing through the thermal conditioning components of the HVAC system. Pressure drops associated with this filtration must be considered in the design and installation of the air-handling system. Particulate filters or air cleaners should have a minimum efficiency of 60% for 3 m particles, which is equivalent to a MERV 6 designated filter according to ASHRAE Standard 52.2. Microbiologicals. Because ventilation can increase the source as well as removal rates of various air pollutants, it is, at best, moderately effective at reducing exposures to many airborne microbiologicals. Ventilation can, however, be part of the moisture balance that is critical to retarding fungal growth on surfaces and spores released into the air, depending on indoor and outdoor conditions. Radon and Other Soil Gases. Buildings are exposed to gases and water that migrate through the soil into occupied spaces through cracks or other leaks. Soil gases vary with time and conditions, and can contain toxins from pesticides, landfills, fuels, or sewers, but the highest-profile pollutant in this category is radon and its radioactivedecay-produced “daughters.” Source control measures, such as differential pressure control and ground-to-building airtightening, are far more effective than ventilation mechanisms at controlling exposure to soil gas in buildings. See Chapter 11 for more information. Volatile Organic Compounds (VOCs). VOCs are ubiquitous in modern life. Examples of products that emit VOCs include manufactured wood products, paints, stains, varnishes, solvents, pesticides, adhesives, wood preservatives, waxes, polishes, cleansers, lubricants, sealants, dyes, air fresheners, fuels, plastics, copy machines, printers, tobacco products, perfumes, cooking by-products, and dry-cleaned clothes. Whenever possible, VOCs and other toxic compounds should be stored outside the occupied space in loosely constructed or ventilated enclosures such as garden sheds or detached garages, and away from occupied buildings’ ventilation intakes and operable windows. When unusual amounts of such compounds are present, consider using additional ventilation should be considered (e.g., extra ventilation during and for a while after construction projects). Reduced-VOC-emitting products are often available. Outdoor Air. Outdoor air may at times contain unacceptably high levels of pollutants, including ozone, pollen, carbon monoxide, particulate matter, odors, or toxic agents. At such times, it may be impossible to provide acceptable indoor air quality using solely outdoor air, and increased ventilation rates can actually decrease indoor air quality. In locations where this problem may be anticipated, provide automatic or manual controls to allow temporary reduction of the ventilation rate. If these events are frequent, consider a means to more effectively clean recirculated air during these times.
Local Exhaust After source elimination, the single most important source control mechanism in dwellings is local exhaust. Kitchens, utility
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16.22 rooms, bathrooms, and other spaces designed to allow specific contaminant releases should be provided with local exhaust that is vented to the outdoors. Workshops, recreation rooms, smoking areas, art studios, greenhouses, and hobby rooms may also require local ventilation and/or air cleaning to remove contaminants generated by the activities involved. Contaminants of concern should be evaluated to determine how much additional ventilation is required. Many of these rooms can be adequately ventilated by following the requirements for kitchens or bathrooms. If unvented combustion appliances must be used, rooms with these appliances should also meet general ventilation requirements for kitchens, because such appliances generate significant amounts of moisture and, often, ultrafine particles, even when operating properly. Mechanical exhaust (i.e., a fan with ducting to the outdoors) with adequate makeup air is the preferred method for providing local ventilation. Normally, such an exhaust system is designed to operate intermittently under manual control to send contaminated air outdoors when occupants recognize the need for ventilation. However, in many circumstances, a continuous, lower-flow-rate exhaust can work as well. Specific provisions for intermittent versus continuous exhaust rates are typically included in ventilation standards such as ASHRAE Standard 62.2; some example information is provided here. Continuous Local Mechanical Exhaust. A continuously operating mechanical exhaust system is intended to operate without occupant intervention. This exhaust may be part of a balanced whole-building mechanical ventilation system. The system should be designed to operate during all hours in which the dwelling is occupied. Override control, to provide higher exhaust rates when needed, should be provided. The minimum delivered ventilation should be at least that given in Table 1. Intermittent Local Mechanical Exhaust. An intermittently operating local mechanical exhaust is intended to be operated as needed by the occupant. Shutoff timers, occupancy controls, multiple-speed fans, and switching integral with room lighting are helpful, provided they do not impede occupant control. The minimum airflow rate should be at least that given in Table 2. Alternatives. Cleaning recirculated air can sometimes be substituted for local exhaust, if it can be shown to be effective and reliable in removing contaminants of concern. Natural ventilation is not generally a suitable method for local exhaust and ventilation air needs in many climates and spaces. Using natural ventilation can cause reentrainment problems when contaminated exhaust or exfiltrating air reenters the building. In milder climates, natural ventilation may be acceptable when the contaminant of concern is related to odor rather than health or safety. Purpose-designed passive exhaust systems have shown acceptable ventilation in some European settings, and may be considered in lieu of mechanical systems. Axley (2001b) discusses evaluation and design of passive residential ventilation systems further.
Whole-House Ventilation Although control of significant sources of pollution in a dwelling is important, whole-house ventilation through centrally introduced, conditioned, and distributed outdoor air may still be needed. Each dwelling should be provided with outdoor air according to code adopted by the local jurisdiction; general guidance according to ASHRAE Standard 62.2 is given in Table 3 of this chapter. The rate is the sum of the Area-Based and Occupancy-Based columns. Design occupancy expectations in the United States can be based on the number of bedrooms as follows: first bedroom, two persons; each additional bedroom, one person. Additional ventilation should be considered when occupant densities exceed one person per 25 m3. Natural whole-house ventilation that relies on occupant operation should not be used to make up any part of the minimum total whole-house ventilation air requirement. However, because
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Continuous Exhaust Airflow Rates Application
Airflow Rate
Enclosed kitchen Restroom
5 ach 10 L/s
Notes Based on kitchen volume Not less than 2 ach
Table 2 Intermittent Exhaust Airflow Rates Application
Airflow Rate
Kitchen
50 L/s 150 L/s
Restroom
25 L/s
Notes With a vented range hood With other exhaust fans, including downdraft Not less than 2 ach
Table 3 Total Ventilation Air Requirements Area Based
Occupancy Based
0.15 L/s per square metre of floor space
3.5 L/s per person, based on normal occupancy
occupancy and sources vary significantly, the capacity to ventilate above minimum rates can be provided by operable exterior openings such as doors and windows.
Air Distribution Ventilation air should be provided to each habitable room through mechanical and natural air distribution. If a room does not have a balance between air supply and return or exhaust, pathways for transfer air should be provided. These pathways may be door undercuts, transfer ducts with grilles, or simply grilles where ducts are not necessary; however, building codes must be checked, because compartmentation requirements for limiting fire and smoke movement may restrict transfer air means or require rated fire and smoke dampers, for example. In houses without central air handlers, special provisions to distribute outdoor air may be required. Rooms in which occupants spend many continuous hours, such as bedrooms, may require special consideration. Local and whole-house ventilation equipment should be chosen to be energy efficient, easy to maintain, reliable, durable, and quiet. Consider using heat recovery, especially in cold climates.
Selection Principles for Residential Ventilation Systems Occupant comfort, energy efficiency, ease of use, service life, initial and life-cycle costs, value-added features, and indoor environmental quality should be considered when selecting a strategy and system. Poorly designed, installed, or maintained HVAC and related systems can be potential contributors to poor indoor air quality. For example, occupants may not use the ventilation systems as intended if operation results in thermal or acoustical discomfort or excessive energy use. The resulting lack of ventilation might produce poor indoor air quality. Therefore, careful design, construction, commissioning, operation, and maintenance is necessary to provide optimum effectiveness. All exhaust, supply, or air handler fans have the potential to change the pressure of an indoor space relative to the outdoors. High-flow fans, such as in the air handler and with some cooking exhausts, can cause significant depressurization, particularly in tightly constructed homes. Considering these effects is essential in design. Excessive depressurization of the living space relative to outdoors may cause backdrafting of combustion appliances and increased migration of contaminants such as radon or other soil gases, car exhaust, or insulation particles into the living space. Depressurization can also result in increased moisture intrusion into building cavities in warm, moist climates and may cause more structural damage and fungal growth. Pressurization of the living space can cause condensation in building cavities in cold climates, also resulting in structural damage. Excess pressure can best be prevented by balancing
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Ventilation and Infiltration ventilation and exhaust systems and tightly sealing duct systems. In addition, adequate pathways must be available for all return air to the air-handling devices. Occupant activities, operation of exhaust fans, and leaky ducts may depressurize the structure. Options to address backdrafting concerns include
and predicted house infiltration rates for these and other models. The average long-term differences between measurements and predictions are generally about 40%, although individual predictions can be inaccurate by 100% or more (Persily 1986; Walker and Wilson 1998).
• Using combustion appliances with sealed combustion systems • Locating combustion appliances in a ventilated room isolated from depressurized zones by well-sealed partitions • Providing supply or makeup air to balance exhaust from the zone • Testing to ensure that depressurization is not excessive
Multicell models of airflow in buildings treat buildings as a series of interconnected zones and assume that air within each zone is well mixed. Several such models have been developed by Allard and Herrlin (1989), Etheridge and Alexander (1980), Feustel and Raynor-Hoosen (1990), Herrlin (1985), Liddament and Allen (1983), Walton (1984, 1989), and Walton and Dols (2003). They are all based on a mass balance for each zone of the building. These mass balances are used to solve for interior static pressures in the building by requiring that inflows and outflows for each zone balance to zero. The user must provide information describing building envelope leakage, values to account for wind pressure on the building envelope, temperatures for each zone, and any mechanical ventilation airflow rates. Wind pressure coefficient data in the literature, air leakage measurement results from the building or its components, and air leakage data from research can be used as estimates. These models not only solve for whole-building and individual zone air change rates, but also determine airflow rates and pressure differences between zones. These interzone airflow rates are useful for predicting pollutant transport within buildings with well mixed zones. Chapter 13 has more details on multizone airflow and IAQ modeling.
The system must also be designed, built, operated, and maintained in a way that discourages growth of biological contaminants. Typical precautions include sloping condensate drain pans toward the drain, keeping condensate drains free of obstructions, keeping cooling coils free of dirt and other obstructions, maintaining humidifiers, and checking for and eliminating any cause of moisture inside ducts. Additional information on residential ventilation system design and IAQ control may be found in ASHRAE Guideline 24.
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16.23
10. SIMPLIFIED MODELS OF RESIDENTIAL VENTILATION AND INFILTRATION This section describes several calculation procedures, ranging from simple estimation techniques to more physical models. Orme (1999) provides a more thorough review of simplified models. For energy modeling or code compliance, a particular building’s air change rate cannot be reliably deduced from the building’s construction or age, or from a simple visual inspection. Some measurement is necessary, such as a pressurization test of envelope airtightness or a detailed quantification of the leakage sites and their magnitude. The air change rate of a building may be calculated given (1) the location and leakage function for every opening in the building envelope and between major building zones, (2) the wind pressure coefficients over the building envelope, and (3) any mechanical ventilation airflow rates. These data are generally unavailable for all except very simple structures or extremely well studied buildings. Therefore, their values must be assumed. The appropriateness of these assumptions influences the accuracy of predictions of air change rates.
Empirical Models These models of residential infiltration are based on statistical fits of infiltration rate data for specific houses. They use pressurization test results to account for house airtightness and take the form of simple relations between infiltration rate, an airtightness rating, and, in most cases, weather conditions. Empirical models account for envelope infiltration only and do not include intentional ventilation. In one approach, the calculated air change rate at 50 Pa based on a pressurization test is simply divided by a constant approximately equal to 20 (Sherman 1987). This technique does not account for the effect of infiltration-driving mechanisms on air exchange. Empirical models that do account for weather effects have been developed by Kronvall (1980), Reeves et al. (1979), and Shaw (1981). The latter two models account for building air leakage using the values of c and n from Equation (40). The only other data required are wind speed and temperature difference. These empirical models predict long-term (one-week) infiltration rates very well in the houses from which they were developed; they do not, however, work as well in other houses because of the building-specific nature of leakage distribution, wind pressure, and internal partitioning. Persily (1986) and Persily and Linteris (1983) compared measured
Multizone Models
Single-Zone Models Several procedures have been developed to calculate building air change rates that are based on physical models of the building interior as a single zone. These single-zone models are only appropriate for buildings with minimal internal resistance to airflow, and are therefore inappropriate for large, multizone buildings. Some models of this type have been developed by Cole et al. (1980), Sherman and Grims-rud (1980), Walker and Wilson (1998), and Warren and Webb (1980). The section on Residential Calculation Examples uses both basic and enhanced models (Bradley 1993; CHBA 1994; Hamlin and Pushka 1994; Palmiter and Bond 1994; Walker and Wilson 1998). The basic model uses effective air leakage area AL at 4 Pa, which can be obtained from a whole-building pressurization test. The enhanced model uses pressurization test results to characterize house air leakage through leakage coefficient c and pressure exponent n. The enhanced model improves on the basic model by using a power law to represent envelope leakage, including a flue as a separate leakage path, and having separate wind effects for houses with crawlspaces or basements instead of slab-on-grade foundations. For both models, the user must provide the wind speed, the temperature difference, information on the distribution of leakage over the building envelope, a wind shelter parameter or local shielding for the basic model, and a terrain coefficient. The predictive accuracy of the enhanced model can be very good, typically ±10% when parameters are well known for the building in question (Palmiter and Bond 1994; Sherman and Modera 1986; Walker and Wilson 1998). All models’ results are sensitive to the user-provided data, and some of these parameters are often quite difficult to determine.
Superposition of Wind and Stack Effects Simplified physical models of infiltration solve the problem of two natural driving forces (wind and stack) separately and then combine them in a process called superposition. Superposition is necessary because each physical process can affect internal and external air pressures on the building’s envelope, and can cause interactions between physical processes that are otherwise independent. An exact
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2017 ASHRAE Handbook—Fundamentals (SI)
solution is impossible, because (1) detailed properties of all the building leaks are unknown and (2) leakage is a nonlinear process. Therefore, most modelers have developed a simplified superposition process to combine stack and wind effects. Sherman (1992b) compared various superposition procedures and derived a generalized superposition equation involving simple leakage distribution parameters, and showed that the result is always subadditive. Typically, only 35% of infiltration from the smaller effect can be added to the larger effect. Depending on the specific building and conditions, that percentage could go as high as 85% or as low as zero. Walker and Wilson (1993) compared several superposition techniques to measured data. Sherman, as well as Walker and Wilson, found quadrature, shown in Equation (47), to be a robust superposition technique: Q =
2
2
• X = 0 (equal amounts of leakage in the floor and ceiling) • Heights of one-, two-, and three-story buildings = 2.5, 5.0, and 7.5 m, respectively Example 2. Estimate the infiltration at winter design conditions for a twostory house in Lincoln, Nebraska. The house has effective air leakage area of 500 cm2 and volume of 340 m3, and the predominant wind is perpendicular to the street (shelter class 3). The indoor air temperature is 20°C. Solution: The 99% design temperature for Lincoln is –19°C. Assume a design wind speed of 6.7 m/s. From Equation (48), with Cs = 0.000 290 from Table 4 and Cw = 0.000 231 from Table 6, the airflow rate caused by infiltration is 500 2 Q = ------------ 0.000 290 39 + 0.000 231 6.7 1000 = 0.0736 m3/s = 265 m3/h
(47)
Qs + Qw
The following sections discuss how superposition is combined with calculation of wind and stack flows to determine total flow.
From Equation (2), air change rate I is equal to Q divided by the building volume: I = (265 m3/h)/340 m3 = 0.78 h–1 = 0.78 ach
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Residential Calculation Examples Basic Model. The following calculations are based on the Sherman and Grimsrud (1980) model, which uses the effective air leakage area at 4 Pa. This leakage area can be obtained from a whole-building pressurization test. Using effective air leakage area, the airflow rate from infiltration is AL Q = ------------ C s T + C w U 2 1000
(48)
Example 3. Predict the average infiltration during a one-week period in January for a one-story house in Portland, Oregon. During this period, the average indoor/outdoor temperature difference is 17 K, and average wind speed is 2.7 m/s. The house has volume of 255 m3 and effective air leakage area of 690 cm2, and it is located in an area with buildings and trees within 10 m in most directions (shelter class 4). Solution: From Equation (48), the airflow rate caused by infiltration is 690 2 Q = ------------ 0.000 145 17 + 0.000 104 2.7 1000 = 0.0392 m3/s = 141 m3/h
where airflow rate, m3/s effective air leakage area, cm2 stack coefficient, (L/s)2/(cm4 ·K) average indoor-outdoor temperature difference for time interval of calculation, K Cw = wind coefficient, (L/s)2/[cm4 ·(m/s)2] U = average wind speed measured at local weather station for time interval of calculation, m/s Q AL Cs T
= = = =
Table 4 shows values of Cs for one-, two-, and three-story houses. The value of wind coefficient Cw depends on the local shelter class of the building (Table 5) and the building height. Table 6 shows values of Cw for one-, two-, and three-story houses in shelter classes 1 to 5. In calculating values in Tables 4 and 6, the following assumptions were made: • Terrain used for converting meteorological to local wind speeds is that of a rural area with scattered obstacles • R = 0.5 (half the building leakage in the walls) Table 4
Basic Model Stack Coefficient Cs House Height (Stories)
Stack coefficient
Example 4. Estimate the average infiltration over the heating season in a two-story house with volume of 310 m3 and leakage area of 848 cm2. The house is located on a lot with several large trees but no other close buildings (shelter class 3). Average wind speed during the heating season is 3.2 m/s, and the average indoor/outdoor temperature difference is 20 K. Solution: From Equation (48), the airflow rate from infiltration is 848 2 Q = ------------ 0.000 290 20 + 0.000 231 3.2 1000 = 0.077 m3/s = 276 m3/h The average air change rate is therefore I = 276/310 = 0.89 h–1 = 0.89 ach
Enhanced Model. This section presents a simple, single-zone approach to calculating air infiltration rates in houses based on the Walker and Wilson (1998) model. The airflow rate from stack and wind infiltration is
Two
Three
Qs = cCs T n
(49)
0.000 145
0.000 290
0.000 435
Qw = cCw (sU)2n
(50)
Shelter Class
Description
1 2 3
No obstructions or local shielding Typical shelter for an isolated rural house Typical shelter caused by other buildings across street from building under study Typical shelter for urban buildings on larger lots where sheltering obstacles are more than one building height away Typical shelter produced by buildings or other structures immediately adjacent (closer than one house height: e.g., neighboring houses on same side of street, trees, bushes)
5
I = 141/255 = 0.55 h–1 = 0.55 ach
One
Table 5 Local Shelter Classes
4
The air change rate is therefore
where Qs = stack airflow rate, m3/s
Table 6 Basic Model Wind Coefficient Cw House Height (Stories)
Shelter Class
One
Two
Three
1 2 3 4 5
0.000 319 0.000 246 0.000 174 0.000 104 0.000 032
0.000 420 0.000 325 0.000 231 0.000 137 0.000 042
0.000 494 0.000 382 0.000 271 0.000 161 0.000 049
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Ventilation and Infiltration Qw c Cs Cw s T n
= = = = = = =
16.25
wind airflow rate, m3/s flow coefficient, m3/(s·Pan) stack coefficient, (Pa/K)n wind coefficient, (Pa·s2/m)n shelter factor indoor – outdoor temperature difference, K pressure exponent
Qs = (0.051)(0.089)[20 – (–19)]0.67 = 0.053 m3/s The wind flow is calculated using Equation (50): Qw = (0.051)(0.156)(0.64 3.95)1.34 = 0.027 m3/s Substituting Qs and Qw into Equation (47) gives Q = 0.059 m3/s = 214 m3/h. From Equation (2), air change rate I is equal to Q divided by the building volume:
In calculating tabulated values of Cs , Cw , and s, the assumptions were • Each story is 2.5 m high. • The flue is 15 cm in diameter and reaches 2 m above the upper ceiling. • The flue is unsheltered. • Half of envelope leakage (not including the flue) is in the walls and one-quarter each is at the floor and ceiling, respectively. • n = 0.67
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The following examples use typical values for terrain factors, house height, and wind speed measurement height, wind speed multiplier G given in Table 7, and use a relationship based on equations found in Chapter 24. Example 5. Estimate the infiltration at winter design conditions for a two-story slab-on-grade house with a flue in Lincoln, Nebraska. The house has a flow coefficient of c = 0.051 m3/(s · Pan) and a pressure exponent of n = 0.67 (corresponding to effective leakage area of 500 cm2 at 4 Pa). The building volume is 340 m3. The 97.5% design temperature is –19°C, and design wind speed is 6.7 m/s. Solution: For a slab-on-grade two-story house with a flue, Table 8 gives Cs = 0.089 (Pa/K)n and Cw = 0.156 (Pa·s2/m2)n. The house is maintained at 20°C indoors. The building wind speed is determined by taking the design wind speed Umet and applying by the wind speed multiplier G from Table 7: U = GUmet = 0.59(6.7) = 3.95 m/s From Table 5, the shelter class for a typical urban house is 4. Table 9 gives the shelter factor for a two-story house with a flue and shelter class 4 as s = 0.64. The stack flow is calculated using Equation (49):
Table 7 Enhanced Model Wind Speed Multiplier G House Height (Stories) Wind speed multiplier G
One
Two
Three
0.48
0.59
0.67
Table 8 Enhanced Model Stack and Wind Coefficients One Story
Cs Cw for slab-ongrade Cw for crawlspace
Two Story
No Flue
With Flue
0.054 0.156
0.069 0.142
0.128
0.128
Three Story
No Flue
With Flue
No Flue
With Flue
0.078 0.170
0.089 0.156
0.098 0.170
0.107 0.167
0.142
0.142
0.151
0.154
I = (214 m3/h)/340 m3 = 0.63 h–1 = 0.63 ach Example 6. Estimate the average infiltration over a one-week cold-weather period for a single-story house with a crawlspace in Redmond, Washington. The house has a flow coefficient of c = 0.078 m3/(s·Pa n) and a pressure exponent of n = 0.6 (corresponding to effective leakage area of 690 cm2 at 4 Pa). The building volume is 255 m3. During this period, the average indoor-to-outdoor temperature difference is 16 K, and wind speed is 2.7 m/s. The house is electrically heated and has no flue. Solution: For a single-story house with no flue, Cs =0.054 (Pa/K) n. For a crawlspace, Cw = 0.128 (Pa·s2/m2)n. From Table 7, for a one-story house, G = 0.48. U = GUmet = 0.48(2.7) = 1.3 m/s Table 9 gives shelter factor s = 0.50 for a house with no flue and shelter class 4. Stack flow is calculated using Equation (49): Qs = (0.078)(0.054)(16)0.6 = 0.022 m3/s Wind flow is calculated using Equation (50): Qw = (0.078)(0.128)(0.50 1.3)1.34 = 0.006 m3/s Substituting Qs and Qw into Equation (47) gives Q = 0.023 m3/s = 83 m3/h. From Equation (2), air change rate I is equal to Q divided by the building volume: I = (83 m3/h)/(255 m3) = 0.32 h–1 = 0.32 ach Example 7. Estimate the infiltration for a three-story house in San Francisco, California. The house has a flow coefficient of c = 0.102 m3/(s · Pan) and a pressure exponent of n = 0.67 (corresponding to effective leakage area of 1000 cm2 at 4 Pa). The building volume is 395 m3. The indoor/ outdoor temperature difference is 5 K and wind speed is 4.47 m/s. The house has a flue and a crawlspace. Solution: For a three-story house with a flue, Cs = 0.107(Pa/K)n. For a crawlspace, Cw = 0.154 (Pa· s2/m2)n. From Table 7, for a three-story house, G = 0.67. U = GUmet = 0.67(4.47) = 3.0 m/s The prevailing wind blows along the row of houses parallel to the street, so the house has a shelter class of 5. Table 9 gives the shelter factor for a three-story house with a flue and shelter class 5 as s = 0.43. Qs = (0.102)(0.107)(5)0.67 = 0.032 m3/s Qw = (0.102)(0.154)(0.43 3.0)1.34 = 0.022 m3/s Substituting Qs and Qw in Equation (47) gives Q = 0.039 m3/s = 140 m3/h. The air changes per hour are I = (140 m3/h)/(395 m3) = 0.35 h–1 = 0.35 ach
Combining Residential Infiltration and Mechanical Ventilation
Table 9 Enhanced Model Shelter Factor s Shelter Class
No Flue
One Story with Flue
Two Story with Flue
Three Story with Flue
1 2 3 4 5
1.00 0.90 0.70 0.50 0.30
1.10 1.02 0.86 0.70 0.54
1.07 0.98 0.81 0.64 0.47
1.06 0.97 0.79 0.61 0.43
Significant infiltration and mechanical ventilation often occur simultaneously in residences. The pressure difference from Equation (31) can be used for each building leak, and the flow network (including mechanical ventilation) for the building can be solved to find the airflow through all the leaks while accounting for the effect of the mechanical ventilation. However, for simplified models, natural infiltration and mechanical ventilation are usually determined
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2017 ASHRAE Handbook—Fundamentals (SI)
separately and require a superposition method to combine the flow rates. Sherman (1992b) compared various procedures and derived a generalized superposition equation that involves simple leakage distribution parameters. The result is always subadditive. For small, unbalanced fans, typically only half the mechanically introduced outdoor airflow contributes to the total, but this fraction can be anywhere between 0 and 100%, depending on leakage distribution. When the makeup air fan’s flow rate is large, infiltration may be ignored, because the building becomes either neutrally or positively pressurized. In special cases when the leakage distribution is known and highly skewed, it may be necessary to work through the superposition method in more detail. For example, in a wind-dominated situation, a makeup air fan has a much bigger effect than an exhaust fan on changing the total ventilation rate; the same is true for houses with high neutral pressure levels in cold climates. For the general case, when details are not known or can be assumed to be broad and typical, the following superposition gives good results: 2 2 Qcomb = Qbal + Q unbal + Q infiltration
(51)
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Typical Practice The preceding sections on estimating infiltration in low-rise residences represent current analytical techniques typically used for research and remediation purposes, but most small residential buildings are designed and constructed without direct involvement of ventilation engineers. Contractors, who typically prepare these buildings’ designs, are required to follow mandates in various codes and standards, and they apply experience-based rules of thumb when determining, for example, exhaust needs. Often, leaky buildings or air quality problems result. Research and experience have shown that tightening building envelopes and using mechanical ventilation with heat recovery can improve indoor air quality and reduce energy consumption. Retaining the services of a ventilation engineer before construction begins is advisable in some situations.
11.
COMMERCIAL AND INSTITUTIONAL AIR LEAKAGE Envelope Leakage ASTM Standard E779 and CGSB Standard 149.10 include methods to measure the airtightness of building envelopes of singlezone buildings. Although many multizone buildings can be treated as single-zone buildings for testing by opening interior doors or by inducing equal pressures in adjacent zones, these standards provide no guidelines for dealing with problems arising in tall buildings, such as stack and wind effects. Tall buildings require refinement and extensions of established procedures because they have obstacles to accurate measurement not present in small buildings, including large envelope leakage area, interfloor leakage, vertical shafts, and large wind and stack pressures. Chapter 4 in the 2015 ASHRAE Handbook—HVAC Applications discusses tall buildings and their challenges in detail. For conducting a pressurization test in a large building, the building’s own air-handling equipment sometimes can be used to induce test pressures, as described in CGSB Standard 149.15. In other cases, a large fan is brought to the building to perform the test, as described by CIBSE Standard TM23. ASHRAE RP-1478’s final report (Anis and Brennan 2014) discusses how to test the envelope leakage of large commercial buildings. In the past, building envelopes of large commercial buildings were often assumed to be quite airtight. Tamura and Shaw (1976a) found that, assuming a flow exponent n of 0.65 in Equation (40), air leakage measurements in eight Canadian office buildings with sealed windows ranged from 610 to 2440 cm3/(s·m2). Persily and Grot (1986) performed whole-building pressurization tests in large
office buildings that showed that pressurization airflow rate divided by building volume was relatively low compared to that of houses. However, if these airflow rates are normalized by building envelope area instead of by volume, the results indicate envelope airtightness levels similar to those in typical North American houses. The same study also looked at eight U.S. office buildings and found air leakage ranging from 1080 to 5220 cm3/(s·m2) at 75 Pa. This means that the office buildings’ envelopes were leakier than expected. Typical air leakage values per unit wall area at 75 Pa were 500, 1500, and 3000 cm3/(s·m2) for tight, average, and leaky walls, respectively (Tamura and Shaw 1976a). Emmerich and Persily (2014) summarized available measured airtightness data for almost 400 buildings, including about 70 constructed between 2000 and 2010. The average air leakage for the buildings was 20% tighter than the average for the 228 buildings included in a similar 2011 analysis. The data were analyzed to determine factors that affect airtightness (e.g., building type, height). Recent additions to the database, previously reported by Emmerich and Persily (2005), include numerous buildings constructed to meet the specifications of sustainable building programs such as U.S. Green Building Council’s LEED® rating system, as well as buildings designed and constructed with air retarders. The overall average airtightness reported was 13.1 m3/(h·m2) for six-sided (i.e., including slabs and basement surfaces) envelope surface areas at 75 Pa. The data show only weak trends related to year of construction, height, floor area, wall construction, or building type, but do demonstrate that buildings designed and constructed with attention to airtightness are much tighter than typical commercial buildings. The analysis found that the 79 buildings with air retarders had an average air leakage almost 70% less than the average for the 290 buildings not specified as having air retarders, thus demonstrating the critical need to design and construct commercial buildings with dedicated air retarders to support sustainable building design. However, the wide variation among the measures taken to limit or reduce air leakage among these buildings and the lack of detailed descriptions of the air retarders make it difficult to predict a specific level of airtightness that will result from use of a specific air retarder approach. In the United States, commercial building construction practices are addressed by various standards, codes, and green building program requirements, and Table 10 summarizes some of the relevant air leakage limits from these requirements. Both ASHRAE Standards 90.1 and 189.1 require continuous air barriers (CABs) for most commercial buildings. Since 2010, Standard 90.1 requires the CAB to meet either a material tightness limit [0.02 L/(s·m2) under a pressure differential of 75 Pa] or an assembly tightness limit [0.2 L/(s·m2) under a pressure differential of 75 Pa], but it does not include a whole-building tightness limit or a requirement for wholebuilding pressurization testing. The building commissioning requirements of Standard 189.1-2014 include a whole-building test demonstrating the building meets a tightness limit of 1.25 L/(s·m2) under a pressure differential of 75 Pa or the implementation of a rigorous envelope commissioning program. The 2012 International Energy Conservation Code® (IECC®) (ICC 2012a) has requirements with options for a CAB with material or assembly tightness or a whole-building test. The 2012 International Green Construction Code® (IgCC®) (ICC 2012b) includes the same requirements as the 2012 IECC but also includes a wholebuilding testing requirement consistent with the U.S. Army Corps of Engineers (USACE 2009) value. Many U.S. state building codes currently or will include requirements for continuous air barriers either through reference to IECC, IGCC, ASHRAE Standard 90.1 or 189.1, or their own independent requirements; a current list of requirements is available at www.airbarrier.org. Since 2009, the USACE has required that conditioned buildings be built or retrofitted to include a continuous air barrier to control air leakage through the building envelope (USACE 2009). The
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Table 10 Summary of Building Airtightness Data Air Leakage, Standard or Code ASHRAE 90.1 ASHRAE 189.1 IECC IgCC USACE GSA
L/(s·m2)
at 75 Pa
Material
Assembly
Whole Building*
0.02 0.02 0.02 Same as IECC 0.02 0.02
0.2 0.2 0.2 Same as IECC — 0.2
— 1.25 2.0 1.25 1.25 2.0
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Notes: IECC= International Energy Conservation Code IgCC = International Green Construction Code USACE = U.S. Army Corps of Engineers GSA = U.S. General Services Administration *Whole-building limits are based on six-sided enclosure, including slab and belowgrade walls.
specification requires whole-building testing with a maximum leakage of 1.25 L/(s·m2) at 75 Pa based on the six-sided building enclosure area (including slab and subgrade walls). The average tightness for a set of 285 new and retrofitted USACE buildings was reported to be 0.9 L/(s·m2) (Zhivov 2013). Also, the U.S. General Services Administration (GSA 2010) since 2010 requires all new U.S. federal buildings for the Public Buildings Service to include an air barrier with the whole building having an air leakage rate of not more than 2.0 L/(s·m2) at 75 Pa. Grot and Persily (1986) also found that eight recently constructed office buildings had infiltration rates ranging from 0.1 to 0.6 ach when outdoor air intakes were deactivated. Infiltration rates exhibited varying degrees of weather dependence, generally much lower than that measured in houses. Infiltration in commercial buildings can have many negative consequences, including reduced thermal comfort, interference with proper operation of mechanical ventilation systems, degraded indoor air quality, moisture damage of building envelope components, and increased energy consumption. These results suggest strongly that commercial buildings’ envelopes require tighter construction, and that continuous air barrier systems should be used in the enclosures of all conditioned buildings. Since 1997, the Building Environment and Thermal Envelope Council of the National Institute of Building Sciences has sponsored several symposia on air barriers for buildings in North American climates. Others have also published articles on the importance of limiting air leakage in commercial buildings (Anis 2001; Ask 2003; Fennell and Haehnel 2005). Envelope leakage in commercial buildings also depends on HVAC system operation. Often, commercial buildings and their HVAC systems are in operation during normal daytime business hours but switch into “unoccupied” operation at nights and on weekends and holidays. If pressurized while their HVAC systems operate, infiltration is often very low or even eliminated in buildings with tight envelopes. However, in unoccupied mode, this pressurization is often lost, so infiltration and potentially moisture intrusion may be significant at times.
Air Leakage Through Internal Partitions In large buildings, air leakage associated with internal partitions becomes very important. Walls, doors, floors, ceilings, plenums, chases, transfer ducts and grilles, and elevator and stairway enclosures are the major separations of concern in these buildings. Their leakage characteristics are needed in multizone models to predict infiltration through exterior walls and airflow patterns in a building. These internal resistances are also important in the event of a fire to predict smoke movement patterns and evaluate smoke management systems. Table 11 gives air leakage areas calculated at 75 Pa with CD = 0.65 for different internal partitions of commercial buildings (Klote et al. 2012). Figure 13 presents examples of measured air leakage rates of
Table 11 Air Leakage Areas for Internal Partitions in Commercial Buildings (at 75 Pa and CD = 0.65) Construction Element Stairwell walls Elevator shaft walls
Floors AL = air leakage area
Wall Tightness
Area Ratio
Tight Average Loose Tight Average Loose
AL/Aw 0.14 104 0.11 103 0.35 103 0.18 103 0.84 103 0.18 102 AL/Af
Average
0.52 104
Aw = wall area
Af = floor area
Fig. 13 Air Leakage Rates of Elevator Shaft Walls elevator shaft walls (Tamura and Shaw 1976b), the type of data used to derive the values in Table 11. Consult Chapter 53 of the 2015 ASHRAE Handbook—HVAC Applications for performance models and applications of smoke control systems. As with the exterior shell, making significant effort to seal unintentional openings in internal partitions in large buildings can improve thermal comfort and energy efficiency, as well as improve the building’s performance in a fire. Leakage paths, intentional or otherwise, at the top of elevator shafts are often equivalent to orifice areas of 0.4 to 1.0 m2. Air leakage rates through stair tower and elevator doors are shown in Figure 14 as a function of average crack width around the door. Sealing elevator and stair doors well, and possibly using indoor vestibules for them, can reduce air as well as smoke flow rates significantly. Air leakage areas associated with other openings in commercial buildings are also important for air movement calculations. These include interior doors and partitions, suspended ceilings in buildings where space above the ceiling is used for air supply or return, and other components of the air distribution system.
Air Leakage Through Exterior Doors Door infiltration depends on the type and use of door, room, and building, and on air speed and pressure differentials. In residences
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Fig. 15 Airflow Coefficient for Automatic Doors
Fig. 14 Air Leakage Rate of Door Versus Average Crack Width and small buildings where doors are used infrequently, a closed door is assumed, and air exchange can be estimated based on air leakage through cracks between door, frame, and sill. Where exterior door are used frequently, airflow through them increases significantly as door-opening frequency increases. Consider using vestibules or revolving doors for high-frequency applications.
Air Leakage Through Automatic Doors Automatically swinging, sliding, rotating, or overhead doors are a major source of air leakage. They are normally installed where large numbers of people use the doors or where bulk goods or vehicles are transported through the doorways. These doors stay open longer with each use than manually operated doors. Air leakage through an automatic door can be reduced by installing a vestibule. However, pairs of automatic doors on the inside and outside of a vestibule normally have overlapping open periods, even when used by only one person at a time. Therefore, it is important that designers include airflow through automatic doors when calculating heating and cooling loads in adjacent spaces. To calculate the average airflow rate through an automatic door, the designer must consider the area of the door, the pressure difference across it, the discharge coefficient of the door when it is open, and the fraction of time that it is open. Obtaining the discharge coefficient is complicated by the fact that it changes as the door opens and closes. To simplify this calculation, ASHRAE research project RP-763 (Yuill 1996) developed Figure 15 to combine the discharge coefficients of doors as they open and close with the fraction of time that doors are open at a particular level of use. This figure presents an overall airflow coefficient as a function of the number of people using a door per hour. To obtain the average infiltration rate through an automatic door, multiply this coefficient by the door’s opening area and by the square root of the pressure difference between the outdoor and indoor air at the door’s location. The pressure difference across a door in a building is difficult to predict accurately and depends on wind pressure on the building, stack effect caused by the indoor/outdoor temperature difference, and effects of air-handling system operation. It also depends on leakage characteristics of the building’s exterior walls and of internal partitions.
Fig. 16 Pressure Factor for Automatic Doors Two simplified design methods are presented here. The first method uses practical assumptions to determine design values for Rp, the square root of the pressure difference across the automatic door, given in Figure 16. The second method requires explicit calculation of envelope pressures. In Figure 16, airflows shown for outdoor air temperatures of 27 and 38°C, represented by dotted lines, are outward flows. They intercept the vertical axis at a lower point than the other lines because wind pressure coefficients on the building’s downwind face, where the greatest outward flows occur, are lower than on the upward face. In many buildings, air pressure in the building is controlled by varying the flow rate through return or exhaust fan(s) or by controlling the relief air dampers. These systems are usually set to maintain a pressure slightly above ambient in the lobby, but in a large building, multiple sensors may be used to regulate air pressure on each floor independently, for example. Subtracting the interior pressure maintained from the wind pressure gives the net pressure for estimating airflow through an exterior door. Method 1. For the first method, the infiltration rate through the automatic door is Q = CA ARp where Q = airflow rate, L/s
(52)
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CA = airflow coefficient from Figure 15, L/(s·m2 ·Pa0.5) A = area of the door opening, m2 Rp = pressure factor from Figure 16, Pa0.5
Air Exchange Through Air Curtains
Method 2. Airflow Q is Q = CA A p
(53)
where Q CA A p
= = = =
airflow rate, L/s airflow coefficient from Figure 15, L/(s·m2 ·Pa0.5) area of the door opening, m2 pressure difference across door, Pa
12.
To find p, it is necessary to find the pressure differential created by both wind and stack effect. To give the largest possible pressure difference across the door, there are no interactions between the two natural pressures: p = pw – ps
(54)
where pw = wind-induced surface pressure relative to static pressure, Pa ps = pressure difference caused by stack effect, Pa
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Air curtains are jets of air projected across envelope openings with the intention of reducing air exchange and the entrance of dust and insects, for example. They are commonly applied to loading dock doorways and high-use building entrances. Performance of air curtains strongly depends on factors such as jet characteristics, wind, and building pressurization. More discussion on air curtain performance is available in the research literature and in Chapter 20 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment.
Example Calculations Find the maximum possible winter infiltration through an automatic door on the ground floor of a 20-story building. The area of the door is 0.91 2.1 m = 1.9 m2. Each floor is 4 m high. Approximately 300 people per hour pass through the door. The design wind conditions are 6.7 m/s, indoor temperature is 21°C, and outdoor temperature is –7°C. The airflow coefficient from Figure 15, using the line for doors without vestibules, is approximately 306 L/(s·m2 ·Pa0.5). Method 1: The pressure factor from Figure 16 is 9.6 Pa0.5. Equation (52) gives the door’s airflow as Q = 306(1.9)9.6 = 5600 L/s Method 2: The worst possible case for wind surface pressure coefficient Cp at any point and in any position on the ground floor of the building is inferred from the figures in Chapter 24 to be about 0.75. Using this in Equation (25), together with the specified wind speed, results in pw = 20 Pa. Assume that H is one-half the door height (1.1 m). To have maximum pressure across the door, assume the neutral pressure plane is located halfway up the building such that 1 4m HNPL = --- 20 stories ------------ = 40 m 2 story Substituting these values into Equation (24) gives ps = –47 Pa. This is the maximum stack pressure difference given no internal resistance to airflow. To find the actual stack pressure difference, it is necessary to multiply this by a draft coefficient. For this example, the coefficient is assumed to be 0.9, which is the highest value that has been found for tall buildings. Therefore, ps = 0.9(–47 Pa) = – 42 Pa. The total pressure is then p = 20 –(–42) = 62 Pa. Substituting into Equation (53), Q = 306(1.9) 62 = 4580 L/s If the building has a vestibule, the airflow coefficient is read from Figure 15 using the line for doors with vestibules, and it is approximately 208 L/(s·m2 ·Pa0.5), reducing airflow to 3100 L/s into the building.
Standard-sized revolving doors are often used where there is high use, but typically not where people would often be carrying large bags or pushing carts. Dols et al. (2014) applied a coupled building energy and airflow simulation tool to simulate airflow through entry doors in a prototype medium office building created by Ng et al. (2012) and found that using a vestibule reduced infiltration through the building entrance by 23%. More information on predicting infiltration rates through standard and extra-large revolving doors is still needed.
COMMERCIAL AND INSTITUTIONAL VENTILATION
ASHRAE Standard 62.1 contains requirements on ventilation and indoor air quality for commercial, institutional, and high-rise residential buildings. These requirements address system and equipment issues, design ventilation rates, commissioning and systems start-up, and operation and maintenance. The user’s manual for Standard 62.1-2010 (ASHRAE 2010b) provides details to help design, install, and operate building systems to meet requirements. The design requirements include two alternative procedures: • The prescriptive ventilation rate procedure (VRP) contains a table of outdoor air ventilation requirements for a variety of space types, with adjustments for air distribution in rooms and systems serving multiple spaces. These requirements consist of both a per-person rate and a per-floor-area rate. Minimum outdoor air ventilation rates are based, in part, on research by BergMunch et al. (1986), Cain et al. (1983), Iwashita et al. (1989), and Yaglou et al. (1936), as well as years of experience of designers and building operators. • The performance indoor air quality procedure (IAQP) seeks acceptable indoor air quality by controlling indoor contaminant concentrations through source control, air cleaning, and ventilation. It allows for either or both improved indoor air quality and reduced energy consumption. Chapter 29 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment has information on air cleaning. The ventilation rate procedure is by far the more commonly used because of its prescriptive nature. Combining source control and local exhaust, as opposed to dilution with ventilation air, is the method of choice in many industrial environments. Industrial ventilation is discussed in Chapters 31 and 32 of the 2015 ASHRAE Handbook—HVAC Applications and in Industrial Ventilation: A Manual of Recommended Practice (ACGIH 2016). Ventilation of medical facilities, where high indoor air quality is expected, is discussed in ASHRAE Standard 62.1, Chapter 8 of the 2015 ASHRAE Handbook—HVAC Applications, and other publications [e.g., FGI (2014)]. Commercial and institutional building ventilation systems are typically designed to provide slight pressurization to reduce infiltration. This pressurization is achieved by having the outdoor or makeup airflow rate higher than the exhaust or relief airflow rate. In these buildings, infiltration is therefore usually neglected in HVAC design except in areas such as lobbies and loading docks, where infiltration can be important because of doors. However, as discussed previously, this little to no infiltration may only be achieved in practice with very tight envelope construction (e.g., including a continuous air barrier). As discussed in the section on Driving Mechanisms for Ventilation and Infiltration, wind and stack effect can also cause significant infiltration and exfiltration. Ventilation airflow rates for commercial and institutional buildings are typically determined using procedures in ASHRAE Standard 62.1. In these procedures for designing mechanical ventilation systems, no credit is given for infiltration. However, weather-driven pressure differentials may be significant and need to be considered when designing the ventilation system.
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Ventilation Rate Procedure Per ASHRAE Standard 62.1, the design ventilation rate is determined starting with a table of minimum ventilation requirements for different space types. These requirements are expressed as an outdoor airflow rate per occupant or per unit floor area, or often both, depending on space type. These ventilation rates are based on air pollutants generated by people, activities, and building materials and furnishings. The rates are then adjusted for various parameters (e.g., multiple zones, type of room air distribution). The HVAC designer faces several challenges in designing an air distribution system to deliver outdoor air to building occupants. The first is to determine whether the outdoor air is acceptable for use, and to design a system for cleaning the air if it is not acceptable. A second is to design an air intake and distribution system that will deliver the required level of outdoor air to the occupied portions of the building, and not just admit it to an air handler. This outdoor air must be delivered not only at design conditions, but throughout the year. The task is complicated by weather-related variations in indoor/outdoor pressure difference. Other complications include pressure variations caused by building components such as intermittent exhaust fans or dirty filters, and probably most significantly by supply flow variations associated with variable-air-volume (VAV) systems (Janu et al. 1995; Mumma and Wong 1990). This delivery issue is related to the discussion in the section on Air Change Effectiveness.
Multiple Spaces Many commercial and institutional buildings have multiplezone, recirculating ventilation systems wherein one or more conventional air handlers condition a mixture of outdoor and recirculated air (supply air) to more than one ventilation zone. Each zone may have a different outdoor air fraction required by ASHRAE Standard 62.1, but each air handler can typically only provide one outdoor air fraction. Therefore, the zone that requires the greatest outdoor air fraction (the critical zone) defines the outdoor air intake rate of the air handler. Consequently, all other zones receive more outdoor air than their minimum requirement. These zones can be considered to have unused ventilation air, which could be returned to the air handler and recirculated; thus, the outdoor air fraction at the air handler could be reduced while still meeting the needs of the critical zone. A method to address multiple-zone recirculating systems, based on the system ventilation efficiency equation (SVEE), is provided in the ventilation rate procedure of Standard 62.1. ASHRAEsponsored research experimentally tested the validity of this equation and confirmed that the SVEE is a valid predictor of ventilation distribution in a building (Yuill et al. 2012). Secondary Path Systems. Some systems circulate unused ventilation air through paths other than through the central air handler. Common examples include transfer fans and fan-powered boxes. Warden (1995) suggested that Standard 62.1 should allow for the increased distribution efficiency that is possible with these systems, and presented a generalized SVEE that includes secondary air paths. A form of this equation is given in Equation (A-3) of ASHRAE Standard 62.1-2010. Equation (A-3) depends on the ventilating ability of the secondary air Er . The standard previously described Er in terms of the proportion of average system return air directly recirculated from the critical zone. Yuill et al. (2008) argued that the formulation of Equation (A-3) implicitly defines Er as a descriptor of the vitiation of the secondary air, and presented a formal definition of Er , a version of which was adopted for Standard 62.1. Under this definition, if secondary air is drawn from an area that is less contaminated than the building average, values of Er greater than 1 are possible. Yuill et al. (2012) measured Er in several locations in an office building and found results ranging from 0.74 to 1.01. Results in the same building
2017 ASHRAE Handbook—Fundamentals (SI) could have ranged from 0.14 to 1.13 if the fan-powered boxes had been located differently.
Survey of Ventilation Rates in Office Buildings Relatively few measurements of as-built office building ventilation performance have been conducted, and those data generally have not used consistent measurement methods or involved representative collections of buildings. The U.S. Environmental Protection Agency (EPA) Building Assessment Survey and Evaluation (BASE) study involved indoor environmental measurements, including ventilation, in 100 randomly selected office buildings using a standardized protocol (EPA 2003). Persily et al. (2005) analyzed the BASE data and found that outdoor air ventilation rates measured using duct traverses at air handler intakes were higher than might be expected, with a mean of about 55 L/s per person. However, these elevated values are partially explained by low occupant density (mean of about four persons per 100 m2) and high outdoor air fractions (mean of about 35%). Considering only values that correspond to minimum outdoor air intake, the mean ventilation rate was 11 L/s per workstation. About one-half the ventilation rates under minimum outdoor air intake were below 9.4 L/s per person. Another key outcome of this study is documentation of measured airflow rates that are quite different from their design values. This finding highlights the need for good system commissioning and maintenance to achieve design intent. Designing and configuring systems to encourage regular maintenance by providing easy access to key system components is also important.
13.
OFFICE BUILDING EXAMPLE
Ventilation and infiltration principles from this chapter, Standard 62.1-2010, and elsewhere are applied to a conventional office building in Atlanta, Georgia. The infiltration, local exhaust, or ventilation airflow rates determined can be used later in the design process (1) as input for the heating and cooling load calculations; (2) for sizing fans, ducts, and dampers; and (3) for inclusion in the construction documents’ air-handling units (AHUs) schedules and specifications. This example relies on the 2010 edition of ASHRAE Standard 62.1; because this and other standards are updated frequently, users should check for the latest edition.
Location The example building is about 13 km northeast of downtown Atlanta, and is close to a major highway and its access roads. Atlanta’s climate is hot and humid in the summer, and has relatively mild winters. The average annual outdoor air temperature is about 15.9°C and the heating kelvin-days per year, base 18.3°C (HKD18.3), are about 1814 (Rock 2005). From Chapter 14, the winter 99% design outdoor air (OA) temperature is –5°C, whereas the 1% cooling dry-bulb temperature is 32.8°C, with a mean coincident wet-bulb temperature of 23.3°C. The 99.6 and 0.4% design wind speeds are 5.4 m/s in the winter and 4 m/s in the summer, both from the northwest. Warm and humid winds also travel north from the Gulf of Mexico, and occasionally the wind is from the Atlantic Ocean from the southeast.
Building The approximately 2835 m2 building is a two-story, flat-roofed, slab-on-grade commercial office building with a substantial roof overhang in each direction. Materials and construction quality are average commercial grade. Double-paned windows, and similar spandrel glass, are fixed in their metal curtain wall frames; all windows are nonoperable. The remaining portions of the exterior walls are brick. There are relatively few exterior doors, as described later in this example. The building is surrounded by black asphalt driveways, a parking lot, and some vegetation. The nearby highway is across a parallel two-lane access road, to the northwest.
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Ventilation and Infiltration Occupancy The building is occupied during normal weekday business hours, and occasionally for special weekend events. Night and weekend thermostat setbacks are used. On the perimeter of the building are mostly single-person offices and conference rooms. The core of the building is mainly open plan with cubicle workspaces, as well as various support rooms, restrooms, two stair towers, and an elevator. There is a large mailroom on the first floor and a lunchroom on the second. Occupant density is high during workdays. The overhead fluorescent lighting is typical of such office buildings, and there are significant computing, printing, and copying equipment loads. Smoking is not allowed in the building. The building is owner occupied. Owners generally have longterm interests in minimizing costs, and in maximizing indoor air quality and thermal comfort so that workers’ productivity is high.
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Infiltration For this example, assume a conventional all-air overhead HVAC system, and that the building is well sealed. Consequently, a slightly positive overall building pressurization is assumed during occupied hours. Because water condensation in the exterior envelope of the building is possible, air pressurization should be as low as is practical, and continuous air and moisture retarders should be installed. As a more expensive alternative to slight pressurization, the automatic control system could actively manage the dampers’ positions and fans’ operation to maintain an average neutral pressurization, relative to the outdoors. In either case, a good assumption is that infiltration is minimized, the windows and spandrel glass are fixed and well sealed, and the exterior doors are normally kept closed. During high-wind conditions beyond design, windward perimeter spaces may have some infiltration loads, but under nonpeak outdoor temperatures, a well-zoned HVAC system should have enough capacity to handle these extra loads. If both the OA temperature and wind are extreme, then these upwind perimeter spaces may become slightly uncomfortable. These extreme conditions are expected to occur only a few hours in a typical year. Spaces with exterior doors can experience significant infiltration loads when people enter and leave. First-floor vestibules on the north and south sides of the building help limit this infiltration through the two main entrances. The double doors from stair tower #2 have infrequent use, and a high level of thermal comfort in stair towers is not typically expected. Thus, brief infiltration surges in stair tower #2 are deemed acceptable. However, the doors from the parking lot to the mailroom are frequently used by staff for shipping, receiving, entrance, and egress, and infiltration loads on the vestibules and mailroom are of concern. Many designers choose to ignore these extra loads in pressurized buildings, because they are transient and not easily characterized; the systems’ capacities are likely sufficient to minimize uncomfortable conditions in these spaces. In this example, however, the HVAC designer is concerned about summertime airborne moisture, especially in the mailroom where books and other publications are stored, because strong, humid, southerly winds easily overcome a slight indoor pressurization when the large doors to the parking lot are open. This chapter and many of its supporting references describe detailed methods for estimating infiltration or air leakage. Typically, pressure differences, openings’ coefficients, and hour-byhour weather data are required to perform these transient calculations, usually using a computer program separate from that used for thermal load calculations. For HVAC design purposes for a building similar to the example, an air change rate of unconditioned outdoor air through infiltration, per space, expressed in air changes per hour or airflow rate (in L/s) is of more immediate use. Either value is then entered into the load calculation program. Unfortunately, accurate air changes per hour are difficult, if not
16.31 impossible, to predict, so design estimates must be made. For example, North Vestibule • • • •
Gross floor area 3.4 m × 4 m = 13.6 m2 Room volume 13.6 m2 × 2.7 m = 36.7 m3 ACHinf 1.0, so Qinf (36.7 m3 × 1000 L/m3 × 1.0)/3660 s/h 10.2 L/s outdoor air
Either 1.0 ach or 10.2 L/s of infiltration is then used as input for the load calculation program for this space. The 1.0 ach assumption was made by the designer during on-site observation that these particular manually operated exterior doors have low usage. If passage rates were known, Yuill’s (1996) flow rate estimation method would have been used instead. South Vestibule • • • •
Gross floor area 2.4 m × 3 m = 7.2 m2 Room volume 7.2 m2 × 2.7 m = 19.4 m3 ACHinf 2.0, so Qinf (19.4 m3 × 1000 L/m3 × 2.0)/3600 s/h 10.8 L/s outdoor air
In practice, this back entrance from the parking lot on the southeast side of the building is the primary means of entrance and egress, and as such, the estimated infiltration for it is increased to 2.0 ach, compared to the north vestibule’s 1.0 ach. In colder U.S. climates, it is common practice for low-cost commercial buildings to have only space heating, and not cooling, in stair towers and vestibules. However, for this building in the Southeast, the designer decided to provide cooling for these vestibules. Thus, the estimated infiltration rates are applied to both the heating and cooling load calculations for these spaces. The building’s mailroom, which also has exterior doors, is to be heated and cooled, too. Mailroom • • • •
Gross floor area (15.5 m × 6.7 m) + (10 m × 3 m) = 134 m2 Room volume 134 m2 × 2.7 m = 362 m3 ACHinf 0.5, so Qinf (362 m3 × 1000 L/m3 × 0.5)/3600 s/h 50.3 L/s outdoor air
Even though the mailroom has only a single layer of doors to the outdoors, and not a vestibule, the designer estimated the infiltration at a lower rate (0.5 ach) than those for the vestibules. This is because of the mailroom’s large interior volume relative to its exterior doorway’s area. Note that no estimate of air changes will be accurate at all times; this portion of HVAC design is still largely an art because of the many unknowns and variability of weather and building use. For improved energy conservation, all exterior doors must be extremely well weatherstripped and have automatic closers. A sign indicating that the doors should be kept closed when not in use should be placed on the mailroom’s exterior doors. High-quality gaskets and sealants for the windows and spandrel glass are also required to minimize infiltration.
Local Exhausts (This section assumes that ANSI/ASHRAE Standard 62.1-2010 has been adopted into the local building code without modification.) At least 10 rooms require direct, powered air exhaust: two restrooms per floor, a darkroom, three designated photocopy spaces, and two janitors’ closets. The restrooms have three flushable fixtures each, so from Table 6-4 of Standard 62.1-2010, with intermittent use, each restroom requires Qea = 3 units × 25 L/s per unit = 75 L/s Also from Table 6-4, the darkroom on the second floor needs Gross floor area 3 m × 4.6 m = 13.8 m2
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16.32
2017 ASHRAE Handbook—Fundamentals (SI) Qea = 13.8 m2 × 5.0 L/(s·m2) = 69 L/s
Similarly, the designated photocopy areas need 2.5 L/(s·m2), so First floor, plan east: 7.4 m2 × 2.5 L/(s·m2) = 18.5 L/s exhaust air First floor, plan southwest: 14.9 m2 × 2.5 L/(s·m2) = 37.3 L/s exhaust air Second floor, plan east: 10.4 m2 × 2.5 L/(s·m2) = 26 L/s exhaust air The two small janitors’ closets, one on each floor, also require exhaust:
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5.6 m2 × 5.0 L/(s·m2) = 28 L/s exhaust air These local exhaust airflow rates are then entered into the load calculation program. They are room loads, attached to each particular space, and are not combined and entered as systems-level loads. The load calculation program evaluates the room loads, appropriately combines them, and then finds the systems-level loads for various peak hours. Some local code authorities amend the requirements of Standard 62.1, or have not yet adopted the most current version, so significant deviations from these examples are possible. For example, in much of the United States, janitorial closets and photocopy rooms have not been required to have local exhausts. Standard 62.1-2010 recognized that these spaces can be significant sources of airborne pollutants, and some direct exhaust from each of them can be very beneficial for improving indoor air quality.
Ventilation (This section also assumes that ANSI/ASHRAE Standard 62.12010 has been adopted into local code without changes.) Ventilation air is needed to maintain acceptable indoor air quality. The example building is well sealed, natural ventilation is not used, and no credit for any infiltration is taken toward ventilation air requirements, as is typical for conventional commercial buildings. Thus, minimum ventilation air required by Standard 62.1 is provided mechanically through the building’s AHUs. Because smoking is not allowed in the building, no extra ventilation for environmental tobacco smoke (ETS) is needed. However, considering outdoor air pollution from the major highway nearby as well as metropolitan Atlanta’s smog, some outdoor air pretreatment may be considered later in the design process. Standard 62.1 has two methods for determining needed ventilation airflow rates: the performance IAQ procedure (IAQP), and the prescriptive ventilation rate procedure (VRP). Most HVAC designers of conventional buildings with normal occupancies and outdoor air conditions use the VRP, which is appropriate for this example building. Required ventilation air, which is solely conditioned outdoor air, is admitted to this building through two air-handling units; each AHU serves one floor. Flow rates of outdoor air are input values for, and carried through to the results of, the load calculation simulation. Energy needed to condition the outdoor air ultimately is a systemslevel load, because all of this ventilation air is conditioned by the AHUs before its introduction to the building. Commercial load calculation programs often provide suggested values of ventilation airflow rates and occupancy schedules, but may not have been updated to reflect the latest VRP requirements and procedures of Standard 62.1. As such, it is difficult to present an example here; instead, a sample check using some assumed values for the first-floor executive director’s office is given. It is assumed that this room is a separate thermal zone because of its use and its location on the southwest corner of the building and thus two afternoon solar exposures. Executive Director’s Office • Gross floor area 3.7 m × 6.4 m = 23.7 m2
• Room volume 23.7 m2 × 2.7 m = 64 m3 • Assumed supply air Qsa = 194 L/s The supply airflow rate was estimated using 7.93 m2/kW, a sensible heat factor of 0.9, a cooling supply (12.8°C) to room (23.9°C) air temperature difference of 11.1 K, 1000 W per kW of cooling, and the sensible heat equation 1.23 × L/s × T. From Table 6-1 of Standard 62.1-2010, the office’s population P can be estimated as P = 23.7 m2 × 5 occupants/100 m2 = 1.19 In this case, however, there is only one regular occupant of the space. The needed ventilation airflow rate to the breathing zone Vbz is then found from the table as follows: Vbz = RpPz + RaAz Vbz = [2.5 L/(s·person) × 1 person] + [0.3 L/(s·m2) × 23.7 m2] = 9.61 L/s where Rp = outdoor airflow rate required per person, from Standard 62.1’s Table 6-1, L/s Pz = zone population (largest number of people expected to occupy the zone during typical use) Ra = outdoor airflow rate required per unit area, from Standard 62.1’s Table 6-1, L/s Az = occupiable floor area of zone, m2
Note that Standard 62.1’s VRP includes a building component RaAz, as well as the traditional per-person people component. Because this is a conventional office building, with ceiling plenums and no raised floors, overhead air supply and return is assumed. The cooling mode, not heating, is dominant in this and most other U.S. office buildings that have high internal heat gains and well-sealed and insulated envelopes. From the standard’s Table 6-2, with ceiling supply of cool air, the zone air distribution effectiveness Ez is estimated as 1.0. From Equation (6-2) of the standard, the design zone outdoor airflow rate Voz is then Voz = Vbz/Ez = (9.61 L/s)/1.0 = 9.61 L/s But this is still not the amount of outdoor air that must be conditioned by the air handler: the rate must be adjusted for inefficiencies and recirculation in the air distribution system. Because single-duct VAV with terminal reheat air distribution systems were initially planned by the designer, Standard 62.1’s multiple-zone recirculating systems adjustment is needed. For this thermal zone, the primary outdoor air fraction Zp for its VAV terminal unit and downstream is Zp = Voz/Vpz = (9.61 L/s)/(194 L/s) = 0.05, or 5% However, for VAV systems, the minimum expected primary airflow rate should be used. In this case, 194 L/s is the peak design airflow rate. Designers often assume about 30% of this peak flow as the minimum in VAV systems, so for this space, 194 × 0.3 = 58.2 L/s. The adjusted primary outdoor air fraction is then Zp = Voz/Vpz = (9.61 L/s)/(58.2 L/s) = 0.17, or 17% The preceding calculations need to be performed for every thermal zone on each air handler. Then, for each system, the highest primary outdoor air fraction is used to estimate the air distribution systems’ ventilation effectiveness; the 62.1 user’s manual (ASHRAE 2010b) includes a spreadsheet for doing these calculations. Increasingly, load calculation programs include the necessary routines and data.
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Ventilation and Infiltration For the purposes of this example, 0.17 is assumed to be the maximum Zp, so, from Table 6.3 of Standard 62.1, the system ventilation efficiency Ev is 0.9. If, instead, the standard’s Appendix A method for determining Ev were used, a value closer to 1.0 for perfect mixing would likely result for this example’s conventional overhead allair cooling system. Table 6-3’s value of 0.9 is likely somewhat conservative, but is obtained quickly for design purposes. Next, the uncorrected outdoor air intake flow rate Vou is needed; Standard 62.1’s Equation (6-6) includes diversity factor D to adjust the people component of the flow rate. All zones’ flow rates are needed to perform this calculation. For this example, the uncorrected outdoor air intake flow rate for the first floor’s AHU was estimated from floor area, an occupancy of 5 people per 100 m2, and 9.4 L/s per person, and is assumed to be 720 L/s. The adjusted outdoor air intake flow rate Vot for this AHU is then
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Vot = Vou/Ev = (720 L/s)/0.9 = 800 L/s outdoor air After load calculations are complete, these assumed airflow rates can be replaced with actual values for each zone, and the outdoor airflow rate can be updated. Repeating the load calculations may be necessary. The final value of the adjusted outdoor air intake flow rate is then reported on the AHU’s equipment schedule so that testing, adjusting, and balancing (TAB) personnel and others can use this information to ensure that the system admits the desired flow rate of ventilation air. The information is also used to select air cleaners, dampers, coils, ducts, and fans. For more examples on determining ventilation air rates for commercial buildings, see the user’s manual for Standard 62.1 (ASHRAE 2010b). For low-rise residential buildings, consult ASHRAE Standard 62.2 and its user’s manual (ASHRAE 2010a).
14. A c cp C C
CA CD Cp Cs Cv Cw E ELA F F f g G h H i I Ii Im IDD L n N NL p P q Q Q R s
= = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = =
SYMBOLS
area, m2 or cm2 flow coefficient, m3/(s·Pan) specific heat, J/(kg·K) or kJ/(kg·K) concentration, ppm time averaged concentration airflow coefficient for automatic doors, L/(s·m2·Pa0.5) discharge coefficient pressure coefficient stack flow coefficient, (L/s)2/(cm4 ·K) or (Pa/K) n effectiveness of openings wind flow coefficient, (L/s)2[cm4 ·(m/s)2] or [Pa/(m/s)]n system efficiency effective leakage area tracer gas injection rate, m3/s time-averaged contaminant source strength, m3/s fractional on-time gravitational acceleration, m/s2 wind speed multiplier, Table 7 specific enthalpy, kJ/kg height, m hour of year air change rate, 1/time instantaneous air change rate, 1/time effective air change rate, 1/time infiltration degree-days, K·day/yr leakage area, m2 or cm2 pressure exponent number of discrete time periods in period of interest normalized leakage pressure, Pa parameter, or number of people heat rate, W volumetric flow rate, m3/s effective volumetric flow rate, m3/s outdoor airflow rate, L/s shelter factor
16.33 = = = = = = = age = = = = S t T U V W I
source strength, m3/s time temperature, °C or K wind speed, m/s volume, m3, or ventilation airflow rate, L/s humidity ratio, kg/kg air change effectiveness age of air, s, min, or h air density, kg/m3 time constant, s, min, or h wind angle, degrees
Subscripts a b ba bz c ca da depress e ea f i inf H ka l la L ma met n N NPL o oa ot ou oz p press r s sa S w v z
= = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = =
area base bypass air breathing zone calculated recirculated air dry air depressurization effective exhaust air floor indoor or time counter for summation (instantaneous) infiltration building height, eaves or roof makeup air latent relief air leakage or local mixed air meteorological station location normalized nominal neutral pressure level outdoor, initial condition, or reference outdoor air adjusted outdoor air uncorrected outdoor air zone outdoor pressure, or primary pressurization reference sensible or stack supply air space or source wind or water ventilation zone
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ACGIH. 2016. Industrial ventilation: A manual of recommended practice, 29th ed. American Conference of Governmental Industrial Hygienists, Cincinnati, OH. Ackerman, M.Y., J.D. Dale, and D.J. Wilson. 2006. Infiltration heat recovery, part 1: Field studies in an instrumented test building (RP-1169). ASHRAE Transactions 112(2):597-608. Paper QC-06-056. AIVC. 1994. An analysis and data summary of the AIVC’s numerical database. Technical Note 44. International Energy Agency Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. Akins, R.E., J.A. Peterka, and J.E. Cermak. 1979. Averaged pressure coefficients for rectangular buildings, vol. 1, Proceedings of the Fifth International Wind Engineering Conference, Fort Collins, pp. 369-380. Allard, F., and M. Herrlin. 1989. Wind-induced ventilation. ASHRAE Transactions 95(2):722-728. Paper VA-89-10-2.
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16.34 Anis, W. 2001. The impact of airtightness on system design. ASHRAE Journal 43(12):31-35. Anis, W., and T. Brennan. 2014. Measuring airtightness of mid and high-rise non-residential buildings. ASHRAE Research Project RP-1478, Final Report. Apte, M.G., W.J. Fisk, and J.M. Daisey. 2000. Associations between indoor CO2 concentrations and sick building syndrome symptoms in US office buildings: An analysis of the 1994-1996 BASE study data. Indoor Air 10(4):246-257. ASHRAE. 2003. Risk management guidance for health, safety and environmental security under extraordinary incidents. Report, Presidential Ad Hoc Committee for Building Health and Safety Under Extraordinary Incidents. ASHRAE. 2010a. Standard 62.2-2010 user’s manual. ASHRAE. 2010b. Standard 62.1-2010 user’s manual. ASHRAE. 2012. Method of testing general ventilation air-cleaning devices for removal efficiency by particle size. ANSI/ASHRAE Standard 52.22012. ASHRAE. 2013. Thermal environmental conditions for human occupancy. ASHRAE Standard 55-2013. ASHRAE. 2010. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2010. ASHRAE. 2016. Ventilation and acceptable indoor air quality in low-rise residential buildings. ANSI/ASHRAE Standard 62.2-2016. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE Standard 90.1-2016. ASHRAE. 2004. Air leakage performance for detached single-family residential building. ANSI/ASHRAE Standard 119-1988 (RA 2004) (withdrawn). ASHRAE. 2002. Measuring air-change effectiveness. ANSI/ASHRAE Standard 129-97 (RA 2002). ASHRAE. 2013. Ventilation of health care facilities. ANSI/ASHRAE/ ASHE Standard 170-2013. ASHRAE. 2014. Standard for the design of high-performance green buildings. ANSI/ASHRAE/IES/USGBC Standard 189.1-2014. ASHRAE. 2015. Ventilation and indoor air quality in low-rise residential buildings. ASHRAE Guideline 24-2015. Ask, A. 2003. Ventilation and air leakage. ASHRAE Journal 45(11):28-36. ASTM. 2012. Test method for determining rate of air leakage through exterior windows, curtain walls, and doors under specified pressure differences across the specimen. Standard E283-04 (R2012). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Test method for determining air change in a single zone by means of a tracer gas dilution. Standard E741-11. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Test method for determining air leakage rate by fan pressurization. Standard E779-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Test method for field measurement of air leakage through installed exterior windows and doors. Standard E783-02 (R2010). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2009. Practices for air leakage site detection in building envelopes and air barrier systems. Standard E1186-03 (R2009). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Test methods for determining airtightness of buildings using an orifice blower door. Standard E1827-11. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Guide for assessing depressurization-induced backdrafting and spillage from vented combustion appliances. Standard E1998-11. American Society for Testing and Materials, West Conshohocken, PA. Axley, J.W. 2001a. Application of natural ventilation for U.S. commercial buildings—Climate suitability, design strategies and methods, modeling studies. GCR-01-820, National Institute of Standards and Technology, Gaithersburg, MD. Axley, J.W. 2001b. Residential passive ventilation systems: Evaluation and design. Technical Note 54. International Energy Agency Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. Axley, J., S. Emmerich, S. Dols, and G. Walton. 2002. An approach to the design of natural and hybrid ventilation systems for cooling buildings. Proceedings of the 9th International Conference on Indoor Air Quality and Climate, Monterey, CA. Available at www.nist.gov/manuscript -publication-search.cfm?pub_id=860870.
2017 ASHRAE Handbook—Fundamentals (SI) Bankvall, C.G. 1987. Air movements and thermal performance of the building envelope. In Thermal insulation: Materials and systems, pp. 124131. F.J. Powell and S.L. Mathews, eds. American Society for Testing and Materials, West Conshohocken, PA. Barley, D. 2001. Overview of residential ventilation activities in the Building America Program (phase I). NREL/TP-550-30107, National Renewable Energy Laboratory, Golden, CO. Batterman, S., G. Hatzivasilis, and C. Jia. 2006. Concentrations and emissions of gasoline and other vapors from residential vehicle garages. Atmospheric Environment 40:1828-1844. Bauman, F., and A. Daly. 2003. Underfloor air distribution design guide. ASHRAE. Berg-Munch, B., G. Clausen, and P.O. Fanger. 1986. Ventilation requirements for the control of body odor in spaces occupied by women. Environmental International 12(1-4):195. Berlad, A.L., N. Tutu, Y. Yeh, R. Jaung, R. Krajewski, R. Hoppe and F. Salzano. 1978. Air intrusion effects on the performance of permeable insulation systems. In Thermal insulation performance, STP 718, pp. 181194. D. McElroy and R. Tye, eds. American Society for Testing and Materials, West Conshohocken, PA. Blomsterberg, A.K., and D.T. Harrje. 1979. Approaches to evaluation of air infiltration energy losses in buildings. ASHRAE Transactions 85(1):797. Paper PH-79-10-1. Bohac, D.L., D.T. Harrje, and L.K. Norford. 1985. Constant concentration infiltration measurement technique: An analysis of its accuracy and field measurements, 176. Proceedings of the ASHRAE/DOE/BTECC Conference on the Thermal Performance of the Exterior Envelopes of Buildings III, Clearwater Beach, FL. Bohac, D.L., D.T. Harrje, and G.S. Horner. 1987. Field study comparisons of constant concentration and PFT infiltration measurements. Proceedings of the 8th IEA Conference of the Air Infiltration and Ventilation Centre, Überlingen, Germany, pp. 47-62. Bradley, B. 1993. Implementation of the AIM-2 infiltration model in HOT2000. Report, for Natural Resources Canada. Buchanan, C.R., and M.H. Sherman. 2000. A mathematical model for infiltration heat recovery. Proceedings of the 21st IEA Conference of the Air Infiltration and Ventilation Centre, The Hague, Netherlands. Report LBNL-44294. Lawrence Berkeley National Laboratory, Berkeley, CA. Cain, W.S., B. Leaderer, R. Isseroff, L. Berglund, R. Huey, and E. Lipsitt. 1983. Ventilation requirements in buildings—I. Control of occupancy odor and tobacco smoke odor. Atmospheric Environment 17(6):11831197. CGSB. 2005. Depressurization test. CAN/CGSB Standard 51.71-2005. Canadian General Standards Board, Ottawa, ON. CGSB. 1986. Determination of the airtightness of building envelopes by the fan depressurization method. Standard 149.10-M86. Canadian General Standards Board, Ottawa, ON. CGSB. 1996. Determination of the overall envelope airtightness of buildings by the fan pressurization method using the building’s air handling systems. Standard 149.15-96. Canadian General Standards Board, Ottawa, ON. Charlesworth, P.S. 1988. Measurement of air exchange rates. Chapter 2 in Air exchange rate and airtightness measurement techniques—An applications guide. International Energy Agency Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. Chastain, J.P. 1987. Pressure gradients and the location of the neutral pressure axis for low-rise structures under pure stack conditions. Unpublished M.S. thesis. University of Kentucky, Lexington. Chastain, J.P., and D.G. Colliver. 1989. Influence of temperature stratification on pressure differences resulting from the infiltration stack effect. ASHRAE Transactions 95(1):256-268. Paper 3230. Chastain, J.P., D.G. Colliver, and P.W. Winner, Jr. 1987. Computation of discharge coefficients for laminar flow in rectangular and circular openings. ASHRAE Transactions 93(2):2259-2283. Paper NT-87-27-1. CHBA. 1994. HOT2000 technical manual. Canadian Home Builders Association, Ottawa, ON. CIBSE. 2000. Testing buildings for air leakage. Standard TM-23. Chartered Institution of Building Services Engineers, London. CIBSE. 2005. Natural ventilation in non-domestic buildings. Chartered Institution of Building Services Engineers, London. Claridge, D.E., and S. Bhattacharyya. 1990. The measured impact of infiltration in a test cell. ASME Journal of Solar Energy Engineering 112: 123-126.
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Ventilation and Infiltration Claridge, D.E., M. Krarti, and S. Bhattacharyya. 1988. Preliminary measurements of the energy impact of infiltration in a test cell. Proceedings of the Fifth Annual Symposium on Improving Building Energy Efficiency in Hot and Humid Climates, pp. 308-317. Cole, J.T., T.S. Zawacki, R.H. Elkins, J.W. Zimmer, and R.A. Macriss. 1980. Application of a generalized model of air infiltration to existing homes. ASHRAE Transactions 86(2):765. Paper DV-80-09-2. Collet, P.F. 1981. Continuous measurements of air infiltration in occupied dwellings. Proceedings of the 2nd IEA Conference of the Air Infiltration Centre, Stockholm, p. 147. CSA. 2016. Residential carbon monoxide alarming devices. CAN/CSA Standard C6.19-01 (R2016). Canadian Standards Association, Toronto. CSA. 2014. Residential mechanical ventilation systems. CAN/CSA-F326M91 (R2014). Canadian Standards Association, Toronto. Cummings, J.B., and J.J. Tooley, Jr. 1989. Infiltration and pressure differences induced by forced air systems in Florida residences. ASHRAE Transactions 96(20):551-560. Paper VA-89-05-4. Cummings, J.B., J.J. Tooley, Jr., and R. Dunsmore. 1990. Impacts of duct leakage on infiltration rates, space conditioning energy use and peak electrical demand in Florida homes. Proceedings of the ACEEE Summer Study, Pacific Grove, CA. American Council for an Energy-Efficient Economy, Washington, D.C. Cutter. 1987. Air-to-air heat exchangers. In Energy design update. Cutter Information Corporation, Arlington, MA. Desrochers, D., and A.G. Scott. 1985. Residential ventilation rates and indoor radon daughter levels. Transactions of the APCA Specialty Conference, Indoor Air Quality in Cold Climates: Hazards and Abatement Measures, Ottawa, p. 362. Diamond, R.C., J.B. Dickinson, R.D. Lipschutz, B. O’Regan, and B. Shohl. 1982. The house doctor’s manual. Report PUB-3017. Lawrence Berkeley National Laboratory, Berkeley, CA. Diamond, R.C., M.P. Modera, and H.E. Feustel. 1986. Ventilation and occupant behaviour in two apartment buildings. Proceedings of the 7th IEA Conference of the Air Infiltration and Ventilation Centre, Stratford-uponAvon, U.K. Report LBL-21862. Lawrence Berkeley National Laboratory, Berkeley, CA. Dickerhoff, D.J., D.T. Grimsrud, and R.D. Lipschutz. 1982. Component leakage testing in residential buildings. Proceedings of the American Council for an Energy-Efficient Economy, 1982 Summer Study, Santa Cruz, CA. Report LBL 14735. Lawrence Berkeley National Laboratory, Berkeley, CA. Dietz, R.N., R.W. Goodrich, E.A. Cote, and R.F. Wieser. 1986. Detailed description and performance of a passive perfluorocarbon tracer system for building ventilation and air exchange measurement. In Measured air leakage of buildings, STP 904, p. 203. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Dols, W. S., L. Wang, S. J. Emmerich, and B. J. Polidoro. 2014. Development and application of an updated whole-building coupled thermal, airflow and contaminant transport simulation program (TRNSYS/ CONTAM). Journal of Building Performance Simulation: 1-21. D’Ottavio, T.W., G.I. Senum, and R.N. Dietz. 1988. Error analysis techniques for perfluorocarbon tracer derived multizone ventilation rates. Building and Environment 23(40). Ek, C.W., S.A. Anisko, and G.O. Gregg. 1990. Air leakage tests of manufactured housing in the Northwest United States. In Air change rate and airtightness in buildings, STP 1067, pp. 152-164. M.H. Sherman, ed. American Society for Testing and Materials, West Conshohocken, PA. Elmroth, A., and P. Levin. 1983. Air infiltration control in housing. International Energy Agency Air Infiltration Centre, Sint-Stevens-Woluwe, Belgium. Emmerich, S.J. 2006. Simulated performance of natural and hybrid ventilation systems in an office building. HVAC&R Research (now Science and Technology for the Built Environment) 12(4):975-1004. Emmerich, S.J., and A.K. Persily. 2005. Airtightness of commercial buildings in the U.S. Proceedings of the 26th IEA Conference of the Air Infiltration and Ventilation Centre, Brussels, pp 65-70. Emmerich, S.J., and A.K. Persily. 2014. Analysis of US commercial building envelope air leakage database to support sustainable building design. International Journal of Ventilation 12(4):331-344. Available at www .nist.gov/manuscript-publication-search.cfm?pub_id=914293. Emmerich, S.J., J.E. Gorfain, and C. Howard-Reed. 2003. Air and pollutant transport from attached garages to residential living spaces—Literature review and field test. International Journal of Ventilation 2(3):265-276.
16.35 Emmerich, S.J., W. S. Dols, and B. Polidoro. 2012. Application of a natural ventilation system design tool to a school building. Proceeding of SimBuild 2012, pp. 185-192. Available at www.nist.gov/customcf/get _ pdf.cfm?pub_id=911004. Emmerich, S.J., A.K. Persily, and L. Wang. 2013. Modeling and measuring the effects of portable gasoline powered generator exhaust on indoor carbon monoxide level. Technical Note (NIST TN) 1781. U.S. Department of Commerce, National Institute of Standards and Technology. Available at www.nist.gov/manuscript-publication-search.cfm?pub_id=912197. Energy Resource Center. 1982. How to house doctor. University of Illinois, Chicago. EPA. 2003. A standardized EPA protocol for characterizing indoor air quality in large office buildings. U.S. Environmental Protection Agency, Washington, D.C. Available from www.epa.gov/indoor-air-quality -iaq/standardized-epa-protocol-characterizing-indoor-air-quality-large -office. Etheridge, D.W. 1977. Crack flow equations and scale effect. Building and Environment 12:181. Etheridge, D.W., and D.K. Alexander. 1980. The British gas multi-cell model for calculating ventilation. ASHRAE Transactions 86(2):808. Paper DV80-09-4. Etheridge, D.W., and J.A. Nolan. 1979. Ventilation measurements at model scale in a turbulent flow. Building and Environment 14(1):53. Eyre, D., and D. Jennings. 1983. Air-vapour barriers—A general perspective and guidelines for installation. Energy, Mines, and Resources Canada, Ottawa, ON. Farrington, R., D. Martin, and R. Anderson. 1990. A comparison of displacement efficiency, decay time constant, and age of air for isothermal flow in an imperfectly mixed enclosure. Proceedings of the ACEEE 1990 Summer Study on Energy Efficiency in Buildings, pp. 4.35-4.43. American Council for an Energy-Efficient Economy, Washington, D.C. Fennell, H.C., and J. Haehnel. 2005. Setting airtightness standards. ASHRAE Journal 47(9):26-31. Feustel, H.E., and A. Raynor-Hoosen, eds. 1990. Fundamentals of the multizone air flow model—COMIS. Technical Note 29. International Energy Agency Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. FGI. 2014. Guidelines for design and construction of hospital and healthcare facilities. American Institute of Architects, Facilities Guidelines Institute, and U.S. Department of Health and Human Services, Washington, D.C. Fisk, W.J., R.K. Spencer, D.T. Grimsrud, F.J. Offermann, B. Pedersen, and R. Sextro. 1984. Indoor air quality control techniques: A critical review. Report LBL-16493. Lawrence Berkeley National Laboratory, Berkeley, CA. Fisk, W.J., R.J. Prill, and O. Steppanen. 1989. A multi-tracer technique for studying rates of ventilation, air distribution patterns and air exchange efficiencies. Proceedings of Conference on Building Systems—Room Air and Air Contaminant Distribution, pp. 237-240. ASHRAE. Fortmann, R.C., N.L. Nagda, and H.E. Rector. 1990. Comparison of methods for the measurement of air change rates and interzonal airflows to two test residences. In Air change rate and airtightness in buildings, STP 1067, pp. 104-118. M.H. Sherman, ed. American Society of Testing and Materials, West Conshohocken, PA. Foster, M.P., and M.J. Down. 1987. Ventilation of livestock buildings by natural convection. Journal of Agricultural Engineering Research 37:1. Fugler, D. 2004. Garage performance testing. CMHC Research Highlights (April). Canada Mortgage and Housing Corporation, Ottawa, ON. Giesbrecht, P., and G. Proskiw. 1986. An evaluation of the effectiveness of air leakage sealing. In Measured air leakage of buildings, STP 904, p. 312. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Grieve, P.W. 1989. Measuring ventilation using tracer-gases. Brüel and Kjær, Denmark. Grimsrud, D.T., and K.Y. Teichman. 1989. The scientific basis of Standard 62-1989. ASHRAE Journal 31(10):51-54. Grimsrud, D.T., M.H. Sherman, R.C. Diamond, P.E. Condon, and A.H. Rosenfeld. 1979. Infiltration-pressurization correlations: Detailed measurements in a California house. ASHRAE Transactions 85(1):851. Paper PH-79-10-5.
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16.36 Grimsrud, D.T., M.H. Sherman, and R.C. Sonderegger. 1982. Calculating infiltration: Implications for a construction quality standard. Proceedings of the ASHRAE/DOE Conference on the Thermal Performance of the Exterior Envelope of Buildings II, Las Vegas, p. 422. Grot, R.A., and R.E. Clark. 1979. Air leakage characteristics and weatherization techniques for low-income housing. Proceedings of the ASHRAE/ DOE Conference on the Thermal Performance of the Exterior Envelopes of Buildings, p. 178. Orlando, FL. Grot, R.A., and A.K. Persily. 1986. Measured air infiltration and ventilation rates in eight large office buildings. In Measured air leakage of buildings, STP 904, p. 151. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. GSA. 2010. Facilities standards for the public buildings service. PBS-P100. U.S. General Services Administration, Washington, D.C. Hamlin, T.L. 1991. Ventilation and airtightness in new, detached Canadian housing. ASHRAE Transactions 97(2):904-910. Paper IN-91-12-3. Hamlin, T., and W. Pushka. 1994. Predicted and measured air change rates in houses with predictions of occupant IAQ comfort. Proceedings of the 15th IEA Air Infiltration and Ventilation Centre Conference, Buxton, U.K., pp. 771-775. Harrje, D.T., and G.J. Born. 1982. Cataloguing air leakage components in houses. Proceedings of the ACEEE 1982 Summer Study, Santa Cruz, CA. American Council for an Energy-Efficient Economy, Washington, D.C. Harrje, D.T., and T.A. Mills, Jr. 1980. Air infiltration reduction through retrofitting. In Building air change rate and infiltration measurements. STP 719, p. 89. C.M. Hunt, J.C. King, and H.R. Trechsel, eds. American Society for Testing and Materials, West Conshohocken, PA. Harrje, D.T., G.S. Dutt, and J. Beyea. 1979. Locating and eliminating obscure but major energy losses in residential housing. ASHRAE Transactions 85(2):521. Paper DE-79-03-1. Harrje, D.T., R.A. Grot, and D.T. Grimsrud. 1981. Air infiltration site measurement techniques. Proceedings of the 2nd IEA Conference of the Air Infiltration Centre, p. 113. Stockholm, Sweden. Harrje, D.T., G.S. Dutt, D.L. Bohac, and K.J. Gadsby. 1985. Documenting air movements and infiltration in multicell buildings using various tracergas techniques. ASHRAE Transactions 91(2):2012-2027. Paper HI-8540-3. Harrje, D.T., R.N. Dietz, M. Sherman, D.L. Bohac, T.W. D’Ottavio, and D.J. Dickerhoff. 1990. Tracer gas measurement systems compared in a multifamily building. In Air change rate and airtightness in buildings, STP 1067, pp. 5-12. M.H. Sherman, ed. American Society for Testing and Materials, West Conshohocken, PA. Heiselberg, P. 2002. Principles of hybrid ventilation. Final Report, International Energy Agency Energy Conservation in Buildings and Community Systems, Annex 35. Hybrid Ventilation Centre, Aalborg University, Denmark. Hekmat, D., H.E. Feustel, and M.P. Modera. 1986. Impacts of ventilation strategies on energy consumption and indoor air quality in single-family residences. Energy and Buildings 9(3):239. Herrlin, M.K. 1985. MOVECOMP: A static-multicell-airflow-model. ASHRAE Transactions 91(2B):1989. Paper HI-85-40-1. Holton, J.K., M.J. Kokayko, and T.R. Beggs. 1997. Comparative ventilation system evaluations. ASHRAE Transactions 103(2):675-692. Paper PH97-08-1. Honma, H. 1975. Ventilation of dwellings and its disturbances. Faibo Grafiska, Stockholm, Sweden. Hopkins, L.P., and B. Hansford. 1974. Air flow through cracks. Building Service Engineer 42(September):123. Hunt, C.M. 1980. Air infiltration: A review of some existing measurement techniques and data. In Building air change rate and infiltration measurements, STP 719, p. 3. C.M. Hunt, J.C. King, and H.R. Trechsel, eds. American Society for Testing and Materials, West Conshohocken, PA. ICC. 2012a. International energy conservation code® (IECC®). International Code Council, Washington, D.C. ICC. 2012b. International green construction code® (IgCC®). International Code Council, Washington, D.C. ISO. 2006. Thermal performance of buildings—Determination of air permeability of buildings—Fan pressurization method. Standard 9972-2006. International Organization for Standardization, Geneva.
2017 ASHRAE Handbook—Fundamentals (SI) Iwashita, G., K. Kimura., et al. 1989. Pilot study on addition of olf units for perceived air pollution sources. Proceedings of the SHASE Annual Meeting, pp. 3221-3324. Society of Heating, Air-Conditioning and Sanitary Engineers of Japan, Tokyo. Jacobson, D.I., G.S. Dutt, and R.H. Socolow. 1986. Pressurization testing, infiltration reduction, and energy savings. In Measured air leakage of buildings, STP 904, p. 265. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Janssen, J.E. 1989. Ventilation for acceptable indoor air quality. ASHRAE Journal 31(10):40-48. Janu, G.J., J.D. Wegner, and C.G. Nesler. 1995. Outdoor airflow control for VAV systems. ASHRAE Journal 37(4):62-68. Johnson, T., and T. Long. 2005. Determining the frequency of open windows in residences: A pilot study in Durham, North Carolina during varying temperature conditions. Journal of Exposure Analysis and Environmental Epidemiology 15(4):329-349. Judkoff, R., J.D. Balcomb, C.E. Handcock, G. Barker, and K. Subbarao. 1997. Side-by-side thermal tests of modular offices: A validation study of the STEM method. Report. National Renewable Energy Laboratory, Golden, CO. Jump, D.A., I.S. Walker, and M.P. Modera. 1996. Field measurements of efficiency and duct retrofit effectiveness in residential forced air distribution systems. Proceedings of the 1996 ACEEE Summer Study, pp. 1.147-1.156. American Council for an Energy-Efficient Economy, Washington, D.C. Kiel, D.E., and D.J. Wilson. 1986. Gravity driven airflows through open doors, 15.1. Proceedings of the 7th IEA Conference of the Air Infiltration and Ventilation Centre, Stratford-upon-Avon, U.K. Kim, A.K., and C.Y. Shaw. 1986. Seasonal variation in airtightness of two detached houses. In Measured air leakage of buildings, STP 904, p. 17. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Klauss, A.K., R.H. Tull, L.M. Roots, and J.R. Pfafflin. 1970. History of the changing concepts in ventilation requirements. ASHRAE Journal 12(6):51-55. Klote, J.H., J.A. Milke, P.G. Turnbull, A. Kashef, and M.J. Ferreira. 2012. Handbook of smoke control engineering. ASHRAE. Kohonen, R., T. Ojanen, and M. Virtanen. 1987. Thermal coupling of leakage flows and heating load of buildings. Proceedings of the 8th IEA Air Infiltration and Ventilation Centre Conference, Überlingen, Germany, pp. 10.1-10.22. Kreith, F., and R. Eisenstadt. 1957. Pressure drop and flow characteristics of short capillary tubes at low Reynolds numbers. ASME Transactions, pp. 1070-1078. Kronvall, J. 1980. Correlating pressurization and infiltration rate data— Tests of an heuristic model. Lund Institute of Technology, Division of Building Technology, Lund, Sweden. Kumar, R., A.D. Ireson, and H.W. Orr. 1979. An automated air infiltration measuring system using SF6 tracer gas in constant concentration and decay methods. ASHRAE Transactions 85(2):385. Paper DE-2553. Kvisgaard, B., and P.F. Collet. 1990. The user’s influence on air change. In Air change rate and airtightness in buildings, STP 1067, pp. 67-76. M.H. Sherman, ed. American Society for Testing and Materials, West Conshohocken, PA. Lagus, P.L. 1989. Tracer measurement instrumentation suitable for infiltration, air leakage, and airflow pattern characterization. Proceedings of the Conference on Building Systems—Room Air and Air Contaminant Distribution, pp. 97-102. ASHRAE. Lagus, P., and A.K. Persily. 1985. A review of tracer-gas techniques for measuring airflows in buildings. ASHRAE Transactions 91(2B):1075. Paper HI-85-2-1. Lecompte, J.G.N. 1987. The influence of natural convection in an insulated cavity on thermal performance of a wall. In Insulation materials, testing, and applications. American Society for Testing and Materials, West Conshohocken, PA. Li, Y., and P. Heiselberg. 2003. Analysis methods for natural and hybrid ventilation—A critical literature review and recent developments. International Journal of Ventilation 1(4):3-20. Liddament, M.W. 1988. The calculation of wind effect on ventilation. ASHRAE Transactions 94(2):1645-1660. Paper OT-88-13-1.
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Ventilation and Infiltration Liddament, M., and C. Allen. 1983. The validation and comparison of mathematical models of air infiltration. Technical Note 11. International Energy Agency Air Infiltration and Ventilation Centre, Sint-StevensWoluwe, Belgium. Liu, M., and D.E. Claridge. 1992a. The measured energy impact of infiltration under dynamic conditions. Proceedings of the 8th Symposium on Improving Building Systems in Hot and Humid Climates, Dallas. Liu, M., and D.E. Claridge. 1992b. The measured energy impact of infiltration in a test cell. Proceedings of the 8th Symposium on Improving Building Systems in Hot and Humid Climates, Dallas. Liu, M., and D.E. Claridge. 1992c. The energy impact of combined solar radiation/infiltration/conduction effects in walls and attics. Proceedings of the Thermal Performance of Exterior Envelopes of Buildings, 5th ASHRAE/DOE/BTECC Conference, Clearwater Beach, FL. Liu, M., and D.E. Claridge. 1995. Experimental methods for identifying infiltration heat recovery in building. Proceedings of the Thermal Performance of Exterior Envelopes of Buildings, 6th ASHRAE/DOE/BTECC Conference, Clearwater Beach, FL. Lubliner, M., D.T. Stevens, and B. Davis. 1997. Mechanical ventilation in HUD-code manufactured housing in the Pacific Northwest. ASHRAE Transactions 103(1):693-705. Paper PH-97-08-2. Marbek Resource Consultants. 1984. Air sealing homes for energy conservation. Energy, Mines and Resources Canada, Buildings Energy Technology Transfer Program, Ottawa, ON. McDowell, T.P., S. Emmerich, J.W. Thornton, and G. Walton. 2003. Integration of airflow and energy simulation using CONTAM and TRNSYS. ASHRAE Transactions 109(2):757-770. Paper KC-03-10-2. McWilliams, J., and M. Sherman. 2005. Review of literature related to residential ventilation requirements. Paper LBNL-57236. Lawrence Berkeley National Laboratory, Berkeley, CA. Mendell, M.J. 1993. Non-specific symptoms in office workers: A review and summary of the epidemiologic literature. Indoor Air 3 (4):227-236. Modera, M.P. 1989. Residential duct system leakage: Magnitude, impacts, and potential for reduction. ASHRAE Transactions 96(2):561-569. Paper VA-89-05-5. Modera, M.P., and D.J. Wilson. 1990. The effects of wind on residential building leakage measurements. In Air change rate and airtightness in buildings, STP 1067, pp. 132-145. M.H. Sherman, ed. Lawrence Berkeley National Laboratory, Berkeley, CA. Report LBL-24195. Modera, M.P., D. Dickerhoff, R. Jansky, and B. Smith. 1991. Improving the energy efficiency of residential air distribution systems in California. Report LBL-30866. Lawrence Berkeley National Laboratory, Berkeley, CA. Mumma, S.A., and K.M. Shank. 2001. Achieving dry outside air in an energy efficient manner. ASHRAE Transactions 107(1):553-561. Paper AT-01-07-2. Mumma, S.A., and Y.M. Wong. 1990. Analytical evaluation of outdoor airflow rate variation vs. supply airflow rate variation in VAV systems when the outside air damper position is fixed. ASHRAE Transactions 90(1):1197-1208. Murphy, W.E., D.G. Colliver, and L.R. Piercy. 1991. Repeatability and reproducibility of fan pressurization devices in measuring building air leakage. ASHRAE Transactions 97(2):885-895. Paper IN-91-12-1. Nelson, B.D., D.A. Robinson, and G.D. Nelson. 1985. Designing the envelope—Guidelines for buildings (SP-49). Proceedings of the ASHRAE/ DOE/BTECC Conference—Thermal Performance of the Exterior Envelopes of Buildings III, Florida, pp. 1117-1122. Ng, L.C., A. Musser, A.K. Persily, and S.J. Emmerich. 2012. Airflow and indoor air quality models of DOE reference commercial buildings. Technical Note 1734. National Institute of Standards and Technology, Gaithersburg, MD. NRCC. 2010. National Building Code of Canada. National Research Council of Canada, Ottawa, ON. Nylund, P.O. 1980. Infiltration and ventilation. Report D22:1980. Swedish Council for Building Research, Stockholm. Offermann, F., and D. Int-Hout. 1989. Ventilation effectiveness measurements of three supply/return air configurations. Environment International 15(1-6):585-592. Orme, M. 1999. Applicable models for air infiltration and ventilation calculations. Technical Note 51. International Energy Agency Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. Ormerod, R. 1983. Nuclear shelters: A guide to design. Architectural Press, London.
16.37 Palmiter, L., and T. Bond. 1994. Modeled and measured infiltration II—A detailed case study of three homes. Report TR 102511. Electric Power Research Institute, Palo Alto, CA. Palmiter, L., and I. Brown. 1989. The Northwest residential infiltration survey: Description and summary of results. Proceedings of the ASHRAE/ DOE/BTECC/CIBSE Conference—Thermal Performance of the Exterior Envelopes of Buildings IV, Florida, pp. 445-457. Palmiter, L., I.A. Brown, and T.C. Bond. 1991. Measured infiltration and ventilation in 472 all-electric homes. ASHRAE Transactions 97(2):979987. Paper IN-91-15-3. Parekh, A., K. Ruest, and M. Jacobs. 1991. Comparison of airtightness, indoor air quality and power consumption before and after air-sealing of high-rise residential buildings. Proceedings of the 12th IEA Conference of the Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. Parker, G.B., M. McSorley, and J. Harris. 1990. The Northwest residential infiltration survey: A field study of ventilation in new houses in the Pacific Northwest. In Air change rate and airtightness in buildings, STP 1067, pp. 93-103. M.H. Sherman, ed. American Society for Testing and Materials, West Conshohocken, PA. Persily, A. 1982. Repeatability and accuracy of pressurization testing. Proceedings of the ASHRAE/DOE Conference, Thermal Performance of the Exterior Envelopes of Buildings II, Las Vegas. Persily, A.K. 1986. Measurements of air infiltration and airtightness in passive solar homes. In Measured air leakage of buildings, STP 904, p. 46. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Persily, A.K. 1988. Tracer gas techniques for studying building air exchange. Report NBSIR 88-3708. National Institute of Standards and Technology, Gaithersburg, MD. Persily, A.K. 1991. Design guidelines for thermal envelope integrity in office buildings. Proceedings of the 12th IEA Conference of the Air Infiltration and Ventilation Centre, Ottawa, ON. Persily, A.K. 2004. Building ventilation and pressurization as a security tool. ASHRAE Journal 46 (9):18-24. Persily, A.K., and J. Axley. 1990. Measuring airflow rates with pulse tracer techniques. In Air change rate and airtightness in buildings, pp. 31-51. STP 1067. M.H. Sherman, ed. American Society for Testing and Materials, West Conshohocken, PA. Persily, A.K., and R.A. Grot. 1985a. The airtightness of office building envelopes. Proceedings of the ASHRAE/DOE/BTECC Conference on the Thermal Performance of the Exterior Envelopes of Buildings III, Clearwater Beach, FL, p. 125. Persily, A.K., and R.A. Grot. 1985b. Accuracy in pressurization data analysis. ASHRAE Transactions 91(2B):105. Paper HI-85-03-2. Persily, A.K., and R.A. Grot. 1986. Pressurization testing of federal buildings. In Measured air leakage of buildings, STP 904, p. 184. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Persily, A.K., and G.T. Linteris. 1983. A comparison of measured and predicted infiltration rates. ASHRAE Transactions 89(2):183. Paper DC-8304-1. Persily, A.K., J. Gorfain, and G. Brunner. 2005. Ventilation design and performance in U.S. office buildings. ASHRAE Journal 47(4):30-35. Persily, A.K., R.E. Chapman, S. Emmerich, W.S. Dols, H. Davis, P. Lavappa, and A. Rushing. 2007. Building retrofits for increased protection against airborne chemical and biological releases. NISTIR Report 7379. National Institute of Standards and Technology, Gaithersburg, MD. Persily, A.K., A. Musser, and S. Emmerich. 2010. Modeled infiltration rate distributions for U.S. housing. Indoor Air 20(6):473-485. Powell, F., M. Krarti, and A. Tuluca. 1989. Air movement influence on the effective thermal resistance of porous insulations: A literature survey. Journal of Thermal Insulation 12:239-251. Price, P.N., and M.H. Sherman. 2006. Ventilation behavior and household characteristics in new California houses. Report LBNL-59620. RDH. 2013. Air leakage control in multi-unit residential buildings. Canada Mortgage and Housing Corporation. Reardon, J.T., and C.-Y. Shaw. 1997. Evaluation of five simple ventilation strategies suitable for houses without forced-air heating. ASHRAE Transactions 103(1):731-744. Paper PH-97-08-5. Reeves, G., M.F. McBride, and C.F. Sepsy. 1979. Air infiltration model for residences. ASHRAE Transactions 85(1):667. Paper PH-79-07A-1.
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16.38 Riley, M. 1990. Indoor air quality and energy conservation: The R-2000 home program experience. Proceedings of Indoor Air ’90: International Conference on Indoor Air Quality and Climate, Ottawa, vol. 5, p. 143. Robison, P.E., and L.A. Lambert. 1989. Field investigation of residential infiltration and heating duct leakage. ASHRAE Transactions 95(2):542550. Paper VA-89-05-3. Rock, B.A. 1992. Characterization of transient pollutant transport, dilution, and removal for the study of indoor air quality. Ph.D. dissertation, University of Colorado at Boulder. University Microfilms International. Rock, B.A. 2005. A user-friendly model and coefficients for slab-on-grade load and energy calculations. ASHRAE Transactions 111(2):122-136. Paper 4795. Rock, B.A. 2006. Ventilation for environmental tobacco smoke. Elsevier Science, New York, and ASHRAE. Rock, B.A., and D. Zhu. 2002. Designer’s guide to ceiling-based air diffusion. ASHRAE. Rock, B.A., M.J. Brandemuehl, and R. Anderson. 1995. Toward a simplified design method for determining the air change effectiveness. ASHRAE Transactions 101(1):217-227. Paper 3852. Rudd, A.F. 1998. Design/sizing methodology and economic evaluation of central-fan-integrated supply ventilation systems. ACEEE 1998 Summer Study on Energy Efficiency in Buildings. 23-28 August, Pacific Grove, CA. American Council for an Energy Efficient Economy, Washington, D.C. Russell, M., M. Sherman, and A. Rudd. 2005. Review of residential ventilation technologies. Paper LBNL-576. Lawrence Berkeley National Laboratory, Berkeley, CA. Sandberg, M.H. 1981. What is ventilation efficiency? Building and Environment 16:123-135. Seppanen, O.A., W.J. Fisk and M.J. Mendell. 1999. Association of ventilation rates and CO2 concentrations with health and other responses in commercial and institutional buildings. Indoor Air 9(4):226-252. Shaw, C.Y. 1981. A correlation between air infiltration and air tightness for a house in a developed residential area. ASHRAE Transactions 87(2):333. Paper CI-2655. Shaw, C.Y., and W.C. Brown. 1982. Effect of a gas furnace chimney on the air leakage characteristic of a two-story detached house. Proceedings of the 3rd IEA Conference of the Air Infiltration Centre, London. Sherman, M.H. 1986. Infiltration degree-days: A statistic for quantifying infiltration-related climate. ASHRAE Transactions 92(2):161-181. Paper 2986. Sherman, M.H. 1987. Estimation of infiltration from leakage and climate indications. Energy and Buildings 10(1):81. Sherman, M.H. 1989a. Uncertainty in airflow calculations using tracer gas measurements. Building and Environment 24(4):347-354. Sherman, M.H. 1989b. On the estimation of multizone ventilation rates from tracer gas measurements. Building and Environment 24(4):355-362. Sherman, M.H. 1990. Tracer gas techniques for measuring ventilation in a single zone. Building and Environment 25(4):365-374. Sherman, M.H. 1991. Single-zone stack-dominated infiltration modeling. Proceedings of the 12th IEA Conference of the Air Infiltration and Ventilation Centre, Ottawa, ON, pp. 297-314. Sherman, M.H. 1992a. A power law formulation of laminar flow in short pipes. Journal of Fluids Engineering 114:601-605. Report LBL-29414, Lawrence Berkeley National Laboratory, Berkeley, CA. Sherman, M.H. 1992b. Superposition in infiltration modeling. Indoor Air 2:101-114. Sherman, M.H., and D. Dickerhoff. 1989. Description of the LBL multitracer measurement system. Proceedings of the ASHRAE/DOE/BTECC/ CIBSE Conference—Thermal Performance of the Exterior Envelopes of Buildings IV, pp. 417-432. Sherman, M.H., and D.J. Dickerhoff. 1998. Airtightness of U.S. dwellings. ASHRAE Transactions 104(2):1359-1367. Paper TO-98-25-1. Sherman, M.H., and D.T. Grimsrud. 1980. Infiltration-pressurization correlation: Simplified physical modeling. ASHRAE Transactions 86(2):778. Paper DV-80-09-3. Sherman, M.H., and N. Matson. 1997. Residential ventilation and energy characteristics. ASHRAE Transactions 103(1):717-730. Paper PH-9708-4. Sherman, M.H., and M.P. Modera. 1986. Comparison of measured and predicted infiltration using the LBL infiltration model. In Measured air leakage of buildings, STP 904, p. 325. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA.
2017 ASHRAE Handbook—Fundamentals (SI) Sherman, M.H., and D.J. Wilson. 1986. Relating actual and effective ventilation in determining indoor air quality. Building and Environment 21(3/ 4):135. Sherman, M.H., D.T. Grimsrud, P.E. Condon, and B.V. Smith. 1980. Air infiltration measurement techniques. Proceedings of the 1st IEA Conference of the Air Infiltration Centre, London. Report LBL-10705. Lawrence Berkeley National Laboratory, Berkeley, CA. Sibbitt, B.E., and T. Hamlin. 1991. Meeting Canadian residential ventilation standard requirements with low-cost systems. Canada Mortgage and Housing Corporation, Ottawa, ON. Sinden, F.W. 1978a. Wind, temperature and natural ventilation—Theoretical considerations. Energy and Buildings 1(3):275. Sinden, F.W. 1978b. Multi-chamber theory of air infiltration. Building and Environment 13:21-28. Sorensen, J.H., and B.M. Vogt. 2001. Will duct tape and plastic really work? Issues related to expedient sheltering-in-place. Report ORNL/TM-2001/ 154. Oak Ridge National Laboratory, Oak Ridge, TN. Sundell, J., H. Levin, W.W. Nazaroff, W.S. Cain, W.J. Fisk, D.T. Grimsrud, and F. Gyntelberg. 2011. Ventilation rates and health: Multidisciplinary review of the scientific literature. Indoor Air 21(3):191-204. Tamura, G.T., and C.Y. Shaw. 1976a. Studies on exterior wall air tightness and air infiltration of tall buildings. ASHRAE Transactions 82(1):122. Paper DA-2388. Tamura, G.T., and C.Y. Shaw. 1976b. Air leakage data for the design of elevator and stair shaft pressurization system. ASHRAE Transactions 82(2):179. Paper SE-2413. Tamura, G.T., and A.G. Wilson. 1966. Pressure differences for a nine-story building as a result of chimney effect and ventilation system operation. ASHRAE Transactions 72(1):180. Tamura, G.T., and A.G. Wilson. 1967a. Pressure differences caused by chimney effect in three high buildings. ASHRAE Transactions 73(2):II.1.1. Tamura, G.T., and A.G. Wilson. 1967b. Building pressures caused by chimney action and mechanical ventilation. ASHRAE Transactions 73(2):II.2.1. Timusk, J., A.L. Seskus, and K. Linger. 1992. A systems approach to extend the limit of envelope performance. Proceedings of the 6th ASHRAE/ DOE/BTECC Conference—Thermal Performance of Exterior Envelopes of Buildings, Clearwater Beach, FL. Turk, B.T., D.T. Grimsrud, J.T. Brown, K.L. Geisling-Sobotka, J. Harrison, and R.J. Prill. 1989. Commercial building ventilation rates and particle concentrations. ASHRAE Transactions 95(1):422-433. Paper 3248. USACE. 2009. U.S. Army Corps of Engineers air leakage test protocol for building envelopes. Air Barrier Association of America, Walpole, MA, and U.S. Army Corps of Engineers. Verschoor, J.D., and J.O. Collins. 1986. Demonstration of air leakage reduction program in navy family housing. In Measured air leakage of buildings, STP 904, p. 294. H.R. Trechsel and P.L. Lagus, eds. American Society for Testing and Materials, West Conshohocken, PA. Walker, I.S. 1999. Distribution system leakage impacts on apartment building ventilation rates. ASHRAE Transactions 105(1):943-950. Paper CH99-14-2. Walker, I.S., and T.W. Forest. 1995. Field measurements of ventilation rates in attics. Building and Environment 30(3):333-347. Walker, I.S., and D.J. Wilson. 1993. Evaluating models for superposition of wind and stack effects in air infiltration. Building and Environment 28(2):201-210. Walker, I.S., and D.J. Wilson. 1994. Practical methods for improving estimates of natural ventilation rates. Proceedings of the 15th IEA Conference of the Air Infiltration and Ventilation Centre, Buxton, U.K., pp. 517526. Walker, I.S., and D.J. Wilson. 1998. Field validation of algebraic equations for stack and wind driven air infiltration calculations. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 4(2):119-140. Walker, I.S., D.J. Wilson., and M.H. Sherman. 1997. A comparison of the power law to quadratic formulations for air infiltration calculations. Energy and Buildings 27(3). Walker, I., M. Sherman, J. Siegel, D. Wang, C. Buchanan, and M. Modera. 1999. Leakage diagnostics, sealant longevity, sizing and technology transfer in residential thermal distribution systems: Part II. Report LBNL-42691.
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Ventilation and Infiltration
16.39
Walton, G.N. 1984. A computer algorithm for predicting infiltration and interroom airflows. ASHRAE Transactions 90(1B):601. Paper AT-8411-3. Walton, G.N. 1989. Airflow network models for element-based building airflow modeling. ASHRAE Transactions 95(2):611-620. Paper VA-89-065. Walton, G., and W.S. Dols. 2003. CONTAM 2.1 supplemental user guide and program documentation. NISTIR Report 7049, National Institute of Standards and Technology, Gaithersburg, MD. Warden, D. 1995. Outdoor air: Calculation and delivery. ASHRAE Journal 37(6):54-63. Warren, P.R., and B.C. Webb. 1980. The relationship between tracer gas and pressurization techniques in dwellings. Proceedings of the 1st IEA Conference of the Air Infiltration Centre, London. Warren, P.R., and B.C. Webb. 1986. Ventilation measurements in housing. CIBSE Symposium, Natural Ventilation by Design. Chartered Institution of Building Services Engineers, London. Weidt, J.L., J. Weidt, and S. Selkowitz. 1979. Field air leakage of newly installed residential windows. Proceedings of the ASHRAE/DOE Conference—Thermal Performance of the Exterior Envelopes of Buildings Orlando, FL, p. 149. Weschler, C.J. 2000. Ozone in indoor environments: Concentration and chemistry. Indoor Air 10:269-288. Wilson, D.J., and I.S. Walker. 1991. Wind shelter effects on air infiltration for a row of houses. Proceedings of the 12th IEA Conference of the Air Infiltration and Ventilation Centre, Ottawa, ON, pp. 335-346. Wilson, D.J., and I.S. Walker. 1992. Feasibility of passive ventilation by constant area vents to maintain indoor air quality in houses. Proceedings of Indoor Air Quality ’92, ASHRAE/ACGIH/AIHA Conference, San Francisco. Wilson, D.J., and I.S. Walker. 1993. Infiltration data from the Alberta Home Heating Research Facility. Technical Note 41. Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. Wiren, B.G. 1984. Wind pressure distributions and ventilation losses for a single-family house as influenced by surrounding buildings—A wind tunnel study. Proceedings of the Air Infiltration Centre Wind Pressure Workshop, Brussels, pp. 75-101. Wolf, S. 1966. A theory of the effects of convective air flow through fibrous thermal insulation. ASHRAE Transactions 72(1):III 2.1-III 2.9. Yaglou, C.P., and W.N. Witheridge. 1937. Ventilation requirements. ASHVE Transactions 43:423. Yaglou, C.P., E.C. Riley, and D.I. Coggins. 1936. Ventilation requirements. ASHVE Transactions 42:133. Yuill, G.K. 1986. The variation of the effective natural ventilation rate with weather conditions. Proceedings of the Solar Energy Society of Canada Renewable Energy Conference ’86, pp. 70-75. Yuill, G.K. 1991. The development of a method of determining air change rates in detached dwellings for assessing indoor air quality. ASHRAE Transactions 97(2):896-903. Paper IN-91-12-2. Yuill, G.K. 1996. Impact of high use automatic doors on infiltration. ASHRAE Research Project RP-763, Final Report.
Yuill, G.K., and G.M. Comeau. 1989. Investigation of the indoor air quality, air tightness and air infiltration rates of a random sample of 78 houses in Winnipeg. Proceedings of IAQ ’89, The Human Equation—Health and Comfort, pp. 122-127. ASHRAE. Yuill, G.K., M.R. Jeanson, and C.P. Wray. 1991. Simulated performance of demand-controlled ventilation systems using carbon dioxide as an occupancy indicator. ASHRAE Transactions 97(2):963-968. Paper IN-91-151. Yuill, D.P., G.K. Yuill, and A.H. Coward. 2008. Measurement and analysis of vitiation of secondary air in air distribution systems (RP-1276). HVAC&R Research (now Science and Technology for the Built Environment) 14(3):345-357. Yuill, D.P., G.K. Yuill, and A.H. Coward. 2012. Experimental validation of the multiple-zone system ventilation efficiency equation of ANSI/ ASHRAE Standard 62.1 (RP-1276). HVAC&R Research (now Science and Technology for the Built Environment) 18(3). Zhivov, A. 2013. Air tightness in new and retrofitted US army buildings. Air Barrier Association of America, Walpole, MA.
BIBLIOGRAPHY AIVC. [Ongoing]. AIRBASE bibliographic database. International Energy Agency Air Infiltration and Ventilation Centre, Sint-Stevens-Woluwe, Belgium. www.aivc.org/resources/airbase. Brennan, T., W. Turner, G. Fisher, B. Thompson, and B. Ligman. 1992. Fan pressurization of school buildings. Proceedings of Thermal Performance of the Exterior Envelopes of Buildings V, pp. 643-645. ASHRAE. Colliver, D.G., W.E. Murphy, and W. Sun. 1992. Evaluation of the techniques for the measurement of air leakage of building components. ASHRAE Research Project RP-438, Final Report. University of Kentucky, Lexington. Lamming, S., and J. Salmon. 1998. Wind data for design of smoke control systems (RP-816). ASHRAE Transactions 104(1):742-751. Paper 4171. Lstiburek, J.W. 2005. Understanding air barriers. ASHRAE Journal 47(7): 24-30. Persily, A.K., W.S. Dols, S.J. Nabinger, and S. Kirchner. 1991. Preliminary results of the environmental evaluation of the Federal Records Center in Overland, Missouri. NISTIR Report 4634. National Institute of Standards and Technology, Gaithersburg, MD. Potter, N. 2001. Air tightness testing—A guide for clients and contractors. Technical Note 19/2001. Building Services Research and Information Association, Bracknell, U.K. Sherman, M.H. 1995. The use of blower door data. Indoor Air 5:215-224. Yuill, D.P., G.K. Yuill, and A.H. Coward. 2007. A study of multiple space effects on ventilation system efficiency in Standard 62.1-2004 and experimental validation of the multiple spaces equation. ASHRAE Research Project RP-1276, Final Report. Yuill, D.P., G.K. Yuill, and A.H. Coward. 2009. Experimental validation of the multiple spaces equation of ASHRAE Standard 62.1. ASHRAE Research Project RP-1276, Final Report.
Related Commercial Resources
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Related Commercial Resources CHAPTER 17
RESIDENTIAL COOLING AND HEATING LOAD CALCULATIONS Residential Features ................................................................ Calculation Approach.............................................................. Other Methods ......................................................................... Residential Heat Balance (RHB) Method ................................ Residential Load Factor (RLF) Method ..................................
17.1 17.1 17.2 17.2 17.2
Common Data and Procedures ................................................ 17.2 Cooling Load............................................................................ 17.8 Heating Load.......................................................................... 17.11 Load Calculation Example..................................................... 17.12 Symbols .................................................................................. 17.14
HIS chapter covers cooling and heating load calculation procedures for residential buildings, including detailed heatbalance methods that serve as the basis for cooling load calculation. Simple cooling load procedures, suitable for hand calculations, are provided for typical cases. Straightforward heating load calculation procedures are also included. Procedures in this chapter are based on the same fundamentals as the nonresidential methods in Chapter 18. However, many characteristics distinguish residential loads, and Chapter 18’s procedures should be applied with care to residential applications. Additional information about residential heating and cooling is found in Chapter 1 of the 2015 ASHRAE Handbook—HVAC Applications and Chapter 10 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment.
• Dehumidification Issues. Dehumidification occurs during cooling unit operation only, and space condition control is usually limited to use of room thermostats (sensible heat-actuated devices). Excessive sensible capacity results in short-cycling and severely degraded dehumidification performance.
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T
1.
RESIDENTIAL FEATURES
With respect to heating and cooling load calculation and equipment sizing, the following unique features distinguish residences from other types of buildings: • Smaller Internal Heat Gains. Residential system loads are primarily imposed by heat gain or loss through structural components and by air leakage or ventilation. Internal heat gains, particularly those from occupants and lights, are small compared to those in commercial or industrial structures. • Varied Use of Spaces. Use of spaces in residences is more flexible than in commercial buildings. Localized or temporary temperature excursions are often tolerable. • Fewer Zones. Residences are generally conditioned as a single zone or, at most, a few zones. Typically, a thermostat located in one room controls unit output for multiple rooms, and capacity cannot be redistributed from one area to another as loads change over the day. This results in some hour-to-hour temperature variation or swing that has a significant moderating effect on peak loads, because of heat storage in building components. • Greater Distribution Losses. Residential ducts are frequently installed in attics or other unconditioned buffer spaces. Duct leakage and heat gain or loss can require significant increases in unit capacity. Residential distribution gains and losses cannot be neglected or estimated with simple rules of thumb. • Partial Loads. Most residential cooling systems use units of relatively small capacity (about 5 to 18 kW cooling, 18 to 32 kW heating). Because loads are largely determined by outdoor conditions, and few days each season are design days, the unit operates at partial load during most of the season; thus, an oversized unit is detrimental to good system performance, especially for cooling in areas of high wet-bulb temperature. The preparation of this chapter is assigned to TC 4.1, Load Calculation Data and Procedures.
17.1 Copyright © 2017, ASHRAE
In addition to these general features, residential buildings can be categorized according to their exposure: • Single-Family Detached. A house in this category usually has exposed walls in four directions, often more than one story, and a roof. The cooling system is a single-zone, unitary system with a single thermostat. Two-story houses may have a separate cooling system for each floor. Rooms are reasonably open and generally have a centralized air return. In this configuration, both air and load from rooms are mixed, and a load-leveling effect, which requires a distribution of air to each room that is different from a pure commercial system, results. Because the amount of air supplied to each room is based on the load for that room, proper load calculation procedures must be used. • Multifamily. Unlike single-family detached units, multifamily units generally do not have exposed surfaces facing in all directions. Rather, each unit typically has a maximum of three exposed walls and possibly a roof. Each living unit has a single unitary cooling system or a single fan-coil unit and the rooms are relatively open to one another. This configuration does not have the same load-leveling effect as a single-family detached house. • Other. Many buildings do not fall into either of the preceding categories. Critical to the designation of a single-family detached building is well-distributed exposure so there is not a shortduration peak; however, if fenestration exposure is predominantly east or west, the cooling load profile resembles that of a multifamily unit. On the other hand, multifamily units with both east and west exposures or neither east nor west exposure exhibit load profiles similar to single-family detached.
2.
CALCULATION APPROACH
Variations in the characteristics of residences can lead to surprisingly complex load calculations. Time-varying heat flows combine to produce a time-varying load. The relative magnitude and pattern of the heat flows depends on the building characteristics and exposure, resulting in a building-specific load profile. In general, an hourby-hour analysis is required to determine that profile and find its peak. In theory, cooling and heating processes are identical; a common analysis procedure should apply to either. Acceptable simplifications are possible for heating; however, for cooling, different approaches are used. Heating calculations use simple worst-case assumptions: no solar or internal gains, and no heat storage (with all heat losses evaluated instantaneously). With these simplifications, the heating problem is reduced to a basic UA t calculation. The heating
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17.2
2017 ASHRAE Handbook—Fundamentals (SI)
procedures in this chapter use this long-accepted approach, and thus differ only in details from prior methods put forth by ASHRAE and others. The cooling procedures in this chapter were extensively revised in 2005, based on the results of ASHRAE research project RP-1199, also supported by the Air-Conditioning Contractors of America (ACCA) (Barnaby et al. 2004, 2005). Although the complexity of residential cooling load calculations has been understood for decades, prior methods used a cooling load temperature difference/cooling load factor (CLTD/CLF) form requiring only hand-tractable arithmetic. Without such simplification, the procedures would not have been used; an approximate calculation was preferable to none at all. The simplified approaches were developed using detailed computer models and/or empirical data, but only the simplifications were published. Now that computing power is routinely available, it is appropriate to promulgate 24 h, equation-based procedures.
3.
OTHER METHODS
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Several residential load calculation methods have been published in North America over the last 30 years. All use the UAt heating formulation and some variation of the CLTD/CLF approach for cooling. • ACCA. Manual J, 8th edition (ACCA 2016) is widely used in the United States. Cooling loads are calculated using semiempirical heat gain factors derived from experimental data taken at the University of Illinois in the 1950s. These factors, associated overview, and references are found in the 1985 and earlier editions of the ASHRAE Handbook—Fundamentals. The 8th edition retains the underlying factors but provides increased flexibility in their application, in addition to other extensions. • ASHRAE. The 1989 to 2001 editions of the ASHRAE Handbook— Fundamentals contain an updated method based on ASHRAE research project RP-342 (McQuiston 1984). In this work, cooling factors were re-derived using a transfer-function building model that included temperature-swing effects. • F280. This Canadian adaptation of the CLTD/CLF procedure (CAN/CSA Standard F280) also uses cooling methods based on ASHRAE RP-342. Heating procedures include detailed ground heat loss estimates. A key common element of all cooling methods is attention to temperature swing, via empirical data or suitable models. Throughout the literature, it is repeatedly emphasized that direct application of nonresidential methods (based on a fixed set point) results in unrealistically high cooling loads for residential applications.
4.
RESIDENTIAL HEAT BALANCE (RHB) METHOD
A 24 h procedure is required to accurately determine the cooling load profile of a residence. The heat balance (HB) method allows detailed simulation of space temperatures and heat flows. ASHRAE research project RP-1199 adapted HB to residential applications, resulting in the residential heat balance (RHB) method. Although RHB provides the technical basis for this chapter, it is a computeronly technique and is not documented here. HB is described in Chapter 18 and Pedersen et al. 1998; Barnaby et al. (2004, 2005) document RHB enhancements. RP-1199 produced an implementation of the RHB method, called ResHB (Barnaby et al. 2004). This application is derived from the ASHRAE Toolkit for Building Load Calculations (Pedersen et al. 2001) and has the following features: • Multizone. Whereas the original Toolkit code supported a single zone, ResHB can analyze projects that include multiple systems, zones, and rooms.
• Temperature swing. ResHB calculates cooling load with temperature swing. That is, the code searches for sensible capacity sufficient to hold the space temperature within a specified excursion above the set point. • Master/slave control. ResHB allows control of cooling output in “slave” rooms based on the cooling requirements of a “master” room, where the thermostat is located. Rooms with incompatible load profiles will exhibit poor temperature control. • Residential defaults. ResHB includes default values suitable for residential problems. In its current form, ResHB is a research-oriented reference implementation of RHB. ResHB FORTRAN source code is available under license from ASHRAE.
5.
RESIDENTIAL LOAD FACTOR (RLF) METHOD
The procedure presented in this chapter is the residential load factor (RLF) method. RLF is a simplified procedure derived from detailed ResHB analysis of prototypical buildings across a range of climates. The method is tractable by hand but is best applied using a spreadsheet. Two main applications are anticipated: • Education and training. The transparency and simplicity of RLF make it suitable for use in introductory courses on building load calculations. • Quick load estimates. In situations where detailed analysis is impractical, the RLF method is a possible alternative. For example, the method might be implemented as a spreadsheet on a handheld device and used for on-site sizing of replacement cooling equipment. Note that, although room-by-room calculations are possible with the RLF method, computerized methods based on RHB are more suitable for performing full room-level calculations required for equipment selection and distribution system design. RLF was derived from several thousand ResHB cooling load results (Barnaby and Spitler 2005; Barnaby et al. 2004). A range of climates and building types were analyzed. Statistical regression techniques were used to find values for the load factors tabulated in later sections. Factor values were validated by comparing ResHB versus RLF results for buildings not involved in the regression analysis. Within its range of applicability, RLF cooling loads are generally within 10% of those calculated with ResHB. The RLF derivation was repeated for 2009 using the updated temperature profile and clear-sky model (see Chapter 14), resulting in minor revisions to load factors and other coefficients. Additional revisions to Chapter 14 occurred in 2013 and 2017; those changes would alter RLF values very little, so the 2009 factors are retained. The RLF method should not be applied to situations outside the range of underlying cases, as shown in Table 1. Note that the RLF calculation sequence involves two distinct steps. First, the cooling and heating load factors (CFs and HFs) are derived for all project component types. These factors are then applied to the individual components by a single multiplication. (The two-step approach is demonstrated in the Load Calculation Example section.) For a specific location and representative constructions, CFs and HFs can be precalculated and used repeatedly. In essence, the structure of RLF allows assembling location-specific versions of the rigid tables found in prior editions, and also documents the equations used to generate tabulated values. Using these equations, a complete implementation of the RLF method, including CF and HF calculation, is well within the capabilities of current PC spreadsheet applications.
6.
COMMON DATA AND PROCEDURES
The following guidelines, data requirements, and procedures apply to all load calculation approaches, whether heating or cooling, hand-tractable or computerized.
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Residential Cooling and Heating Load Calculations Table 1 Item
Valid Range
Latitude
20 to 60°N
17.3
RLF Limitations
Notes
Also approximately valid for 20 to 60°S with N and S orientations reversed for southern hemisphere. Date July 21 Application must be summer peaking. Buildings in mild climates with significant SE/S/SW glazing may experience maximum cooling load in fall or even winter. Use RHB if local experience indicates this is a possibility. Elevation Less than 2000 m RLF factors assume 50 m elevation. With elevation-corrected Cs, method is acceptably accurate except at very high elevations. Climate Warm/hot Design-day average outdoor temperature assumed to be above indoor design temperature. Construction Lightweight residential construction (wood May be applied to masonry veneer over frame construction; results are conservative. Use or metal framing, wood or stucco siding) RHB for structural masonry or unconventional construction. Fenestration area 0 to 15% of floor area on any façade, 0 to Spaces with high fenestration fraction should be analyzed with RHB. 30% of floor area total Fenestration tilt Vertical or horizontal Skylights with tilt less than 30° can be treated as horizontal. Buildings with significant sloped glazing areas should be analyzed with RHB. Occupancy Residential Applications with high internal gains and/or high occupant density should be analyzed with RHB or nonresidential procedures. Temperature swing 1.7 K Distribution losses Typical Applications with extensive duct runs in unconditioned spaces should be analyzed with RHB.
qt = qs + ql
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General Guidelines Design for Typical Building Use. In general, residential systems should be designed to meet representative maximum-load conditions, not extreme conditions. Normal occupancy should be assumed, not the maximum that might occur during an occasional social function. Intermittently operated ventilation fans should be assumed to be off. These considerations are especially important for cooling-system sizing. Building Codes and Standards. This chapter presentation is necessarily general. Codes and regulations take precedence; consult local authorities to determine applicable requirements. Designer Judgment. Designer experience with local conditions, building practices, and prior projects should be considered when applying the procedures in this chapter. For equipment-replacement projects, occupant knowledge concerning performance of the existing system can often provide useful guidance for achieving a successful design. Verification. Postconstruction commissioning and verification are important steps in achieving design performance. Designers should encourage pressurization testing and other procedures that allow identification and repair of construction shortcomings. Uncertainty and Safety Allowances. Residential load calculations are inherently approximate. Many building characteristics are estimated during design and ultimately determined by construction quality and occupant behavior. These uncertainties apply to all calculation methods, including first-principles procedures such as RHB. It is therefore tempting to include safety allowances for each aspect of a calculation. However, this practice has a compounding effect and often produces oversized results. Typical conditions should be assumed; safety allowances, if applied at all, should be added to the final calculated loads rather than to intermediate components. In addition, temperature swing provides a built-in safety factor for sensible cooling: a 20% capacity shortfall typically results in a temperature excursion of at most about one or two degrees.
Basic Relationships Common air-conditioning processes involve transferring heat via air transport or leakage. The sensible, latent, and total heat conveyed by air on a volumetric basis is qs = Cs Qt
(1)
ql = Cl QW
(2)
qt = Ct Qh
(3)
(4)
where qs, ql, qt = Cs = Cl = Ct = Q t W h
= = = =
sensible, latent, total heat transfer rates, W air sensible heat factor, W/(L·s·K) (1.23 at sea level) air latent heat factor, W/(L·s) (3010 at sea level) air total heat factor, W/(L·s) per kJ/kg enthalpy h (1.2 at sea level) air volumetric flow rate, L/s air temperature difference across process, K air humidity ratio difference across process, kgw /kgda air enthalpy difference across process, kJ/kg
The heat factors Cs, Cl, and Ct are elevation dependent. The sealevel values in the preceding definitions are appropriate for elevations up to about 300 m. Procedures are provided in Chapter 18 for calculating adjusted values for higher elevations.
Design Conditions The initial step in the load calculation is selecting indoor and outdoor design conditions. Indoor Conditions. Indoor conditions assumed for design purposes depend on building use, type of occupancy, and/or code requirements. Chapter 9 and ASHRAE Standard 55 define the relationship between indoor conditions and comfort. Typical practice for cooling is to design for indoor conditions of 24°C db and a maximum of 50 to 65% rh. For heating, 20°C db and 30% rh are common design values. These conditions are the default values used throughout this chapter. Outdoor Conditions. Outdoor design conditions for load calculations should be selected from location-specific climate data in Chapter 14, or according to local code requirements as applicable. Cooling. The 1% design dry-bulb temperature and mean coincident wet bulb temperature from Chapter 14 climate data are generally appropriate. As previously emphasized, oversized cooling equipment results in poor system performance. Extremely hot events are necessarily of short duration (conditions always moderate each night); therefore, sacrificing comfort under typical conditions to meet occasional extremes is not recommended. Load calculations also require the hottest-month dry-bulb temperature daily range, and wind speed. These values can also be found in Chapter 14, although wind speed is commonly assumed to be 3.4 m/s. Typical buildings in middle latitudes generally experience maximum cooling requirements in midsummer (July in the northern hemisphere and January in the southern hemisphere). For this reason, the
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2017 ASHRAE Handbook—Fundamentals (SI)
RLF method is based on midsummer solar gains. However, this pattern does not always hold. Buildings at low latitudes or with significant south-facing glazing (north-facing in the southern hemisphere) should be analyzed at several times of the year using the RHB method. Local experience can provide guidance as to when maximum cooling is probable. For example, it is common for south-facing buildings in mild northern-hemisphere climates to have peak cooling loads in the fall because of low sun angles. Chapter 14 contains monthly temperature data to support calculations for any time of year. Heating. General practice is to use the 99% design dry-bulb temperature from Chapter 14. Heating load calculations ignore solar and internal gains, providing a built-in safety factor. However, the designer should consider two additional factors: • Many locations experience protracted (several-day) cold periods during which the outdoor temperature remains below the 99% value. • Wind is a major determinant of infiltration. Residences with significant leakage (e.g., older houses) may have peak heating demand under conditions other than extreme cold, depending on site wind patterns. Depending on the application and system type, the designer should consider using the 99.6% value or the mean minimum extreme as the heating design temperature. Alternatively, the heating load can be calculated at the 99% condition and a safety factor applied when equipment is selected. This additional capacity can also serve to meet pickup loads under nonextreme conditions. Adjacent Buffer Spaces. Residential buildings often include unconditioned buffer spaces such as garages, attics, crawlspaces, basements, or enclosed porches. Accurate load calculations require the adjacent air temperature. In many cases, a simple, conservative estimate is adequate, especially for heating calculations. For example, it is generally reasonable to assume that, under heating design conditions, adjacent uninsulated garages, porches, and attics are at outdoor temperature. Another reasonable assumption is that the temperature in an adjacent, unheated, insulated room is the mean of the indoor and outdoor temperatures. In cases where a temperature estimate is required, a steady-state heat balance analysis yields the following: C s Qt o + A x U x t x + q tb = ------------------------------------------------------Cs Q + Ax Ux
(5)
where buffer space temperature, °C buffer space infiltration/ventilation flow rate, L/s outdoor air temperature, °C area of xth buffer space surface, m2 U-factor of xth buffer space surface, W/(m2 ·K) air temperature at outside of xth buffer space surface, °C (typically, outdoor air temperature for exterior surfaces, conditioned space temperature for surfaces between buffer space and house, or ground temperature for below-grade surfaces) q = additional buffer space heat gains, W (e.g., solar gains or distribution system losses)
tb Q to Ax Ux tx
= = = = = =
Building Data Component Areas. To perform load calculations efficiently and reliably, standard methods must be used for determining building surface areas. For fenestration, the definition of component area must be consistent with associated ratings. Gross area. It is both efficient and conservative to derive gross surface areas from outer building dimensions, ignoring wall and floor thicknesses. Thus, floor areas should be measured to the outside of adjacent exterior walls or to the centerline of adjacent partitions. When apportioning to rooms, façade area should be divided
at partition centerlines. Wall height should be taken as floor-tofloor. Using outer dimensions avoids separate accounting of floor edge and wall corner conditions. Further, it is standard practice in residential construction to define floor area in terms of outer dimensions, so outer-dimension takeoffs yield areas that can be readily checked against building plans (e.g., the sum of room areas should equal the plan floor area). Although outer-dimension procedures are recommended as expedient for load calculations, they are not consistent with rigorous definitions used in building-related standards (e.g., ASTM Standard E631). However, the inconsistencies are not significant in the load calculation context. Fenestration area. Fenestration includes exterior windows, skylights, and doors. Fenestration U-factor and SHGC ratings (see Table 2) are based on the entire product area, including frames. Thus, for load calculations, fenestration area is the area of the rough opening in the wall or roof, less installation clearances (projected product area Apf ). Installation clearances can be neglected; it is acceptable to use the rough opening as an approximation of Apf . Net area. Net surface area is the gross surface area less fenestration area (rough opening or Apf ) contained within the surface. Volume. Building volume is expediently calculated by multiplying floor area by floor-to-floor height. This produces a conservative estimate of enclosed air volume, because wall and floor volumes are included in the total. More precise calculations are possible but are generally not justified in this context. Construction Characteristics. U-factors. Except for fenestration, construction U-factors should be calculated using procedures in Chapter 27, or taken from manufacturer’s data, if available. U-factors should be evaluated under heating (winter) conditions. Fenestration. Fenestration is characterized by U-factor and solar heat gain coefficient (SHGC), which apply to the entire assembly (including frames). If available, rated values should be used, determined according to procedures set forth by National Fenestration Rating Council (NFRC), Canadian Standards Association (CSA), or other specifying body (see Chapter 15). Ratings can be obtained from product literature, product label, or online listings (NFRC 2017). For unrated products (e.g., in existing construction), the Ufactor and SHGC can be estimated using Table 2 or tables in Chapter 15. Note that fenestration U-factors are evaluated under heating (winter) design conditions but are used in this chapter for both heating and cooling calculations. Relatively few types of glazing are encountered in residential applications. Single-glazed clear, double-glazed clear, and doubleglazed low-emissivity (“low-e”) glass predominate. Single-glazed is now rare in new construction but common in older homes. Tripleglazing, reflective glass, and heat-absorbing glass are encountered occasionally. Acrylic or glass skylights are common. Multipane low-e insulated glazing is available in high- and low-solar-gain variants, as discussed in Chapter 15. Low-solar is now the more common for new construction in all parts of the United States. Properties of windows equipped with storm windows should be estimated from data for a similar configuration with an additional pane. For example, data for clear, double-glazed should be used for a clear single-glazed window with a storm window. Fenestration interior and exterior shading must be included in cooling load calculations, as discussed in the Cooling Load section. Table 2 shows representative window U-factor and SHGC values for common glazing and frame combinations. Consult Chapter 15 for skylight characteristics.
Load Components Below-Grade Surfaces. For cooling calculations, heat flow into the ground is usually ignored because it is difficult to quantify.
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17.5
Table 2 Typical Fenestration Characteristicsa Frame
Aluminum with Thermal Break
Reinforced Vinyl/Aluminum Clad Wood
Wood/Vinyl
Insulated Fiberglass/Vinyl
Reflective
Aluminum
Heat-absorbing
Insulated Fiberglass/Vinyl
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Low-e, high-solar
Wood/Vinyl
Low-e, low-solar
Reinforced Vinyl/Aluminum Clad Wood
Clear
Aluminum with Thermal Break
Glazing Type
U SHGC U SHGC U SHGC
5.91 0.86 2.73 0.76 1.76 0.68
7.24 0.75 4.62 0.67 3.80 0.60
6.12 0.75 3.42 0.67 2.60 0.60
5.14 0.64 3.00 0.57 2.25 0.51
5.05 0.64 2.87 0.57 2.19 0.51
4.61 0.64 5.83 0.57 1.91 0.51
6.42 0.78 3.61 0.69 2.76 0.62
6.07 0.78 3.22 0.69 2.39 0.62
5.55 0.75 2.86 0.67 2.05 0.60
5.55 0.75 2.84 0.67 2.01 0.60
5.35 0.75 2.72 0.67 1.93 0.60
U SHGC U SHGC
1.70 0.41 1.02 0.27
3.83 0.37 3.22 0.25
2.68 0.37 2.07 0.25
2.33 0.31 1.76 0.21
2.21 0.31 1.71 0.21
1.89 0.31 1.45 0.21
2.75 0.38 2.13 0.25
2.36 0.38 1.76 0.25
2.03 0.36 1.44 0.24
2.01 0.36 1.40 0.24
1.90 0.36 1.33 0.24
U SHGC U SHGC
1.99 0.70 1.42 0.62
4.05 0.62 3.54 0.55
2.89 0.62 2.36 0.55
2.52 0.52 2.02 0.46
2.39 0.52 1.97 0.46
2.07 0.52 1.70 0.46
2.99 0.64 2.47 0.56
2.60 0.64 2.10 0.56
2.26 0.61 1.77 0.54
2.24 0.61 1.73 0.54
2.13 0.61 1.66 0.54
U SHGC U SHGC U SHGC
5.91 0.73 2.73 0.62 1.76 0.34
7.24 0.64 4.62 0.55 3.80 0.31
6.12 0.64 3.42 0.55 2.60 0.31
5.14 0.54 3.00 0.46 2.25 0.26
5.05 0.54 2.87 0.46 2.19 0.26
4.61 0.54 2.53 0.46 1.91 0.26
6.42 0.66 3.61 0.56 2.76 0.31
6.07 0.66 3.22 0.56 2.39 0.31
5.55 0.64 2.86 0.54 2.05 0.30
5.55 0.64 2.84 0.54 2.01 0.30
5.35 0.64 2.72 0.54 1.93 0.30
U SHGC U SHGC U SHGC
5.91 0.31 2.73 0.29 1.76 0.34
7.24 0.28 4.62 0.27 3.80 0.31
6.12 0.28 3.42 0.27 2.60 0.31
5.14 0.24 3.00 0.22 2.25 0.26
5.05 0.24 2.87 0.22 2.19 0.26
4.61 0.24 2.53 0.22 1.91 0.26
6.42 0.29 3.61 0.27 2.76 0.31
6.07 0.29 3.22 0.27 2.39 0.31
5.55 0.27 2.86 0.26 2.05 0.30
5.55 0.27 2.84 0.26 2.01 0.30
5.35 0.27 2.72 0.26 1.93 0.30
Center Glazing of Layers IDb Propertyc,d Glazing 1
1a
2
5a
3
29a
2
25a
3
40c
2
17c
3
32c
1
1c
2
5c
3
29c
1
1l
2
5p
3
29c
aData
are from Chapter 15, Tables 4 and 14 for selected combinations.
bID
= Chapter 15 glazing type identifier.
Surfaces adjacent to the ground are modeled as if well insulated on the outside, so there is no overall heat transfer, but diurnal heat storage effects are included. Heating calculations must include loss via slabs and basement walls and floors, as discussed in the Heating Load section. Infiltration. Infiltration is generally a significant component of both cooling and heating loads. Refer to Chapter 16 for a detailed discussion of residential air leakage. The simplified residential models found in that chapter can be used to calculate infiltration rates for load calculations. Infiltration should be evaluated for the entire building, not individual rooms or zones. Natural infiltration leakage rates are modified by mechanical pressurization caused by unbalanced ventilation or duct leakage. These effects are discussed in the section on Combined Ventilation and Infiltration Airflow. Leakage rate. Air leakage rates are specified either as airflow rate Qi , or air exchanges per hour (ACH), related as follows: Qi = ACH(V/3.6)
(6)
3.6Q i ACH = ------------V
(7)
where Qi = infiltration airflow rate, L/s ACH = air exchange rate, changes/h V = building volume, m3
Fixed
Aluminum
Operable
cU
= U-factor, W/(m2 ·K).
dSHGC
= solar heat gain coefficient.
Infiltration airflow rate depends on two factors: • Building effective leakage area (envelope leaks plus other air leakage paths, notably flues) and its distribution among ceilings, walls, floors, and flues. • Driving pressure caused by buoyancy (stack effect) and wind. Using the simplifying assumptions presented in Chapter 16, these factors can be evaluated separately and combined using Equation (8). Qi = AL IDF
(8)
where AL = building effective leakage area (including flue) at reference pressure difference = 4 Pa, assuming discharge coefficient CD = 1, cm2 IDF = infiltration driving force, L/(s·cm2)
The following sections provide procedures for determining AL and IDF. Leakage area. As discussed in Chapter 16, there are several interconvertible ways to characterize building leakage, depending on reference pressure differences and assumed discharge coefficient. This formulation uses the effective leakage area at 4 Pa, assuming CD = 1, designated AL (Sherman and Grimsrud 1980). The only accurate procedure for determining AL is by measurement using a pressurization test (commonly called a blower door
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2017 ASHRAE Handbook—Fundamentals (SI)
test). Numerous field studies have shown that visual inspection is not adequate for obtaining even a crude estimate of leakage. For buildings in design, a pressurization test is not possible and leakage area must be assumed for design purposes. Leakage can be estimated using tabulated component leakage areas found in Chapter 16. A simpler approach is based on an assumed average leakage per unit of building surface area: AL = Aes Aul
(9)
where
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Aes = building exposed surface area, m2 Aul = unit leakage area, cm2/m2 (from Table 3)
Aul is the leakage area per unit surface area; suitable design values are found in Table 3. Field experience indicates that the level of care applied to reducing leakage often depends on winter conditions, because cold-air leakage is readily detected. Thus, lower Aul values are expected in colder climates. Note that the Aul value doubles at each reduced construction quality step in Table 3; very high infiltration loads are typical in older houses. In Equation (9), Aes is the total building surface area at the envelope pressure boundary, defined as all above-grade surface area that separates the outdoors from conditioned or semiconditioned space. Table 4 provides guidance for evaluating Aes. IDF. To determine IDF, use the Chapter 16 methods cited previously. As a further simplification, Barnaby and Spitler (2005) derived the following relationship that yields results approximately equal to the AIM-2 model (Walker and Wilson 1990, 1998; Chapter 16’s enhanced model) at design conditions: I 0 + H t I 1 + I 2 A L , flue A L IDF = -------------------------------------------------------------------------------1000
(10)
where I0, I1, I2 = coefficients, as follows: Cooling 3.4 m/s
Heating 6.7 m/s
25 0.38 0.12
51 0.35 0.23
I0 I1 I2
H = building average stack height, m (typically 2.5 m per story)
Table 3
Unit Leakage Areas Aul , cm2/m2
Construction
Description
Tight
Construction supervised by air-sealing specialist Carefully sealed construction by knowledgeable builder Typical current production housing Typical pre-1970 houses Old houses in original condition
Good Average Leaky Very leaky
0.7 1.4 2.8 5.6 10.4
Table 4 Evaluation of Exposed Surface Area Situation
Include
Ceiling/roof combination (e.g., cathedral ceiling without attic) Ceiling or wall adjacent to attic Wall exposed to ambient
Gross surface area
Ceiling or wall area Gross wall area (including fenestration area) Wall adjacent to unconditioned buf- Common wall area fer space (e.g., garage or porch) Floor over open or vented Floor area crawlspace Floor over sealed crawlspace Crawlspace wall area Floor over conditioned or Above-grade basement semiconditioned basement wall area Slab floor
t = difference between indoor and outdoor temperatures, K AL, f lue = flue effective leakage area at reference pressure difference = 4 Pa, assuming CD = 1, cm2 (total for flues serving furnaces, domestic water heaters, fireplaces, or other vented equipment, evaluated assuming associated equipment is not operating and with dampers in closed position; see Chapter 16)
Building stack height H is the average height difference between the ceiling and floor (or grade, if the floor is below grade). Thus, for buildings with vented crawlspaces, the crawlspace height is not included. For basement or slab-on-grade construction, H is the average height of the ceiling above grade. Generally, there is significant leakage between basements and spaces above, so above-grade basement height should be included whether or not the basement is fully conditioned. With suitable adjustments for grade level, H can also be estimated as V/Acf (conditioned floor area). Equation (10) is valid for typical suburban residential wind sheltering, AL, f lue < AL /2, and at any elevation. Table 5 shows IDF values derived with Equation (10), assuming AL, flue = 0. Verification of leakage. A postconstruction pressurization test is strongly recommended to verify that design leakage assumptions are actually achieved. Excess leaks should be located and repaired. Allocation of infiltration to rooms. Total building infiltration should typically be allocated to rooms according to room volume; that is, it should be assumed that each room has the same air exchange rate as the whole building. In reality, leakage varies by room and over time, depending on outdoor temperature and wind conditions. These effects can either increase or decrease room leakage. In addition, system air mixing tends to redistribute localized leakage to all rooms. Thus, in most cases, there is no reasonable way to assign more or less leakage to specific rooms. An exception is leaky, multistory houses. The preferable and cost-effective response is mitigation of the leakage. If repair is not possible, then for heating load calculation purposes, some leakage can be differentially assigned to lower story and/or windward rooms in proportion to exposed surface area (i.e., adjustment using an “exposure factor”). Multifamily buildings. Usually, the simplified methods in Chapter 16 and this section do not apply to multifamily residences. However, they can be used for row houses that are full building height and have more than one exposed façade. For apartment units subdivided within a former detached residence, the entire building should be analyzed and the resulting exchange rate applied to the apartment volume. In other multifamily structures, infiltration is determined by many factors, including overall building height and degree of sealing between apartments. For low-rise construction, an upper bound for the infiltration rate can be found by evaluating the entire building. As building height increases, leakage problems can be magnified, as discussed in Chapter 16. Estimating leakage rates may require advice from a high-rise infiltration specialist. Ventilation. Whole-building ventilation. Because of energy efficiency concerns, residential construction has become significantly tighter over
Exclude
Table 5 Roof area Exterior wall area Crawlspace wall area Floor area Floor area Slab area
Typical IDF Values, L/(s·cm2)
Heating Design Temperature, °C
Cooling Design Temperature, °C
H, m
–40
2.5 3 4 5 6 7 8
0.10 0.095 0.086 0.077 0.069 0.11 0.10 0.093 0.083 0.072 0.14 0.12 0.11 0.093 0.079 0.16 0.14 0.12 0.10 0.086 0.18 0.16 0.14 0.11 0.093 0.20 0.17 0.15 0.12 0.10 0.22 0.19 0.16 0.14 0.11
–30
–20
–10
0
10
30
35
40
0.060 0.061 0.065 0.069 0.072 0.075 0.079
0.031 0.032 0.034 0.036 0.039 0.041 0.043
0.035 0.038 0.042 0.046 0.050 0.051 0.058
0.040 0.043 0.049 0.055 0.061 0.068 0.074
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17.7
the last several decades. Natural leakage rates are often insufficient to maintain acceptable indoor air quality. ASHRAE Standard 62.2 specifies the required minimum whole-building ventilation rate as
Note that unbalanced duct leakage can produce additional pressurization or depressurization. This effect is discussed in the section on Distribution Losses. Airflow components can be combined with infiltration leakage as follows (Palmiter and Bond 1991; Sherman 1992):
Qv = 0.15Acf + 3.5(Nbr + 1)
(11)
Qvi = max(Q unbal , Qi + 0.5Q unbal )
where where
Qv = required ventilation flow rate, L/s Acf = building conditioned floor area, m2 Nbr = number of bedrooms (not less than 1)
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(14)
Certain mild climates are exempted from this standard; local building authorities ultimately dictate actual requirements. In addition, Standard 62.2 specifies alternative methods for determining ventilation requirements that may result in smaller Qv values. Whole-building ventilation is expected to become more common because of a combination of regulation and consumer demand. The load effect of Qv must be included in both cooling and heating calculations. Heat recovery. Heat recovery devices should be considered part of mechanical ventilation systems. These appliances are variously called heat recovery ventilators (HRVs) or energy recovery ventilators (ERVs) and integrate with residential distribution systems, as described in Chapter 26 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. Either sensible heat or total heat (enthalpy) can be exchanged between the exhaust and intake airstreams. ERV/ HRV units are characterized by their sensible and total effectiveness. Local mechanical exhaust. Kitchen and bathroom exhaust fans are required by Standard 62.2 and are typically present. Exhaust fans that operate intermittently by manual control are generally not included in load calculations. Continuously operating ventilation should be included. Note that exhaust fans induce load only through enhanced infiltration because of building depressurization (see the section on Combined Ventilation and Infiltration Airflow for further discussion). Combustion Air. Fuel-fired boilers, furnaces, and domestic water heaters require combustion air. If the combustion air source is within the building envelope (including in semiconditioned basements), additional infiltration and heating load are induced. Locating the equipment outside of conditioned space (e.g., in a garage or vented mechanical closet) or using sealed-combustion equipment eliminates this load. Combustion air requirements for new forced-draft equipment can be estimated at 0.4 L/(s·kW) or about 12 L/s for a 30 kW heating appliance. The requirements for existing natural draft equipment should be estimated at twice that amount. In many cases, these quantities are relatively small and can be neglected. For cooling load calculations, heating equipment is assumed to be not operating, leaving only any domestic water heaters, for which the combustion air requirements are generally neglected. Combined Ventilation and Infiltration Airflow. Mechanical pressurization modifies the infiltration leakage rate. To assess this effect, overall supply and exhaust flow rates must be determined and then divided into “balanced” and “unbalanced” components. Qbal = min(Qsup, Qexh)
(12)
Qunbal = max(Qsup, Qexh) – Qbal
(13)
where Qbal = balanced airflow rate, L/s Qsup = total ventilation supply airflow rate, L/s Qexh = total ventilation exhaust airflow rate (including any combustion air requirements), L/s Qunbal = unbalanced airflow rate, L/s
Qvi = combined infiltration/ventilation flow rate (not including balanced component), L/s Qi = infiltration leakage rate assuming no mechanical pressurization, L/s
Ventilation/infiltration load. The cooling or heating load from ventilation and infiltration is calculated as follows: qvi,s = Cs[Qvi + (1 – s)Q bal, hr + Q bal, oth]t qvi, l = Cl (Qvi + Qbal,oth )W
(no HRV/ERV)
(15) (16)
qvi,t = Ct6 [Qvi + (1 – t )Q bal,hr + Qbal,oth]h
(17)
qvi,l = qvi,t – qvi,s
(18)
where qvi,s = sensible ventilation/infiltration load, W s = HRV/ERV sensible effectiveness Q bal,hr = balanced ventilation flow rate via HRV/ERV equipment, L/s Q bal,oth = other balanced ventilation supply airflow rate, L/s t = indoor/outdoor temperature difference, K W = indoor/outdoor humidity ratio difference qvi,t = total ventilation/infiltration load, W t = HRV/ERV total effectiveness h = indoor/outdoor enthalpy difference, kJ/kg qvi,l = latent ventilation/infiltration load, W
Distribution Losses. Air leakage and heat losses from duct systems frequently impose substantial equipment loads in excess of building requirements. The magnitude of losses depends on the location of duct runs, their surface areas, surrounding temperatures, duct wall insulation, and duct airtightness. These values are usually difficult to accurately determine at the time of preconstruction load calculations, and must be estimated using assumed values, so that selected equipment capacity is sufficient. Good design and workmanship both reduce duct losses. In particular, locating duct runs within the conditioned envelope (above dropped hallway ceilings, for example) substantially eliminates duct losses. Specific recommendations are found in Chapter 10 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. Good workmanship and correct materials are essential to achieve low leakage. Many common sealing techniques, notably duct tape, have been shown to fail in a few years. Well-constructed duct systems show leakage rates of 5% of fan flow from supply and return runs, whereas 11% or more on each side is more typical. Because of the potentially large load impact of duct leakage, postconstruction verification of airtightness is strongly recommended. Duct losses can be estimated using models specified in ASHRAE Standard 152, Francisco and Palmiter (1999), and Palmiter and Francisco (1997). The allowance for distribution losses is calculated as follows: qd = Fdl qbl
(19)
where qd = distribution loss, W Fdl = duct loss/gain factor, from Table 6 or ASHRAE Standard 152 design efficiencies or a detailed model qbl = total building load, W
Table 6 shows typical duct loss/gain factors calculated for the conditions indicated. These values can provide guidance for hand
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17.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 6 Typical Duct Loss/Gain Factors 1 Story Supply/Return Leakage Insulation (m2 ·K)/W
Duct Location
2 or More Stories
11%/11% R-0
R-0.7
5%/5% R-1.4
R-0
11%/11%
R-0.7
R-1.4
R-0
5%/5%
R-0.7
R-1.4
R-0
R-0.7
R-1.4
No loss (Fdl = 0)
Conditioned space Attic
C H/F H/HP
1.26 0.49 0.56
0.71 0.29 0.37
0.63 0.25 0.34
0.68 0.34 0.34
0.33 0.16 0.19
0.27 0.13 0.16
1.02 0.41 0.49
0.66 0.26 0.35
0.60 0.24 0.33
0.53 0.27 0.28
0.29 0.14 0.17
0.25 0.12 0.15
Basement
C H/F H/HP
0.12 0.28 0.23
0.09 0.18 0.17
0.09 0.16 0.16
0.07 0.19 0.14
0.05 0.10 0.09
0.04 0.08 0.08
0.11 0.24 0.20
0.09 0.17 0.16
0.09 0.15 0.15
0.06 0.16 0.12
0.04 0.09 0.08
0.04 0.08 0.07
Crawlspace
C H/F H/HP
0.16 0.49 0.56
0.12 0.29 0.37
0.11 0.25 0.34
0.10 0.34 0.34
0.06 0.16 0.19
0.05 0.13 0.16
0.14 0.41 0.49
0.12 0.26 0.35
0.11 0.24 0.33
0.08 0.27 0.28
0.06 0.14 0.17
0.05 0.12 0.15
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Values calculated for ASHRAE Standard 152 default duct system surface area using model of Francisco and Palmiter (1999). Values are provided as guidance only; losses can differ substantially for other conditions and configurations. Assumed surrounding temperatures: Heating/furnace (H/F) and heating/heating pump (H/HP): to = 0°C, tattic = 0°C, tb = 18°C, tcrawl = 0°C Cooling (C): to = 35°C, tattic = 49°C, tb = 20°C, tcrawl = 22°C
estimates, and illustrate the need for achieving low duct leakage. To the extent conditions differ from those shown, specific calculations should be made using a method cited previously. Note also that Table 6 cooling factors represent sensible gain only. Duct leakage may also introduce significant latent gain; see ASHRAE Standard 152.
7.
COOLING LOAD
A cooling load calculation determines total sensible cooling load from heat gain (1) through opaque surfaces (walls, floors, ceilings, and doors), (2) through transparent fenestration surfaces (windows, skylights, and glazed doors), (3) caused by infiltration and ventilation, and (4) because of occupancy. The latent portion of the cooling load is evaluated separately. Although the entire structure may be considered a single zone, equipment selection and system design should be based on room-by-room calculations. For proper design of the distribution system, the conditioned airflow required by each room must be known.
Peak Load Computation To select a properly sized cooling unit, the peak or maximum load (block load) for each zone must be computed. The block load for a single-family detached house with one central system is the sum of all the room loads. If the house has a separate system for each zone, each zone block load is required. When a house is zoned with one central cooling system, the system size is based on the entire house block load, whereas zone components, such as distribution ducts, are sized using zone block loads. In multifamily structures, each living unit has a zone load that equals the sum of the room loads. For apartments with separate systems, the block load for each unit establishes the system size. Apartment buildings having a central cooling system with fan-coils in each apartment require a block load calculation for the complete structure to size the central system; each unit load establishes the size of the fan-coil and air distribution system for each apartment. One of the methods for nonresidential buildings discussed in Chapter 18 may be used to calculate the block load.
Opaque Surfaces Heat gain through walls, floors, ceilings, and doors is caused by (1) the air temperature difference across such surfaces and (2) solar gains incident on the surfaces. The heat capacity of typical construction moderates and delays building heat gain. This effect is modeled in detail in the computerized RHB method, resulting in accurate simultaneous load estimates.
Table 7
Opaque Surface Cooling Factor Coefficients
Surface Type
OFt
OFb, K
OFr
Ceiling or wall adjacent to vented attic Ceiling/roof assembly
0.62 1
14.3roof – 4.5 38.3roof – 7.0
–0.19 –0.36
Wall (wood frame) or door with solar exposure Wall (wood frame) or door (shaded) Floor over ambient Floor over crawlspace Slab floor (see Slab Floor section)
1
8.2
–0.36
1 1 0.33
0 0 0
–0.36 –0.06 –0.28
roof = roof solar absorptance (see Table 8).
Table 8 Roof Solar Absorptance roof Color Material Asphalt shingles Tile Metal Elastomeric coating
White
Light
Medium
Dark
0.75 0.30 0.35 0.30
0.75 0.40 0.50
0.85 0.80 0.70
0.92 0.80 0.90
Source: Summarized from Parker et al. (2000).
The RLF method uses the following to estimate cooling load: (20) qopq = A CFopq CFopq = U(OFt t + OFb + OFr DR)
(21)
where qopq = A = CF = U = t = OFt ,OFb ,OFr = DR =
opaque surface cooling load, W net surface area, m2 surface cooling factor, W/m2 construction U-factor, W/(m2 ·K) cooling design temperature difference, K opaque-surface cooling factors (see Table 7) cooling daily range, K
OF factors, found in Table 7, represent construction-specific physical characteristics. OFt values less than 1 capture the buffering effect of attics and crawlspaces, OFb represents incident solar gain, and OFr captures heat storage effects by reducing the effective temperature difference. Note also that CF can be viewed as CF = U CLTD, the formulation used in prior residential and nonresidential methods. Table 7 factors for walls are simplified in two ways. First, the values do not depend on wall orientation. This has minimal effect on total load, because residences typically have a mix of exposures.
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Residential Cooling and Heating Load Calculations Second, only wood frame construction is included. The wood frame values can be used for heavier construction (e.g., masonry), but this overpredicts the wall’s contribution to cooling load and is thus conservative.
17.9 Table 9 Horizontal surfaces Et = 816.2 + 10.20L – 0.1806L2 Ed = min(Et , 170)
Slab Floors
ED = Et – Ed
Slab floors produce a slight reduction in cooling load, as follows: qopq = A CFslab
(22)
CFslab = 1.9 – 1.4 hsrf
(23)
Vertical surfaces = --------- (normalized exposure, 0 – 1) 180 Et = 423.5 + 1240 – 51433 + 31914 + 35.25L + 0.1427L2 – 12.89L – 0.6172L2 + [0. 7802L2/( + 1)] 115.9 4 L 2 Ed = min 356.3 – 37.73 + 0.3022L – --------------------- +1 ED = Et – Ed
where A = area of slab, m2 CFslab = slab cooling factor, W/m2 hsrf = effective surface conductance, including resistance of slab covering material such as carpet =1/(Rcvr + 0.12), W/(m2 ·K). Representative Rcvr values are found in Chapter 6 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. 1.9 = constant, W/m2 1.4 = factor, K
where Et , Ed , ED = peak hourly total, diffuse, and direct irradiance, W/m2 L = site latitude, °N = exposure (surface azimuth), ° from south (–180 to +180)
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Surfaces Adjacent to Buffer Space Heat gain from adjacent unconditioned or semiconditioned spaces can be calculated based on the partition U-factor and temperature difference. Buffer space air temperature tb can be estimated using procedures discussed in the section on Adjacent Buffer Spaces. Generally, simple approximations are sufficient. If the partition surface is large and/or poorly insulated, buffer temperature should be calculated with more care.
Transparent Fenestration Surfaces Cooling load associated with nondoor fenestration is calculated as follows: qfen = A CFfen
(24)
CFfen = U(t – 0.46DR) + PXI × SHGC × IAC × FFs
(25)
where fenestration cooling load, W fenestration area (including frame), m2 surface cooling factor, W/m2 fenestration NFRC heating U-factor, W/(m2 ·K) cooling design temperature difference, K peak exterior irradiance, including shading modifications, W/m2 [see Equations (26) or (27)] SHGC = fenestration rated or estimated NFRC solar heat gain coefficient IAC = interior shading attenuation coefficient, Equation (29) FFs = fenestration solar load factor, Table 13 qfen A CFfen U t PXI
Peak Irradiance Equations
= = = = = =
Peak Exterior Irradiance (PXI). Although solar gain occurs throughout the day, RP-1199 regression studies (Barnaby et al. 2004) showed that the cooling load contribution of fenestration correlates well with the peak-hour irradiance incident on the fenestration exterior. PXI is calculated as follows: PXI = Tx Et (unshaded fenestration)
(26)
PXI = Tx [Ed + (1 – Fshd)ED] (shaded fenestration)
(27)
Table 10 Peak Irradiance, W/m2 Latitude Exposure ED Ed Et
106 98 98 104 117 138 167 203 106 97 85 74 65 56 48 41 34 97 203 183 172 169 174 187 208 237 203
Northeast/Northwest ED Ed Et
442 442 444 447 451 457 465 474 442 149 139 131 124 117 111 106 100 149 590 582 575 571 568 568 570 574 590
East/West
ED Ed Et
524 548 570 590 608 624 638 651 524 178 171 164 159 154 149 145 141 178 702 719 734 748 761 773 783 792 702
Southeast/Southwest ED Ed Et
299 355 407 455 499 540 577 610 299 193 187 182 178 174 170 167 164 193 492 542 589 632 673 710 744 775 492
South
ED Ed Et
21 114 203 286 365 439 509 574 21 197 192 188 185 182 180 177 175 197 218 306 391 471 547 619 686 749 218
Horizontal
ED Ed Et
788 790 782 765 739 705 661 608 788 170 170 170 170 170 170 170 170 170 958 960 952 935 909 875 831 778 958
North
For horizontal or vertical surfaces, peak irradiance values can be obtained from Table 10 for primary exposures, or from Table 9 equations for any exposure. Skylights with slope less than 30° from horizontal should be treated as horizontal. Steeper, nonvertical slopes are not supported by the RLF method. Exterior Attachments. Common window coverings can significantly reduce fenestration solar gain. Table 11 shows transmission values for typical attachments. Permanent Shading. The shaded fraction Fshd can be taken as 1 for any fenestration shaded by adjacent structures during peak hours. Simple overhang shading can be estimated using the following: SLF D oh – X oh Fshd = min 1, max 0, ----------------------------------------- h
where PXI = peak exterior irradiance, W/m2 Et, Ed, ED = peak total, diffuse, and direct irradiance (Table 9 or 10), W/m2 Tx = transmission of exterior attachment (insect screen or shade screen) Fshd = fraction of fenestration shaded by permanent overhangs, fins, or environmental obstacles
20° 25° 30° 35° 40° 45° 50° 55° 60°
where SLF Doh Xoh h
= = = =
shade line factor from Table 12 depth of overhang (from plane of fenestration), m vertical distance from top of fenestration to overhang, m height of fenestration, m
(28)
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2017 ASHRAE Handbook—Fundamentals (SI)
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The shade line factor (SLF) is the ratio of the vertical distance a shadow falls beneath the edge of an overhang to the depth of the overhang, so the shade line equals the SLF times the overhang depth. Table 12 shows SLFs for July 21 averaged over the hours of greatest solar intensity on each exposure. More complex shading situations should be analyzed with the RHB method. Fenestration Solar Load Factors. Fenestration solar load factors FFs depend on fenestration exposure and are found in Table 13. The values represent the fraction of transmitted solar gain that contributes to peak cooling load. It is thus understandable that morning (east) values are lower than afternoon (west) values. Higher values are included for multifamily buildings with limited exposure. Interior Shading. Interior shading significantly reduces solar gain and is ubiquitous in residential buildings. Field studies show that a large fraction of windows feature some sort of shading; for example, James et al. (1997) studied 368 houses and found interior shading in 80% of audited windows. Therefore, in all but special circumstances, interior shading should be assumed when calculating cooling loads. In the RLF method, the interior attenuation coefficient (IAC) model is used, as described in Chapter 15. Residential values from that chapter are consolidated in Table 14. IAC values for
many other configurations are found in Chapter 15, Tables 14A to 14G. In some cases, it is reasonable to assume that a shade is partially open. For example, drapes are often partially open to admit daylight. IAC values are computed as follows: IAC = 1 + Fcl (IACcl – 1)
IAC = interior attenuation coefficient of fenestration with partially closed shade Fcl = shade fraction closed (0 to 1) IACcl = interior attenuation coefficient of fully closed configuration (from Table 14 or Chapter 15, Tables 14A to 14G)
Infiltration and Ventilation See the Common Data and Procedures section.
Internal Gain The contributions of occupants, lighting, and appliance gains to peak sensible and latent loads can be estimated as
Tx
None Exterior insect screen Shade screen
1.0 0.64 (see Chapter 15, Table 13G) Manufacturer shading coefficient (SC) value, typically 0.4 to 0.6
Table 12
qig,s qig,l Acf Noc
Latitude 20° 25° 30° 35° 40° 45° 50° 55° 60°
North Northeast/Northwest East/West Southeast/Southwest South
2.8 1.4 1.2 2.1 20.0
2.1 1.5 1.2 1.8 14.0
1.4 1.6 1.1 2.0 6.9
1.5 1.2 1.1 1.7 4.7
1.7 1.3 1.1 1.5 3.3
1.0 1.3 1.0 1.6 2.7
0.8 0.9 1.0 1.4 2.1
0.9 0.9 0.9 1.2 1.7
(30)
qig,l = 20 + 0.22Acf + 12Noc
(31)
= = = =
sensible cooling load from internal gains, W latent cooling load from internal gains, W conditioned floor area of building, m2 number of occupants (if unknown, estimate as Nbr + 1)
Equations (30) and (31) and their coefficients are derived from Building America (2004) load profiles evaluated at 4:00 PM, as documented by Barnaby and Spitler (2005). Predicted gains are typical for U.S. homes. Further allowances should be considered when unusual lighting intensities or other equipment are in continuous use during peak cooling hours. In critical situations where intermittent high occupant density or other internal gains are expected, a parallel cooling system should be considered. For room-by-room calculations, qig,s should be evaluated for the entire conditioned area, and allocated to kitchen and living spaces.
Shade Line Factors (SLFs)
Exposure
qig,s = 136 + 2.2Acf + 22Noc
where
Table 11 Exterior Attachment Transmission Attachment
(29)
where
0.8 0.8 0.8 1.1 1.4
Air Distribution System: Heat Gain
Note: Shadow length below overhang = SLF Doh.
See the Common Data and Procedures section.
Table 13 Fenestration Solar Load Factors FFs Exposure
Single Family Detached
Multifamily
North Northeast East Southeast South Southwest West Northwest Horizontal
0.44 0.21 0.31 0.37 0.47 0.58 0.56 0.46 0.58
0.27 0.43 0.56 0.54 0.53 0.61 0.65 0.57 0.73
Total Latent Load The latent cooling load is the result of three predominant moisture sources: outdoor air (infiltration and ventilation), occupants, and miscellaneous sources, such as cooking, laundry, and bathing. These components, discussed in previous sections, combine to yield the total latent load: ql = qvi,l + qig,l
(32)
where ql = total latent load, W
Table 14 Interior Attenuation Coefficients (IACcl) Drapes Glazing Layers
Open-Weave Glazing Type (ID*)
Roller Shades
Closed-Weave
Opaque
Blinds
Light
Dark
Light
Dark
White
Translucent Light
Medium
White
1
Clear (1a) Heat absorbing (1c)
0.64 0.68
0.71 0.72
0.45 0.50
0.64 0.67
0.34 0.40
0.44 0.49
0.74 0.76
0.66 0.69
2
Clear (5a) Low-e high-solar (17c) Low-e low-solar (25a) Heat absorbing (5c)
0.72 0.76 0.79 0.73
0.81 0.86 0.88 0.82
0.57 0.64 0.68 0.59
0.76 0.82 0.85 0.77
0.48 0.57 0.60 0.51
0.55 0.62 0.66 0.58
0.82 0.86 0.88 0.83
0.74 0.79 0.82 0.76
*Chapter
15 glazing identifier
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Residential Cooling and Heating Load Calculations
17.11
qvi,l = ventilation/infiltration latent gain, W, from Equation (16) or (18) qig,l = internal latent gain, W, from Equation (31)
well insulated, the adjacent buffer space procedure (see the section on Surfaces Adjacent to Buffer Space) can be used.
Additional latent gains may be introduced through return duct leakage and specific atypical sources. These may be estimated and included. Lstiburek and Carmody (1993) provide data for household moisture sources; however, again note that Equation (31) adequately accounts for normal gains. Because air conditioning systems are usually controlled by a thermostat, latent cooling is a side effect of equipment operation. During periods of significant latent gain but mild temperatures, there is little cooling operation, resulting in unacceptable indoor humidity. Multispeed equipment, combined temperature/humidity control, and dedicated dehumidification should be considered to address this condition.
Summary of RLF Cooling Load Equations Table 15 contains a brief list of equations used in the cooling load calculation procedure described in this chapter.
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8.
HEATING LOAD
Calculating a residential heating load involves estimating the maximum heat loss of each room or space to be heated and the simultaneous maximum (block) heat loss for the building, while maintaining a selected indoor air temperature during periods of design outdoor weather conditions. As discussed in the section on Calculation Approach, heating calculations use conservative assumptions, ignoring solar and internal gains, and building heat storage. This leaves a simple steady-state heat loss calculation, with the only significant difficulty being surfaces adjacent to grade.
Exterior Surfaces Above Grade All above-grade surfaces exposed to outdoor conditions (walls, doors, ceilings, fenestration, and raised floors) are treated identically, as follows: q = A × HF
(33)
HF = Ut
(34) W/m2.
where HF is the heating load factor in Two ceiling configurations are common:
• For ceiling/roof combinations (e.g., flat roof or cathedral ceiling), the U-factor should be evaluated for the entire assembly. • For well-insulated ceilings (or walls) adjacent to vented attic space, the U-factor should be that of the insulated assembly only (the roof is omitted) and the attic temperature assumed to equal the heating design outdoor temperature. The effect of attic radiant barriers can be neglected. In cases where the ceiling or wall is not
Below-Grade and On-Grade Surfaces The Heating Load Calculations section of Chapter 18 includes simplified procedures for estimating heat loss through below-grade walls and below- and on-grade floors. Those procedures are applicable to residential buildings. In more detailed work, Bahnfleth and Pedersen (1990) show a significant effect of the area-to-perimeter ratio. For additional generality and accuracy, see also methods described or cited in Beausoleil-Morrison and Mitalas (1997), CAN/ CSA Standard F280, HRAI (2014), and Krarti and Choi (1996).
Surfaces Adjacent to Buffer Space Heat loss to adjacent unconditioned or semiconditioned spaces can be calculated using a heating factor based on the partition temperature difference: HF = U(ti – tb)
Ventilation and Infiltration Infiltration of outdoor air causes both sensible and latent heat loss. The energy required to raise the temperature of outdoor infiltrating air to indoor air temperature is the sensible component; energy associated with net loss of moisture from the space is the latent component. Determining the volumetric flow Q of outdoor air
Table 15 Summary of RLF Cooling Load Equations Load Source
Equation
Exterior opaque surfaces
Partitions to unconditioned space Ventilation/infiltration Occupants and appliances Distribution Total sensible load
qopq = A CF CF = U(OFt t + OFb + OFr DR) qfen = A CF CF = U(t – 0.46DR) + PXI SHGC IAC FFs q = AUt qs = CsQt qig,s = 136 + 2.2Acf + 22Noc qd = Fdl q qs = qd + q
Latent load Ventilation/infiltration Internal gain
ql = qvi,l + qig,l qvi,l = Cl QW qig,l = 20 + 0.22Acf + 12Noc
Exterior transparent surfaces
(35)
Buffer space air temperature tb can be estimated using procedures discussed in the section on Adjacent Buffer Spaces. Generally, simple approximations are sufficient except where the partition surface is poorly insulated. Crawlspaces and basements are cases where the partition (the house floor) is often poorly insulated; they also involve heat transfer to the ground. Most codes require crawlspaces to be adequately vented year round. However, work highlighting problems with venting crawlspaces (DeWitt 2003) has led to application of sealed crawlspaces with insulated perimeter walls. Equation (5) may be applied to basements and crawlspace by including appropriate ground-related terms in the heat balance formulation. For example, when including below-grade walls, A x = Abw , Ux = Uavg,bw , and t x = tgr should be included as applicable in the summations in Equation (5). Losses from piping or ducting should be included as additional buffer space heat gain. Determining the ventilation or infiltration rate for crawlspaces and basements is difficult. Latta and Boileau (1969) estimated the air exchange rate for an uninsulated basement at 0.67 ach under winter conditions. Field measurements of eight ventilated crawlspaces summarized in Palmiter and Francisco (1996) yielded a median flow rate of 4.6 ach. Clearly, crawlspace infiltration rates vary widely, depending on vent configuration and operation.
Tables and Notes OF factors from Table 7 PXI from Table 9 or 10 plus adjustments FFs from Table 13 t = temperature difference across partition See Common Data and Procedures section Fdl from Table 6
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2017 ASHRAE Handbook—Fundamentals (SI)
entering the building is discussed in the Common Data and Procedures section and in Chapter 16. Determining the resulting sensible and latent loads is discussed in the Ventilation/Infiltration Load subsection.
Humidification In many climates, humidification is required to maintain comfortable indoor relative humidity under heating conditions. The latent ventilation and infiltration load calculated, assuming desired indoor humidity conditions, equals the sensible heat needed to evaporate water at a rate sufficient to balance moisture losses from air leakage. Self-contained humidifiers provide this heat from internal sources. If the heat of evaporation is taken from occupied space or the distribution system, the heating capacity should be increased accordingly.
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Pickup Load For intermittently heated buildings and night thermostat setback, additional heat is required to raise the temperature of air, building materials, and material contents to the specified temperature. The rate at which this additional heat must be supplied is the pickup load, which depends on the structure’s heat capacity, its material contents, and the time in which these are to be heated. Because the design outdoor temperature is generally much lower than typical winter temperatures, under most conditions excess heating capacity is available for pickup. Therefore, many engineers make no pickup allowance except for demanding situations. If pickup capacity is justified, the following guidance can be used to estimate the requirement. Relatively little rigorous information on pickup load exists. Building simulation programs can predict recovery times and required equipment capacities, but a detailed simulation study is rarely practical. Armstrong et al. (1992a, 1992b) developed a model for predicting recovery from setback and validated it for a church and two office buildings. Nelson and MacArthur (1978) studied the relationship between thermostat setback, furnace capacity, and recovery time. Hedrick et al. (1992) compared Nelson and
MacArthur’s results to tests for two test houses and found that the furnace oversizing required for a 2 h recovery time ranges from 20 to 120%, depending on size of setback, building mass, and heating t (colder locations require less oversizing on a percentage basis). The designer should be aware that there are trade-offs between energy savings from thermostat setback and energy penalties incurred by oversizing equipment. Koenig (1978) studied a range of locations and suggested that 30% oversizing allows recovery times less than 4 h for nearly the entire heating season and is close to optimum from an energy standpoint. The preceding guidance applies to residential buildings with fuel-fired furnaces. Additional considerations may be important for other types of heating systems. For air-source heat pumps with electric resistance auxiliary heat, thermostat setback may be undesirable (Bullock 1978). Thermostats with optimum-start algorithms, designed to allow both energy savings and timely recovery to the daytime set point, are becoming routinely available and should be considered in all cases.
Summary of Heating Load Procedures Table 16 lists equations used in the heating load calculation procedures described in this chapter.
9.
LOAD CALCULATION EXAMPLE
A single-family detached house with floor plan shown in Figure 1 is located in Atlanta, GA, USA. Construction characteristics are documented in Table 17. Using the RLF method, find the block (whole-house) design cooling and heating loads. A furnace/airconditioner forced-air system is planned with a well-sealed and well-insulated (R-8 wrap) attic duct system.
Solution Design Conditions. Table 18 summarizes design conditions. Typical indoor conditions are assumed. Outdoor conditions are determined from Chapter 14.
Table 16 Summary of Heating Load Calculation Equations Load Source
Equation
Tables and Notes
Exterior surfaces above grade Partitions to unconditioned buffer space Walls below grade Floors on grade Floors below grade Ventilation/infiltration Total sensible load
q = UAt q = UAt q =Uavg,bw A(tin – tgr) q = Fp pt q =Uavg,bf A(tin – tgr) qvi = Cs Qt qs = q
t = ti – to t = temp. difference across partition
Table 17 Component Roof/ceiling
Example House Characteristics
Description
Flat wood frame ceiling (insulated with R-5.3 fiberglass) beneath vented attic with medium asphalt shingle roof Exterior walls Wood frame, exterior wood sheathing, interior gypsum board, R-2.3 fiberglass insulation Doors Wood, solid core Floor Slab on grade with heavy carpet over rubber pad; R-0.9 edge insulation to 1 m below grade Windows
Construction
See Chapter 18, Equations (41) and (42) See Chapter 18, Equations (37) and (38) From Common Data and Procedures section
Factors U = 0.18 W/(m2 ·K) roof = 0.85 (Table 8) U = 51 W/(m2 ·K)
U = 2.3 W/(m2 ·K) Rcvr = 0.21 (m2 ·K)/W (Table 3, Chapter 6, 2016 ASHRAE Handbook—HVAC Systems and Equipment) Fp = 85 W/(m2 ·K) (estimated from Chapter 18, Table 24) Low-e/low-solar in wood frames. Half fixed, half operable with insect screens Fixed: U = 2.01 W/(m2 ·K); SHGC = 0.36 (Table 2) (except living room picture window, which is fixed). 0.6 m eave overhang on Operable: U = 2.21 W/(m2 ·K); SHGC = 0.31 (Table 2); east and west with eave edge at same height as top of glazing for all windows. Tx = 0.64 (Table 11) Allow for typical closed-weave light drape interior shading, half closed. IACcl = 0.68 (Table 14) Good Aul = 1.4 cm2/m2 (Table 3)
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Residential Cooling and Heating Load Calculations Table 18
17.13
Example House Design Conditions
Item
Heating
Cooling
Notes
Latitude Elevation Indoor temperature Indoor relative humidity Outdoor temperature
— — 20°C N/A –3.5°C
— — 24°C 50% 33°C
33.64°N 315 m
Daily range Outdoor wet bulb Wind speed Design t Moisture difference
N/A N/A 6.7 m/s 23.5 K
9.3°K 23.2°C 3.4 m/s 9K 0.0050 kg/kg
No humidification Cooling: 1% value Heating: 99% MCWB* at 1% Default assumption Psychrometric chart
*MCWB = mean coincident wet bulb.
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Table 19
Fig. 1 Example House Component Quantities. Areas and lengths required for load calculations are derived from plan dimensions (Figure 1). Table 19 summarizes these quantities. Opaque Surface Factors. Heating and cooling factors are derived for each component condition. Table 20 shows the resulting factors and their sources. Window Factors. Deriving cooling factors for windows requires identifying all unique glazing configurations in the house. Equation (25) input items indicate that the variations for this case are exposure, window height (with overhang shading), and frame type (which determines U-factor, SHGC, and the presence of insect screen). CF derivation for all configurations is summarized in Table 21. For example, CF for operable 1 m high windows facing west (the second row in Table 21) is derived as follows: • U-factor and SHGC are found in Table 2. • Each operable window is equipped with an insect screen. From Table 11, Tx = 0.64 for this arrangement. • Overhang shading is evaluated with Equation (28). For west exposure and latitude 34°, Table 12 shows SLF = 1.1. Overhang depth (Doh) is 0.6 m and the window-overhang distance (Xoh) is 0 m. With window height h of 0.9 m, Fs = 0.73 (73% shaded). • PXI depends on peak irradiance and shading. Approximating site latitude as 35°N, Table 10 shows ED = 570 and Ed = 164 W/m2 for west exposure. Equation (27) combines these values with Tx and Fs to find PXI = 0.64[164 + (1 – 0.73)570] = 203 W/m2. • All windows are assumed to have some sort of interior shading in the half-closed position. Use Equation (29) with Fcl = 0.5 and IACcl = 0.68 (per Table 17) to derive IAC = 0.84. • FFs is taken from Table 13 for west exposure. • Finally, inserting the preceding values into Equation (25) gives CF = 2.21(9.1 – 0.46 × 9.3) + 203 × 0.31 × 0.84 × 0.56 = 40.3 W/m2.
Example House Component Quantities
Component
Quantity
Notes
Ceiling
195.3 m2
Overall area less garage area (22.6 × 11) – (7.3 × 7.3) 2 (each 0.9 by 2.1 m)
Doors
3.8 m2
Windows
13.9 m2
Walls, exposed exterior
126.2 m2 gross, Wall height = 2.4 m 108.5 m2 net
Walls, garage
35.0 m2
Floor area
195.3 m2
Floor perimeter
67.2 m
Include perimeter adjacent to garage
Total exposed surface
67.2 m2
Wall gross area (including garage wall) plus ceiling area
Volume
356.5 m3
Envelope Loads. Given the load factors and component quantities, heating and cooling loads are calculated for each envelope element, as shown in Table 22. Infiltration and Ventilation. From Table 3, Aul for this house is 1.4 cm2/m2 of exposed surface area. Applying Equation (9) yields AL = Aes Aul = 356.5 × 1.1 = 499 cm2. Using Table 5, estimate heating and cooling IDF to be 0.073 and 0.035 (L·s)/cm2, respectively [alternatively, Equation (10) could be used to find IDF values]. Apply Equation (8) to find the infiltration leakage rates and Equation (7) to convert the rate to air changes per hour: Qi,h = 499 × 0.073 = 36 L/s (0.28 ach) Qi,c = 499 × 0.035 = 17 L/s (0.13 ach) Calculate the ventilation outdoor air requirement with Equation (11) using Acf = 195.3 m2 and Nbr = 3, resulting in Qv = 43 L/s. For design purposes, assume that this requirement is met by a mechanical system with balanced supply and exhaust flow rates (Qunbal = 0). Find the combined infiltration/ventilation flow rates by summing the balanced ventilation flow with net infiltration flow derived with Equation (14): Qvi,h = 43 + max(0, 36 + 0.5 0) = 79 L/s Qvi,c = 43 + max(0, 17 + 0.5 0) = 60 L/s At Atlanta’s elevation of 313 m, elevation adjustment of heat factors results in a small (4%) reduction in air heat transfer; thus, adjustment is unnecessary, resulting in Cs = 23 W/(L·s·K). Use Equation (15) with Qbal,hr = 0 and Qbal,oth = 0 to calculate the sensible infiltration/ventilation loads: qvi,s,h = 1.23 × 79 × 23 = 2235 W qvi,s,c = 1.23 × 60 × 9.1 = 672 W
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17.14
2017 ASHRAE Handbook—Fundamentals (SI) Table 20 Example House Opaque Surface Factors Heating
Cooling
Component
U, W/(m2 ·K) or Fp, W/(m·K)
HF
Ceiling Wall Garage wall Door
0.18 0.51 0.51 2.3
4.14 11.73 11.73 52.90
Floor perimeter
0.85
19.55 Chapter 18, Equation (42)
Reference
OFt
OFb
OFr
Equation (34)
0.62 1 1 1
7.66 8.20 0.00 8.20
–0.19 –0.36 –0.36 –0.36
1.9
–1.4/(0.21 + 0.12) = –4.24
Floor area
CF
Reference
2.08 Table 7 Equation (21) 7.09 2.93 32.09 –2.34 Equation (23)
Table 21 Example House Window Factors
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Exposure
U,
Height, m
Frame
W/(m2 ·K)
HF
Tx
Fshd
PXI
SHGC
IAC
FFs
CF
Table 2
Eq. (34)
Table 11
Eq. (28)
Eq. (27)
Table 2
Eq. (29)
Table 13
Eq. (25)
West
0.9 0.9 1.8 1.8 2.4
Fixed Operable Fixed Operable Fixed
2.01 2.21 2.01 2.21 2.01
46.2 50.8 46.2 50.8 46.2
1 0.64 1 0.64 1
0.73 0.73 0.37 0.37 0.28
318 203 523 335 574
0.36 0.31 0.36 0.31 0.36
0.84 0.84 0.84 0.84 0.84
0.56 0.56 0.56 0.56 0.56
63.5 40.3 98.3 59.5 107.0
South
1.2 1.2
Fixed Operable
2.01 2.21
46.2 50.8
1 0.64
0.00 0.00
391 250
0.36 0.31
0.84 0.84
0.47 0.47
65.3 41.3
East
0.9 0.9 1.2 1.2
Fixed Operable Fixed Operable
2.01 2.21 2.01 2.21
46.2 50.8 46.2 50.8
1 0.64 1 0.64
0.73 0.73 0.55 0.55
318 203 421 269
0.36 0.31 0.36 0.31
0.84 0.84 0.84 0.84
0.31 0.31 0.31 0.31
39.5 27.1 49.1 32.4
Table 22 Example House Envelope Loads Component Ceiling Wall Garage wall Door Floor perimeter Floor area W-Fixed-0.9 W-Operable-0.9 W-Fixed-1.8 W-Operable-1.8 W-Fixed-2.4 S-Fixed-1.2 S-Operable-1.2 E-Fixed-0.9 E-Operable-0.9 E-Fixed-1.2 E-Operable-1.2
HF
CF
4.14 2.08 11.73 7.12 11.73 2.93 52.9 32.09 19.6 –2.34 46.2 63.5 50.8 40.3 46.2 98.3 50.8 59.5 46.2 107.0 46.2 65.3 50.8 41.3 46.2 39.5 50.8 27.1 46.2 49.1 50.8 32.4
Quantity, m2 or m
Heating Load, W
Cooling Load, W
195.3 108.5 35 3.8 67.2 195.3 0.4 0.4 1.1 1.1 4.3 0.7 0.7 0.4 0.4 2.2 2.2
809 1273 411 201 1317 0 18 20 51 56 199 32 36 18 20 102 112
405 772 103 122 0 –457 25 16 108 65 460 46 29 16 11 108 71
4674
1900
Envelope totals
Table 23
Example House Total Sensible Loads
Item
Heating Load, W
Cooling Load, W
Envelope Infiltration/ventilation Internal gain
4674 2235
1900 672 654
Subtotal Distribution loss
6909 898
3225 871
Total sensible load
7807
4096
Internal Gain. Apply Equation (30) to find the sensible cooling load from internal gain: qig,s = 136 + 2.2 × 195.3 + 22(3 + 1) = 654 W Distribution Losses and Total Sensible Load. Table 23 summarizes the sensible load components. Distribution loss factors Fdl are estimated (from Table 6) at 0.13 for heating and 0.27 for cooling. Latent Load. Use Equation (16) with Cl = 3010 W/(L·s), Qvi,c = 60 L/s, Qbal,oth = 0, and W = 0.0050 to calculate the infiltration/ ventilation latent load = 903 W. Use Equation (31) to find the latent load from internal gains = 111 W. Therefore, the total latent cooling load is 752 W.
10.
SYMBOLS
A = area, m2; ground surface temperature amplitude, °C AL = building effective leakage area (including flue) at 4 Pa, assuming CD = 1, cm2 Cl = air latent heat factor, 3010 W/(L·s) at sea level Cs = air sensible heat factor, 23 W/(L·s·K) at sea level Ct = air total heat factor, 1.2 W/(L·s) (kJ/kg) at sea level CF = cooling load factor, W/m2 Doh = depth of overhang (from plane of fenestration), m DR = daily range of outdoor dry-bulb temperature, K E = peak irradiance for exposure, W/m2 Fdl = distribution loss factor Fp = heat loss coefficient per unit length of perimeter, W/(m·K) Fshd = shaded fraction FF = coefficient for CFfen G = internal gain coefficient hsrf = effective surface conductance, including resistance of slab covering material such as carpet, 1/(Rcvr + 0.12) W/(m2 ·K) h = indoor/outdoor enthalpy difference, kJ/kg H = height, m HF = heating (load) factor, W/m2 I = infiltration coefficient IAC = interior shading attenuation coefficient IDF = infiltration driving force, L/(s·cm2)
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Residential Cooling and Heating Load Calculations k LF OF p PXI q Q R SHGC SLF t Tx t U w W V Xoh z roof
= = = = =
conductivity, W/(m·K) load factor, W/m2 coefficient for CFopq perimeter or exposed edge of floor, m peak exterior irradiance, including shading modifications, W/m2 = heating or cooling load, W = air volumetric flow rate, L/s = insulation thermal resistance, (m2 ·K)/W = fenestration rated or estimated NFRC solar heat gain coefficient = shade line factor = temperature, °C = solar transmission of exterior attachment = design dry-bulb temperature difference (cooling or heating), K = construction U-factor, W/(m2 ·K) (for fenestration, NFRC rated heating U-factor) = width, m = indoor-outdoor humidity ratio difference, kgw /kgda = building volume, m3 = vertical distance from top of fenestration to overhang, m = depth below grade, m = roof solar absorptance = heat/energy recovery ventilation (HRV/ERV) effectiveness
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Subscripts avg b bal bf bl bw br ceil cf cl cvr d D da dl env es exh fen floor gr hr i in ig l o oc oh opq oth pf r rhb s shd slab srf sup t ul unbal v vi w wall x
= = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = = =
average base (as in OFb), basement, building, buffer balanced basement floor building load basement wall bedrooms ceiling conditioned floor closed floor covering diffuse, distribution direct dry air distribution loss envelope exposed surface exhaust fenestration floor ground heat recovery infiltration indoor internal gain latent outdoor occupant overhang opaque other projected product daily range (as in OFr) calculated with RHB method sensible or solar shaded slab surface supply total or temperature (as in OFt) unit leakage unbalanced ventilation ventilation/infiltration water wall xth buffer space surface
17.15 REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ACCA. 2016. Manual J residential load calculations, 8th ed., v. 2.5. Air Conditioning Contractors of America, Arlington, VA. Armstrong, P.R., C.E. Hancock, and J.R. Seem. 1992a. Commercial building temperature recovery—Part 1: Design procedure based on a step response model (RP-491). ASHRAE Transactions 98(1):381-396. Armstrong, P.R., C.E. Hancock, and J.R. Seem. 1992b. Commercial building temperature recovery—Part 2: Experiments to verify step response model (RP-491). ASHRAE Transactions 98(1):397-410. ASHRAE. 2013. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2013. ASHRAE. 2016. Ventilation and acceptable indoor air quality in low-rise residential buildings. ANSI/ASHRAE Standard 62.2-2016. ASHRAE. 2014. Method of test for determining the design and seasonal efficiencies of residential thermal distribution systems. ANSI/ASHRAE Standard 152-2014. ASTM. 1998. Standard terminology of building constructions. Standard E631-93a (1998)e1. American Society for Testing and Materials, West Conshohocken, PA. Bahnfleth, W.P., and C.O. Pedersen 1990. A three-dimensional numerical study of slab-on-grade heat transfer. ASHRAE Transactions 96(2):61-72. Barnaby, C.S., and J.D. Spitler. 2005. Development of the residential load factor method for heating and cooling load calculations. ASHRAE Transactions 111(1):291-307. Barnaby, C.S., J.D. Spitler, and D. Xiao. 2004. Updating the ASHRAE/ ACCA residential heating and cooling load calculation procedures and data (RP-1199). ASHRAE Research Project, Final Report. Barnaby, C.S., J.D. Spitler, and D. Xiao. 2005. The residential heat balance method for heating and cooling load calculations (RP-1199). ASHRAE Transactions 111(1):308-319. Beausoleil-Morrison, I., and G. Mitalas. 1997. BASESIMP: A residentialfoundation heat-loss algorithm for incorporating into whole-building energy-analysis programs. Proceedings of Building Simulation ’97, Prague. Building America. 2004. Building America research benchmark definition, v. 3.1. Available at www.nrel.gov/docs/fy05osti/36429.pdf. Bullock, C.E. 1978. Energy savings through thermostat setback with residential heat pumps. ASHRAE Transactions 84(2):352-363. CSA Group. 2012. Determining the required capacity of residential space heating and cooling appliances. CAN/CSA Standard F280-12. CSA Group, Mississauga, ON, Canada. DeWitt, C. 2003. Crawlspace myths. ASHRAE Journal 45:20-26. Francisco, P.W., and L. Palmiter. 1999 (rev. 2003). Improvements to ASHRAE Standard 152P. Ecotope, Inc., Seattle, WA. Hedrick, R.L., M.J. Witte, N.P. Leslie, and W.W. Bassett. 1992. Furnace sizing criteria for energy-efficient setback strategies. ASHRAE Transactions 98(1):1239-1246. HRAI. 2014. Residential heat loss and gain calculations: Student reference guide. Heating, Refrigerating and Air Conditioning Institute of Canada. Mississauga, ON. James, P., J. Cummings, J. Sonne, R. Vieira, and J. Klongerbo. 1997. The effect of residential equipment capacity on energy use, demand, and runtime. ASHRAE Transactions 103(2):297-303. Koenig, K. 1978. Gas furnace sizing requirements for residential heating using thermostat night setback. ASHRAE Transactions 84(2):335-351. Krarti, M., and S. Choi 1996. Simplified method for foundation heat loss calculation. ASHRAE Transactions 102(1):140-152. Latta, J.K., and G.G. Boileau. 1969. Heat losses from house basements. Canadian Building 19(10):39. Lstiburek, J.L., and J. Carmody. 1993. Moisture control handbook. Van Nostrand Reinhold, New York. McQuiston, F.C. 1984. A study and review of existing data to develop a standard methodology for residential heating and cooling load calculations (RP-342). ASHRAE Transactions 90(2A):102-136. Nelson, L.W., and J.W. MacArthur. 1978. Energy savings through thermostat setback. ASHRAE Transactions 84(2):319-334. NFRC. 2017. NFRC certified products directory. National Fenestration Rating Council, Greenbelt, MD. www.nfrc.org or search.nfrc.org.
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17.16
2017 ASHRAE Handbook—Fundamentals (SI)
Palmiter, L., and T. Bond. 1991. Interaction of mechanical systems and natural infiltration. Proceedings of the 12th AIVC Conference on Air Movement and Ventilation Control Within Buildings. Air Infiltration and Ventilation Centre, Coventry, U.K. Palmiter, L., and P. Francisco. 1996. Modeled and measured infiltration: Phase III. A detailed case study of three homes. Electric Power Research Institute Report TR-106288. Palo Alto, CA. Palmiter, L., and P. Francisco. 1997. Development of a practical method of estimating the thermal efficiency of residential forced-air distribution systems. Electric Power Research Institute Report TR-107744. Palo Alto, CA.
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Parker, D.S., J.E.R. McIlvaine, S.F. Barkaszi, D.J. Beal, and M.T. Anello. 2000. Laboratory testing of the reflectance properties of roofing materials. FSEC-CR670-00. Florida Solar Energy Center, Cocoa.
Pedersen, C.O., D.E. Fisher, J.D. Spitler, and R.J. Liesen. 1998. Cooling and heating load calculation principles. ASHRAE. Pedersen, C.O., R.J. Liesen, R.K. Strand, D.E. Fisher, L. Dong, and P.G. Ellis. 2001. Toolkit for building load calculations. ASHRAE. Sherman, M.H. 1992. Superposition in infiltration modeling. Indoor Air 2: 101-114. Sherman, M.H., and D.T. Grimsrud. 1980. Infiltration-pressurization correlation: Simplified physical modeling. ASHRAE Transactions 86(2):778. Walker, I.S., and D.J. Wilson. 1990. The Alberta air infiltration model. The University of Alberta, Department of Mechanical Engineering, Technical Report 71. Walker, I.S., and D.J. Wilson. 1998. Field validation of equations for stack and wind driven air infiltration calculations. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 4(2).
Related Commercial Resources
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Related Commercial Resources CHAPTER 18
NONRESIDENTIAL COOLING AND HEATING LOAD CALCULATIONS Cooling Load Calculation Principles ...................................... 18.1 Internal Heat Gains ................................................................. 18.3 Infiltration and Moisture Migration Heat Gains ......................................................................... 18.14 Fenestration Heat Gain ......................................................... 18.16 Heat Balance Method ............................................................ 18.16
Radiant Time Series (RTS) Method........................................ Heating Load Calculations .................................................... System Heating and Cooling Load Effects............................. Example Cooling and Heating Load Calculations ................ Previous Cooling Load Calculation Methods........................ Building Example Drawings ..................................................
EATING and cooling load calculations are the primary design basis for most heating and air-conditioning systems and components. These calculations affect the size of piping, ductwork, diffusers, air handlers, boilers, chillers, coils, compressors, fans, and every other component of systems that condition indoor environments. Cooling and heating load calculations can significantly affect first cost of building construction, comfort and productivity of occupants, and operating cost and energy consumption. Simply put, heating and cooling loads are the rates of energy input (heating) or removal (cooling) required to maintain an indoor environment at a desired temperature and humidity condition. Heating and air conditioning systems are designed, sized, and controlled to accomplish that energy transfer. The amount of heating or cooling required at any particular time varies widely, depending on external (e.g., outdoor temperature) and internal (e.g., number of people occupying a space) factors. Peak design heating and cooling load calculations, which are this chapter’s focus, seek to determine the maximum rate of heating and cooling energy transfer needed at any point in time. Similar principles, but with different assumptions, data, and application, can be used to estimate building energy consumption, as described in Chapter 19. This chapter discusses common elements of cooling load calculation (e.g., internal heat gain, ventilation and infiltration, moisture migration, fenestration heat gain) and two methods of heating and cooling load estimation: heat balance (HB) and radiant time series (RTS).
Many cooling load components vary widely in magnitude, and possibly direction, during a 24 h period. Because these cyclic changes in load components often are not in phase with each other, each component must be analyzed to establish the maximum cooling load for a building or zone. A zoned system (i.e., one serving several independent areas, each with its own temperature control) needs to provide no greater total cooling load capacity than the largest hourly sum of simultaneous zone loads throughout a design day; however, it must handle the peak cooling load for each zone at its individual peak hour. At some times of day during heating or intermediate seasons, some zones may require heating while others require cooling. The zones’ ventilation, humidification, or dehumidification needs must also be considered.
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H
1. COOLING LOAD CALCULATION PRINCIPLES Cooling loads result from many conduction, convection, and radiation heat transfer processes through the building envelope and from internal sources and system components. Building components or contents that may affect cooling loads include the following: • External: Walls, roofs, windows, skylights, doors, partitions, ceilings, and floors • Internal: Lights, people, appliances, and equipment • Infiltration: Air leakage and moisture migration • System: Outdoor air, duct leakage and heat gain, reheat, fan and pump energy, and energy recovery
1.1
TERMINOLOGY
The variables affecting cooling load calculations are numerous, often difficult to define precisely, and always intricately interrelated. The preparation of this chapter is assigned to TC 4.1, Load Calculation Data and Procedures.
18.1 Copyright © 2017, ASHRAE
18.22 18.30 18.41 18.44 18.57 18.61
Heat Flow Rates In air-conditioning design, the following four related heat flow rates, each of which varies with time, must be differentiated. Space Heat Gain. This instantaneous rate of heat gain is the rate at which heat enters into and/or is generated within a space. Heat gain is classified by its mode of entry into the space and whether it is sensible or latent. Entry modes include (1) solar radiation through transparent surfaces; (2) heat conduction through exterior walls and roofs; (3) heat conduction through ceilings, floors, and interior partitions; (4) heat generated in the space by occupants, lights, and appliances; (5) energy transfer through direct-with-space ventilation and infiltration of outdoor air; and (6) miscellaneous heat gains. Sensible heat is added directly to the conditioned space by conduction, convection, and/or radiation. Latent heat gain occurs when moisture is added to the space (e.g., from vapor emitted by occupants and equipment). To maintain a constant humidity ratio, water vapor must condense on the cooling apparatus and be removed at the same rate it is added to the space. The amount of energy required to offset latent heat gain essentially equals the product of the condensation rate and latent heat of condensation. In selecting cooling equipment, distinguish between sensible and latent heat gain: every cooling apparatus has different maximum removal capacities for sensible versus latent heat for particular operating conditions. In extremely dry climates, humidification may be required, rather than dehumidification, to maintain thermal comfort. Radiant Heat Gain. Radiant energy must first be absorbed by surfaces that enclose the space (walls, floor, and ceiling) and objects in the space (furniture, etc.). When these surfaces and objects become warmer than the surrounding air, some of their heat transfers to the air by convection. The composite heat storage capacity of these surfaces and objects determines the rate at which their respective surface temperatures increase for a given radiant input, and thus governs the relationship between the radiant portion of heat gain and its corresponding part of the space cooling load (Figure 1). The thermal storage effect is critical in differentiating between instantaneous heat gain for a given space and its cooling load at that moment. Predicting
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18.2
2017 ASHRAE Handbook—Fundamentals (SI) instantaneously stored. This time lag must be considered when calculating cooling load, because the load required for the space can be much lower than the instantaneous heat gain being generated, and the space’s peak load may be significantly affected. Accounting for the time delay effect is the major challenge in cooling load calculations. Several methods, including the two presented in this chapter, have been developed to take the time delay effect into consideration.
1.2 Fig. 1 Origin of Difference Between Magnitude of Instantaneous Heat Gain and Instantaneous Cooling Load
COOLING LOAD CALCULATION METHODS
This chapter presents two load calculation methods that vary significantly from previous methods. The technology involved, however (the principle of calculating a heat balance for a given space) is not new. The first of the two methods is the heat balance (HB) method; the second is radiant time series (RTS), which is a simplification of the HB procedure. Both methods are explained in their respective sections. Cooling load calculation of an actual, multiple-room building requires a complex computer program implementing the principles of either method.
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Cooling Load Calculations in Practice
Fig. 2
Thermal Storage Effect in Cooling Load from Lights
the nature and magnitude of this phenomenon to estimate a realistic cooling load for a particular set of circumstances has long been of interest to design engineers; the Bibliography lists some early work on the subject. Space Cooling Load. This is the rate at which sensible and latent heat must be removed from the space to maintain a constant space air temperature and humidity. The sum of all space instantaneous heat gains at any given time does not necessarily (or even frequently) equal the cooling load for the space at that same time. Space Heat Extraction Rate. The rates at which sensible and latent heat are removed from the conditioned space equal the space cooling load only if the room air temperature and humidity are constant. Along with the intermittent operation of cooling equipment, control systems usually allow a minor cyclic variation or swing in room temperature; humidity is often allowed to float, but it can be controlled. Therefore, proper simulation of the control system gives a more realistic value of energy removal over a fixed period than using values of the space cooling load. However, this is primarily important for estimating energy use over time; it is not needed to calculate design peak cooling load for equipment selection. Cooling Coil Load. The rate at which energy is removed at a cooling coil serving one or more conditioned spaces equals the sum of instantaneous space cooling loads (or space heat extraction rate, if it is assumed that space temperature and humidity vary) for all spaces served by the coil, plus any system loads. System loads include fan heat gain, duct heat gain, and outdoor air heat and moisture brought into the cooling equipment to satisfy the ventilation air requirement.
Time Delay Effect Energy absorbed by walls, floor, furniture, etc., contributes to space cooling load only after a time lag. Some of this energy is still present and reradiating even after the heat sources have been switched off or removed, as shown in Figure 2. There is always significant delay between the time a heat source is activated, and the point when reradiated energy equals that being
Load calculations should accurately describe the building. All load calculation inputs should be as accurate as reasonable, without using safety factors. Introducing compounding safety factors at multiple levels in the load calculation results in an unrealistic and oversized load. Variation in heat transmission coefficients of typical building materials and composite assemblies, differing motivations and skills of those who construct the building, unknown infiltration rates, and the manner in which the building is actually operated are some of the variables that make precise calculation impossible. Even if the designer uses reasonable procedures to account for these factors, the calculation can never be more than a good estimate of the actual load. Frequently, a cooling load must be calculated before every parameter in the conditioned space can be properly or completely defined. An example is a cooling load estimate for a new building with many floors of unleased spaces for which detailed partition requirements, furnishings, lighting, and layout cannot be predefined. Potential tenant modifications once the building is occupied also must be considered. Load estimating requires proper engineering judgment that includes a thorough understanding of heat balance fundamentals. Perimeter spaces exposed to high solar heat gain often need cooling during sunlit portions of traditional heating months, as do completely interior spaces with significant internal heat gain. These spaces can also have significant heating loads during nonsunlit hours or after periods of nonoccupancy, when adjacent spaces have cooled below interior design temperatures. The heating loads involved can be estimated conventionally to offset or to compensate for them and prevent overheating, but they have no direct relationship to the spaces’ design heating loads. Correct design and sizing of air-conditioning systems require more than calculation of the cooling load in the space to be conditioned. The type of air-conditioning system, ventilation rate, reheat, fan energy, fan location, duct heat loss and gain, duct leakage, heat extraction lighting systems, type of return air system, and any sensible or latent heat recovery all affect system load and component sizing. Adequate system design and component sizing require that system performance be analyzed as a series of psychrometric processes. System design could be driven by either sensible or latent load, and both need to be checked. In a sensible-load-driven space (the most common case), the cooling supply air has surplus capacity to dehumidify, but this is usually permissible. For a space driven by latent
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Nonresidential Cooling and Heating Load Calculations load (e.g., an auditorium), supply airflow based on sensible load is likely not to have enough dehumidifying capability, so subcooling and reheating or some other dehumidification process is needed. This chapter is primarily concerned with a given space or zone in a building. When estimating loads for a group of spaces (e.g., for an air-handling system that serves multiple zones), the assembled zones must be analyzed to consider (1) the simultaneous effects taking place; (2) any diversification of heat gains for occupants, lighting, or other internal load sources; (3) ventilation; and/or (4) any other unique circumstances. With large buildings that involve more than a single HVAC system, simultaneous loads and any additional diversity also must be considered when designing the central equipment that serves the systems. Methods presented in this chapter are expressed as hourly load summaries, reflecting 24 h input schedules and profiles of the individual load variables. Specific systems and applications may require different profiles.
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1.3 DATA ASSEMBLY Calculating space cooling loads requires detailed building design information and weather data at design conditions. Generally, the following information should be compiled. Building Characteristics. Building materials, component size, external surface colors, and shape are usually determined from building plans and specifications. Configuration. Determine building location, orientation, and external shading from building plans and specifications. Shading from adjacent buildings can be determined from a site plan or by visiting the proposed site, but its probable permanence should be carefully evaluated before it is included in the calculation. The possibility of abnormally high ground-reflected solar radiation (e.g., from adjacent water, sand, or parking lots) or solar load from adjacent reflective buildings should not be overlooked. Outdoor Design Conditions. Obtain appropriate weather data, and select outdoor design conditions. Chapter 14 provides information for many weather stations; note, however, that these design dry-bulb and mean coincident wet-bulb temperatures may vary considerably from data traditionally used in various areas. Use judgment to ensure that results are consistent with expectations. Also, consider prevailing wind velocity and the relationship of a project site to the selected weather station. Recent research projects have greatly expanded the amount of available weather data (e.g., ASHRAE 2012). In addition to the conventional dry bulb with mean coincident wet bulb, data are now available for wet bulb and dew point with mean coincident dry bulb. Peak space load generally coincides with peak solar or peak dry bulb, but peak system load often occurs at peak wet-bulb temperature. The relationship between space and system loads is discussed further in following sections of the chapter. To estimate conductive heat gain through exterior surfaces and infiltration and outdoor air loads at any time, applicable outdoor dryand wet-bulb temperatures must be used. Chapter 14 gives monthly cooling load design values of outdoor conditions for many locations. These are generally midafternoon conditions; for other times of day, the daily range profile method described in Chapter 14 can be used to estimate dry- and wet-bulb temperatures. Peak cooling load is often determined by solar heat gain through fenestration; this peak may occur in winter months and/or at a time of day when outdoor air temperature is not at its maximum. Indoor Design Conditions. Select indoor dry-bulb temperature, indoor relative humidity, and ventilation rate. Include permissible variations and control limits. Consult ASHRAE Standard 90.1 for energy-savings conditions, and Standard 55 for ranges of indoor conditions needed for thermal comfort. Internal Heat Gains and Operating Schedules. Obtain planned density and a proposed schedule of lighting, occupancy, internal
18.3 equipment, appliances, and processes that contribute to the internal thermal load. Areas. Use consistent methods for calculation of building areas. For fenestration, the definition of a component’s area must be consistent with associated ratings. Gross surface area. It is efficient and conservative to derive gross surface areas from outer building dimensions, ignoring wall and floor thicknesses and avoiding separate accounting of floor edge and wall corner conditions. Measure floor areas to the outside of adjacent exterior walls or to the centerline of adjacent partitions. When apportioning to rooms, façade area should be divided at partition centerlines. Wall height should be taken as floor-to-floor height. The outer-dimension procedure is expedient for load calculations, but it is not consistent with rigorous definitions used in buildingrelated standards. The resulting differences do not introduce significant errors in this chapter’s procedures. Fenestration area. As discussed in Chapter 15, fenestration ratings [U-factor and solar heat gain coefficient (SHGC)] are based on the entire product area, including frames. Thus, for load calculations, fenestration area is the area of the rough opening in the wall or roof. Net surface area. Net surface area is the gross surface area less any enclosed fenestration area.
2.
INTERNAL HEAT GAINS
Internal heat gains from people, lights, motors, appliances, and equipment can contribute the majority of the cooling load in a modern building. As building envelopes have improved in response to more restrictive energy codes, internal loads have increased because of factors such as increased use of computers and the advent of dense-occupancy spaces (e.g., call centers). Internal heat gain calculation techniques are identical for both heat balance (HB) and radiant time series (RTS) cooling-load calculation methods, so internal heat gain data are presented here independent of calculation methods.
2.1
PEOPLE
Table 1 gives representative rates at which sensible heat and moisture are emitted by humans in different states of activity. In high-density spaces, such as auditoriums, these sensible and latent heat gains comprise a large fraction of the total load. Even for shortterm occupancy, the extra sensible heat and moisture introduced by people may be significant. See Chapter 9 for detailed information; however, Table 1 summarizes design data for common conditions. The conversion of sensible heat gain from people to space cooling load is affected by the thermal storage characteristics of that space because some percentage of the sensible load is radiant energy. Latent heat gains are usually considered instantaneous, but research is yielding practical models and data for the latent heat storage of and release from common building materials.
2.2
LIGHTING
Because lighting is often a major space cooling load component, an accurate estimate of the space heat gain it imposes is needed. Calculation of this load component is not straightforward; the rate of cooling load from lighting at any given moment can be quite different from the heat equivalent of power supplied instantaneously to those lights, because of heat storage.
Instantaneous Heat Gain from Lighting The primary source of heat from lighting comes from lightemitting elements, or lamps, although significant additional heat may be generated from ballasts and other appurtenances in the
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18.4
2017 ASHRAE Handbook—Fundamentals (SI)
Table 1 Representative Rates at Which Heat and Moisture Are Given Off by Human Beings in Different States of Activity Total Heat, W Adjusted, M/F a
Sensible Heat, W
Latent Heat, W
Degree of Activity
Location
Seated at theater Seated, very light work
Theater Offices, hotels, apartments
115 130
105 115
70 70
35 45
Moderately active office work Standing, light work; walking Walking, standing Sedentary work
Offices, hotels, apartments Department store; retail store Drug store, bank Restaurantc
140 160 160 145
130 130 145 160
75 75 75 80
55 55 70 80
Light bench work Moderate dancing Walking 4.8 km/h; light machine work
Factory Dance hall Factory
235 265 295
220 250 295
80 90 110
Bowlingd Heavy work Heavy machine work; lifting Athletics
Bowling alley Factory Factory Gymnasium
440 440 470 585
425 425 470 525
170 170 185 210
Notes: 1. Tabulated values are based on 24°C room dry-bulb temperature. For 27°C room dry bulb, total heat remains the same, but sensible heat values should be decreased by approximately 20%, and latent heat values increased accordingly. 2. Also see Table 4, Chapter 9, for additional rates of metabolic heat generation. 3. All values are rounded to nearest 5 W.
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Adult Male
= = = =
High V
60
27
58
38
140 160 185
49
35
255 255 285 315
54
19
and assumes that gain from an adult female is 85% of that for an adult male, and gain from a child is 75% of that for an adult male. b Values approximated from data in Table 6, Chapter 9, where V is air velocity with limits shown in that table. c Adjusted heat gain includes 18 W for food per individual (9 W sensible and 9 W latent). d Figure one person per alley actually bowling, and all others as sitting (117 W) or standing or walking slowly (231 W).
(1)
where qel W Ful Fsa
Low V
a Adjusted heat gain is based on normal percentage of men, women, and children for the application listed,
luminaires. Generally, the instantaneous rate of sensible heat gain from electric lighting may be calculated from qel = WFul Fsa
% Sensible Heat that is Radiantb
heat gain, W total light wattage, W lighting use factor lighting special allowance factor
The total light wattage is obtained from the ratings of all lamps installed, both for general illumination and for display use. Ballasts are not included, but are addressed by a separate factor. Wattages of magnetic ballasts are significant; the energy consumption of highefficiency electronic ballasts might be insignificant compared to that of the lamps. The lighting use factor is the ratio of wattage in use, for the conditions under which the load estimate is being made, to total installed wattage. For commercial applications such as stores, the use factor is generally 1.0. The special allowance factor is the ratio of the lighting fixtures’ power consumption, including lamps and ballast, to the nominal power consumption of the lamps. For incandescent lights, this factor is 1. For fluorescent lights, it accounts for power consumed by the ballast as well as the ballast’s effect on lamp power consumption. The special allowance factor can be less than 1 for electronic ballasts that lower electricity consumption below the lamp’s rated power consumption. Use manufacturers’ values for system (lamps + ballast) power, when available. For high-intensity-discharge lamps (e.g. metal halide, mercury vapor, high- and low-pressure sodium vapor lamps), the actual lighting system power consumption should be available from the manufacturer of the fixture or ballast. Ballasts available for metal halide and high-pressure sodium vapor lamps may have special allowance factors from about 1.3 (for low-wattage lamps) down to 1.1 (for high-wattage lamps). An alternative procedure is to estimate the lighting heat gain on a per-square-metre basis. Such an approach may be required when final lighting plans are not available. Table 2 shows the maximum lighting power density (LPD) (lighting heat gain per square metre) allowed by ASHRAE Standard 90.1-2013 for a range of space types.
In addition to determining the lighting heat gain, the fraction of lighting heat gain that enters the conditioned space may need to be distinguished from the fraction that enters an unconditioned space; of the former category, the distribution between radiative and convective heat gain must be established. Fisher and Chantrasrisalai (2006) and Zhou et al. (2016) experimentally studied 12 luminaire types and recommended several categories of luminaires, as shown in Table 3. The table provides a range of design data for the conditioned space fraction, short-wave radiative fraction, and long-wave radiative fraction under typical operating conditions: airflow rate of 5 L/(s·m2), supply air temperature between 15 and 16.7°C, and room air temperature between 22 and 24°C. The recommended fractions in Table 3 are based on lighting heat input rates range of 9.7 to 28 W/m2. For higher design power input, the lower bounds of the space and short-wave fractions should be used; for design power input below this range, the upper bounds of the space and short-wave fractions should be used. The space fraction in the table is the fraction of lighting heat gain that goes to the room; the fraction going to the plenum can be computed as 1 – the space fraction. The radiative fraction is the radiative part of the lighting heat gain that goes to the room. The convective fraction of the lighting heat gain that goes to the room is 1 – the radiative fraction. Using values in the middle of the range yields sufficiently accurate results. However, values that better suit a specific situation may be determined according to the notes for Table 3. Table 3’s data apply to both ducted and nonducted returns. However, application of the data, particularly the ceiling plenum fraction, may vary for different return configurations. For instance, for a room with a ducted return, although a portion of the lighting energy initially dissipated to the ceiling plenum is quantitatively equal to the plenum fraction, a large portion of this energy would likely end up as the conditioned space cooling load and a small portion would end up as the cooling load to the return air. If the space airflow rate is different from the typical condition [i.e., about 5 L/(s·m2)], Figure 3 can be used to estimate the lighting heat gain parameters. Design data shown in Figure 3 are only applicable for the recessed fluorescent luminaire without lens. Although design data presented in Table 3 and Figure 3 can be used for a vented luminaire with side-slot returns, they are likely not applicable for a vented luminaire with lamp compartment returns, because in the latter case, all heat convected in the vented luminaire
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Nonresidential Cooling and Heating Load Calculations Table 2 Common Space Types* Atrium 12.2 m high
Licensed for single user. © 2017 ASHRAE, Inc.
12.2 m high
LPD,
W/m2
1.1/m total height 4.3 + 0.7/m total height
Audience Seating Area 6.8 In auditorium 8.9 In convention center 7.1 In gymnasium 12.3 In motion picture theater 3.1 In penitentiary 26.2 In performing arts theater 16.5 In religious building 4.7 In sports arena 4.7 All other audience seating areas Banking Activity Area 11.9 Breakroom (See Lounge/Breakroom) Classroom/Lecture Hall/Training Room 14.5 In penitentiary 13.4 All other classrooms/lecture halls/ training rooms Conference/Meeting/Multipurpose Room
13.3
Confinement Cells Copy/Print Room Corridorb In facility for visually impaired (and not used primarily by staff)c
8.8 7.8
In hospital In manufacturing facility All other corridors Courtroom Computer Room Dining Area In penitentiary In facility for visually impaired (and not used primarily by staff)c
10.7 4.4 7.1 18.6 18.4
In bar/lounge or leisure dining In cafeteria or fast food dining In family dining All other dining areas Electrical/Mechanical Roomf Emergency Vehicle Garage Food Preparation Area Guest Room Laboratory In or as classroom All other laboratories Laundry/Washing Area
9.9
10.4 28.5 W/m2 11.6 7.0 9.6 7.0 4.6 6.1 13.1 9.8 15.5 19.5 6.5
Source: ASHRAE Standard 90.1-2013. aIn cases where both a common space type and a buildingspecific type are listed, the building-specific space type applies. bIn corridors, extra lighting power density allowance is granted when corridor width is 2.4 m and is not based on room/corridor ratio (RCR).
18.5
Lighting Power Densities Using Space-by-Space Method Common Space Typesa
LPD, W/m2
Building-Specific Space Types*
5.1
Health Care Facility In exam/treatment room In imaging room In medical supply room In nursery In nurses’ station In operating room In patient room In physical therapy room In recovery room Library In reading area In stacks Manufacturing Facility In detailed manufacturing area In equipment room In extra-high-bay area (15.2 m floor-to-ceiling height)
Loading Dock, Interior Lobby In facility for the visually impaired (and not used primarily by staff)c For elevator In hotel In motion picture theater In performing arts theater All other lobbies Locker Room Lounge/Breakroom In health care facility All other lounges/breakrooms Office Enclosed Open plan Parking Area, Interior Pharmacy Area Restroom In facility for the visually impaired (and not used primarily by staff)c
19.4 7.0 11.5 6.4 21.6 9.7 8.1 10.0 7.9 12.0 10.6 2.1 18.1 13.1
All other restrooms 10.6 15.5 Sales Aread 5.9 Seating Area, General Stairway Space containing stairway determines LPD and control requirements for stairway. Stairwell Storage Room 4.65 m2 All other storage rooms Vehicular Maintenance Area Workshop
7.4 13.3 6.8 7.3 17.2
Building-Specific Space Types* LPD, W/m2 Facility for Visually Impairedc In chapel (used primarily by residents) In recreation room/common living room (and not used primarily by staff)
23.8 26.0
Automotive (See Vehicular Maintenance Area) Convention Center: Exhibit Space Dormitory/Living Quarters Fire Station: Sleeping Quarters Gymnasium/Fitness Center In exercise area In playing area cA
15.7 4.2 0.22 7.8 13.0
facility for the visually impaired one that can be documented as being designed to comply with light levels in ANSI/IES RP-28 and is (or will be) licensed by local/ state authorities for either senior long-term care, adult daycare, senior support, and/or people with special visual needs.
is likely to go directly to the ceiling plenum, resulting in zero convective fraction and a much lower space fraction. Therefore, the design data should only be used for a configuration where conditioned air is returned through the ceiling grille or luminaire side slots.
LPD, W/m2 18.0 16.3 9.5 7.6 26.8 6.7 9.9 12.4 11.5 18.4 13.9 8.0 11.3
In high-bay area (7.6 to 15.2 m floor-to-ceiling height)
13.3
In low-bay area (7.6 m floor-to-ceiling height)
12.9
Museum In general exhibition area In restoration room Performing Arts Theater, Dressing Room Post Office, Sorting Area Religious Buildings In fellowship hall In worship/pulpit/choir area Retail Facilities In dressing/fitting room In mall concourse Sports Arena, Playing Area For Class I facility For Class II facility For Class III facility For Class IV facility Transportation Facility In baggage/carousel area In an airport concourse At a terminal ticket counter Warehouse—Storage Area For medium to bulky, palletized items For smaller, hand-carried itemse
11.4 11.0 6.6 10.2 6.9 16.5 7.7 11.9 39.7 25.9 19.4 13.0 5.7 3.9 8.7 6.2 10.2
dFor
accent lighting, see section 9.6.2(b) of ASHRAE Standard 90.1-2013. eSometimes called a picking area. fAn additional 5.7 W/m2 is allowed only if this additional lighting is controlled separately from the base allowance of 4.5 W/m2.
For other luminaire types, it may be necessary to estimate the heat gain for each component as a fraction of the total lighting heat gain by using judgment to estimate heat-to-space and heat-to-return percentages.
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18.6
2017 ASHRAE Handbook—Fundamentals (SI) Table 3
Luminaire Category
Space Fraction
Radiative Fraction Notes
Recessed fluorescent luminaire without lens
0.64 to 0.74
0.48 to 0.68
• Use middle values in most situations • May use higher space fraction, and lower radiative fraction for luminaire with side-slot returns • May use lower values of both fractions for direct/indirect luminaire • May use higher values of both fractions for ducted returns
Recessed fluorescent luminaire with lens
0.40 to 0.50
0.61 to 0.73
• May adjust values in the same way as for recessed fluorescent luminaire without lens
Downlight compact fluorescent luminaire
0.12 to 0.24
0.95 to 1.0
• Use middle or high values if detailed features are unknown • Use low value for space fraction and high value for radiative fraction if there are large holes in luminaire’s reflector
Downlight incandescent luminaire
0.70 to 0.80
0.95 to 1.0
• Use middle values if lamp type is unknown • Use low value for space fraction if standard lamp (i.e. A-lamp) is used • Use high value for space fraction if reflector lamp (i.e. BR-lamp) is used
1.0
0.5 to 0.57
• Use lower value for radiative fraction for surface-mounted luminaire • Use higher value for radiative fraction for pendant luminaire
Recessed LED troffer partial aperture diffuser
0.49 to 0.64
0.37 to 0.47
• Use middle value in most cases. • May use higher space fraction for ducted return configuration and lower space fraction for high supply air temperature. • May use higher radiant value for ducted return configuration and lower value for large supply airflow rate.
Recessed LED troffer uniform diffuser
0.44 to 0.66
0.32 to 0.41
• Use middle value in most cases. • May use higher space fraction for smaller supply airflow rate and lower value for larger supply airflow rate. • May use higher radiant value for ducted return configuration and lower value for larger supply airflow rate.
0.59
0.51
Recessed LED downlight
0.40 to 0.56
0.15 to 0.18
• Use middle value in most cases. • May use higher space fraction value for high supply air temperature and lower value for smaller air flowrate. • May use higher radiant value for dimming control and lower value for large supply air flowrate.
Recessed LED retrofit kit 2×4
0.41 to 0.53
0.31 to 0.42
• Use middle value in most cases. • May use higher space fraction value for large supply air flowrate and lower value for ducted return configuration. • May use higher radiant value for ducted return configuration and lower value for larger supply airflow rate.
Recessed LED color tuning fixture
0.53 to 0.56
0.40 to 0.42
Use middle value in most cases.
Non-in-ceiling fluorescent luminaire
Licensed for single user. © 2017 ASHRAE, Inc.
Lighting Heat Gain Parameters for Typical Operating Conditions
Recessed high-efficacy LED troffer
High-bay LED fixture
1.0
0.42 to 0.51
Use middle value in most cases.
Linear pendant LED fixture
1.0
0.55 to 0.60
Use middle value in most cases.
Sources: Fisher and Chantrasrisalai (2006); Zhou et al. 2016.
radiant time factors (RTFs) may be more appropriate than nonsolar RTFs. (Solar RTFs are calculated assuming most solar radiation is intercepted by the floor; nonsolar RTFs assume uniform distribution by area over all interior surfaces.) This effect may be significant for rooms where lighting heat gain is high and for which solar RTFs are significantly different from nonsolar RTFs.
2.3
ELECTRIC MOTORS
Instantaneous sensible heat gain from equipment operated by electric motors in a conditioned space is calculated as qem = (P/EM)FUM FLM
(2)
where Fig. 3
Lighting Heat Gain Parameters for Recessed Fluorescent Luminaire Without Lens (Fisher and Chantrasrisalai 2006)
Because of the directional nature of downlight luminaires, a large portion of the short-wave radiation typically falls on the floor. When converting heat gains to cooling loads in the RTS method, the solar
qem P EM FUM FLM
= = = = =
heat equivalent of equipment operation, W motor power rating, W motor efficiency, decimal fraction <1.0 motor use factor, 1.0 or decimal fraction <1.0 motor load factor, 1.0 or decimal fraction <1.0
The motor use factor may be applied when motor use is known to be intermittent, with significant nonuse during all hours of operation
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Nonresidential Cooling and Heating Load Calculations Table 4A Minimum Nominal Full-Load Efficiency for 60 Hz NEMA General-Purpose Electric Motors (Subtype I) Rated 600 V or Less (Random Wound)*
Licensed for single user. © 2017 ASHRAE, Inc.
Number of Poles
Table 4B Minimum Average Full-Load Efficiency for Polyphase Small Electric Motors*
Open Drip-Proof Totally Enclosed Motors Fan-Cooled Motors 2
4
6
2
4
6
Synchronous Speed (RPM) 3600 1800 1200
3600 1800 1200
Motor Kilowatts 0.8 1.1 1.5 2.2 3.7 5.6 7.5 11.1 14.9 18.7 22.4 29.8 37.3 44.8 56.0 74.6 93.3 111.9 149.2
77.0 84.0 85.5 86.5 88.5 89.5 90.2 91.0 91.0 91.7 91.7 92.4 93.0 93.6 93.6 94.1 95.0 95.0 95.4
77.0 84.0 85.5 85.5 86.5 88.5 89.5 90.2 91.0 91.7 91.7 92.4 93.0 93.6 93.6 93.6 94.1 94.1 95.0
85.5 86.5 86.5 89.5 89.5 91.0 91.7 93.0 93.0 93.6 94.1 94.1 94.5 95.0 95.0 95.4 95.4 95.8 95.8
82.5 86.5 87.5 88.5 89.5 90.2 91.7 91.7 92.4 93.0 93.6 94.1 94.1 94.5 94.5 95.0 95.0 95.4 95.4
85.5 86.5 86.5 89.5 89.5 91.7 91.7 92.4 93.0 93.6 93.6 94.1 94.5 95.0 95.4 95.4 95.4 95.8 96.2
82.5 87.5 88.5 89.5 89.5 91.0 91.0 91.7 91.7 93.0 93.0 94.1 94.1 94.5 94.5 95.0 95.0 95.8 95.8
Source: ASHRAE Standard 90.1-2013. *Nominal efficiencies established in accordance with NEMA Standard MG1. Design A and Design B are National Electric Manufacturers Association (NEMA) design class designations for fixed-frequency small and medium AC squirrel-cage induction motors.
(e.g., overhead door operator). For conventional applications, its value is 1.0. The motor load factor is the fraction of the rated load delivered under the conditions of the cooling load estimate. Equation (2) assumes that both the motor and driven equipment are in the conditioned space. If the motor is outside the space or airstream, qem = PFUM FLM
(3)
When the motor is inside the conditioned space or airstream but the driven machine is outside, 1.0 – EM qem = P --------------------- FUMFLM EM
18.7
(4)
Equation (4) also applies to a fan or pump in the conditioned space that exhausts air or pumps fluid outside that space. Table 4A and 4B gives minimum efficiencies and related data representative of typical electric motors from ASHRAE Standard 90.1-2013. If electric motor load is an appreciable portion of cooling load, the motor efficiency should be obtained from the manufacturer. Also, depending on design, maximum efficiency might occur anywhere between 75 to 110% of full load; if under- or overloaded, efficiency could vary from the manufacturer’s listing.
Overloading or Underloading Heat output of a motor is generally proportional to motor load, within rated overload limits. Because of typically high no-load motor current, fixed losses, and other reasons, FLM is generally assumed to be unity, and no adjustment should be made for underloading or overloading unless the situation is fixed and can be accurately established, and reduced-load efficiency data can be obtained from the motor manufacturer.
Full-Load Efficiency for Motors Manufactured on or after March 9, 2015, % Number of Poles
Open Motors
Synchronous Speed (RPM)
2
4
6
3600
1800
1200
Motor Kilowatts 0.19 0.25 0.37 0.56 0.75 1.1 1.5 2.2
65.6 69.5 73.4 76.8 77.0 84.0 85.5 85.5
69.5 73.4 78.2 81.1 83.5 86.5 86.5 86.9
67.5 71.4 75.3 81.7 82.5 83.8 N/A N/A
*Average full-load efficiencies established in accordance with 10 CFR 431.
Radiation and Convection Unless the manufacturer’s technical literature indicates otherwise, motor heat gain normally should be equally divided between radiant and convective components for the subsequent cooling load calculations.
2.4
APPLIANCES
A cooling load estimate should take into account heat gain from all appliances (electrical, gas, or steam). Because of the variety of appliances, applications, schedules, use, and installations, estimates can be very subjective. Often, the only information available about heat gain from equipment is that on its nameplate, which can overestimate actual heat gain for many types of appliances, as discussed in the section on Office Equipment.
Cooking Appliances These appliances include common heat-producing cooking equipment found in conditioned commercial kitchens. Marn (1962) concluded that appliance surfaces contributed most of the heat to commercial kitchens and that when appliances were installed under an effective hood, the cooling load was independent of the fuel or energy used for similar equipment performing the same operations. Gordon et al. (1994) and Smith et al. (1995) found that gas appliances may exhibit slightly higher heat gains than their electric counterparts under wall-canopy hoods operated at typical ventilation rates. This is because heat contained in combustion products exhausted from a gas appliance may increase the temperatures of the appliance and surrounding surfaces, as well as the hood above the appliance, more so than the heat produced by its electric counterpart. These higher-temperature surfaces radiate heat to the kitchen, adding moderately to the radiant gain directly associated with the appliance cooking surface. Marn (1962) confirmed that, where appliances are installed under an effective hood, only radiant gain adds to the cooling load; convective and latent heat from cooking and combustion products are exhausted and do not enter the kitchen. Gordon et al. (1994) and Smith et al. (1995) substantiated these findings. Chapter 33 of the 2015 ASHRAE Handbook—HVAC Applications has more information on kitchen ventilation. Sensible Heat Gain for Hooded Cooking Appliances. To establish a heat gain value, nameplate energy input ratings may be used with appropriate usage and radiation factors. Where specific rating data are not available (nameplate missing, equipment not yet purchased, etc.), representative heat gains listed in Tables 5A to 5E (Swierczyna et al. 2008, 2009) for a wide variety of commonly
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18.8
2017 ASHRAE Handbook—Fundamentals (SI)
encountered equipment items. In estimating appliance load, probabilities of simultaneous use and operation for different appliances located in the same space must be considered. Radiant heat gain from hooded cooking equipment can range from 15 to 45% of the actual appliance energy consumption (Gordon et al. 1994; Smith et al. 1995; Swierczyna et al. 2008; Talbert et al. 1973). This ratio of heat gain to appliance energy consumption may be expressed as a radiation factor, and it is a function of both appliance type and fuel source. The radiation factor FR is applied to the average rate of appliance energy consumption, determined by applying usage factor FU to the nameplate or rated energy input. Marn (1962) found that radiant heat temperature rise can be substantially reduced by shielding the fronts of cooking appliances. Although this approach may not always be practical in a commercial kitchen, radiant gains can also be reduced by adding side panels or partial enclosures that are integrated with the exhaust hood. Heat Gain from Meals. For each meal served, approximately 15 W of heat, of which 75% is sensible and 25% is latent, is transferred to the dining space.
Heat Gain for Generic Appliances. The average rate of appliance energy consumption can be estimated from the nameplate or rated energy input qinput by applying a duty cycle or usage factor FU . Thus, sensible heat gain qs for generic electric, steam, and gas appliances installed under a hood can be estimated using one of the following equations: qs = qinput FU FR
(5)
qs = qinput FL
(6)
or where FL is the ratio of sensible heat gain to the manufacturer’s rated energy input. However, ASHRAE research (Swierczyna et al. 2008, 2009) showed the design value for heat gain from a hooded appliance at idle (ready-to-cook) conditions based on its energy consumption rate is, at best, a rough estimate. When appliance heat gain measurements during idle conditions were regressed against energy consumption rates for gas and electric appliances, the appliances’ emissivity, insulation, and surface cooling (e.g., through ventilation rates) scattered the data points widely, with large devia-
Licensed for single user. © 2017 ASHRAE, Inc.
Table 5A Recommended Rates of Radiant and Convective Heat Gain from Unhooded Electric Appliances During Idle (Ready-to-Cook) Conditions Energy Rate, W Appliance insulateda
Cabinet: hot serving (large), hot serving (large), uninsulated proofing (large)a proofing (small-15 shelf) Cheesemelterb Coffee brewing urn Drawer warmers, 2-drawer (moist holding)a Egg cookerb Espresso machinea Food warmer: steam table (2-well-type) Freezer (small) Fryer, countertop, open deep fatb Griddle, countertopb Hot dog rollerb Hot plate: single element, high speedb Hot-food case (dry holding)a Hot-food case (moist holding)a Induction hob, countertopb Microwave oven: commercialb Oven: countertop conveyorized bake/finishingb Paninib Popcorn popperb Rapid-cook oven (quartz-halogen)a Rapid-cook oven (microwave/convection)b Reach-in refrigeratora Refrigerated prep tablea Rice cookerb Soup warmerb Steamer (bun)b Steamer, countertopb Toaster: 4-slice pop up (large): cooking contact (vertical)b conveyor (large) small conveyorb Tortilla grillb Waffle ironb
Rated
Standby
1993 1993 5099 4191 2400 3810 1202 2380 2403 1495 791 4600 8000 1600 1100 9115 9115 5000 1700 5000 1800 850 12 016 5700 1407 586 1550 800 1500 8300 1788 2600 9613 1745 2200 2700
352 1026 410 1143 976 352 147 249 352 1026 322 431 1771 1240 982 733 967 0 0 3932 673 115 0 1141 352 264 82 390 200 344 879 759 3019 1702 1034 267
Rate of Heat Gain, W Sensible Sensible Radiant Convective 117 205 352 0 443 59 0 65 117 88 147 202 848 267 314 264 264 0 0 718 195 28 0 96 88 176 14 0 32 0 59 180 879 358 254 60
Sources: Swierczyna et al. (2008, 2009), with the following exceptions as noted. aSwierczyna et al. (2009) only. bAdditions and updates from ASHRAE research project RP-1631 (Kong and Zhang 2016; Zhang et al. 2016).
234 821 0 264 533 88 0 184 234 176 176 229 923 973 668 469 528 0 0 3214 478 87 0 1045 264 88 68 53 168 248 410 579 2139 1344 780 207
Latent
Total
Usage Factor FU
0 0 59 879 0 205 59 0 0 762 0 0 0 0 0 0 176 0 0 0 0 0 0 0 0 0 0 337 0 96 293 0 0 0 0 0
352 1026 410 1143 976 352 59 249 352 1026 322 431 1771 1240 982 733 967 0 0 3932 673 115 0 1141 352 264 82 390 200 344 762 759 3019 1702 1034 267
0.18 0.51 0.08 0.27 0.41 0.08 0.12 0.10 0.15 0.69 0.41 0.09 0.22 0.77 0.89 0.08 0.11 0.00 0 0.79 0.37 0.14 0 0.20 0.25 0.45 0.05 0.49 0.13 0.04 0.49 0.29 0.31 0.98 0.47 0.10
Radiation Factor FR 0.33 0.20 0.86 0.00 0.45 0.17 0.00 0.26 0.33 0.08 0.45 0.47 0.48 0.22 0.32 0.36 0.27 0.00 0.00 0.18 0.29 0.24 0.00 0.08 0.25 0.67 0.17 0.00 0.16 0.00 0.07 0.24 0.29 0.21 0.25 0.22
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Nonresidential Cooling and Heating Load Calculations
18.9
Table 5B Recommended Rates of Radiant and Convective Heat Gain from Unhooded Electric Appliances during Cooking Conditions Energy Rate, W
Licensed for single user. © 2017 ASHRAE, Inc.
Appliance Cheesemelter Egg cooker Fryer, countertop, open deep fryer Griddle, countertop Hot dog roller Hot plate, single burner Induction hob, countertop Oven, conveyor Microwave Rapid cook Panini grill Popcorn popper Rice cooker Soup warmer Steamer (bun) Steamer, countertop Toaster, conveyor Vertical Tortilla grill Waffle maker
Rated
Rate of Heat Gain, W
Cooking
2400 2380 4600 8000 1600 1100 5000 5000 1700 5700 1800 850 1550 800 1500 8300 1745 2600 2200 2700
Sensible Radiant Sensible Convective
2714 1191 3818 3280 1577 985 653 4292 2363 2310 1374 576 1159 842 791 7731 1705 1841 2194 1180
443 65 202 848 267 313 0 718 0 96 195 28 14 0 32 0 358 180 254 60
1094 369 492 631 611 627 318 2454 934 1234 718 236 95 85 240 499 974 715 1267 357
Latent 599 630 1629 1277 679 44 335 193 995 771 150 192 44 716 511 6934 373 322 673 559
Usage Radiation Factor FU Factor FR
Total 2136 1065 2323 2757 1556 985 653 3365 1929 2102 1062 457 153 801 783 7433 1705 1218 2194 975
1.13 0.50 0.83 0.41 0.99 0.90 0.13 0.86 1.39 0.41 0.76 0.68 0.75 1.05 0.53 0.93 0.98 0.71 1.00 0.44
0.16 0.05 0.05 0.26 0.17 0.32 0.00 0.17 0.00 0.04 0.14 0.05 0.01 0.00 0.04 0.00 0.21 0.10 0.12 0.05
Source: ASHRAE research project RP-1631 (Zhang et al. 2015).
Table 5C Recommended Rates of Radiant Heat Gain from Hooded Electric Appliances During Idle (Ready-to-Cook) Conditions Energy Rate, W Appliance Broiler: underfired 900 mm Cheesemelter* Fryer, kettle Open deep-fat, 1-vat Pressure Griddle, double-sided 900 mm (clamshell down)* (Clamshell up)* Flat 900 mm Small 900 mm* Induction cooktop* Induction wok* Oven, combi: combi-mode* Combi: convection mode Oven, convection full-size Convection half-size* Pasta cooker* Range top, top off/oven on* 3 elements on/oven off 6 elements on/oven off 6 elements on/oven on Range, hot-top Rotisserie* Salamander* Steam kettle, large (225 L), simmer lid down* small (150 L), simmer lid down* Steamer, compartment, atmospheric* Tilting skillet/braising pan
Rated 10 814 3 605 29 014 14 008 13 511 21 218 21 218 17 115 8 997 21 013 3 488 16 411 16 412 12 103 5 510 22 010 4 865 15 005 15 005 19 870 15 826 11 107 7 004 32 414 21 599 9 789 9 642
Standby 9 056 3 488 528 821 791 2 022 3 370 3 370 1 788 0 0 1 612 1 612 1 964 1 084 2 491 1 172 4 513 9 730 10 668 15 035 4 044 6 829 762 528 4 484 1 553
Rate of Heat Gain, W Sensible Radiant 3165 1348 147 293 147 410 1055 1319 791 0 0 234 410 440 147 0 293 1846 4074 4250 3458 1319 2051 29 88 59 0
Usage Factor FU
Radiation Factor FR
0.84 0.97 0.02 0.06 0.06 0.10 0.16 0.20 0.20 0.00 0.00 0.10 0.10 0.16 0.20 0.11 0.24 0.30 0.65 0.54 0.95 0.36 0.97 0.02 0.02 0.46 0.16
0.35 0.39 0.28 0.36 0.19 0.20 0.31 0.39 0.44 0.00 0.00 0.15 0.25 0.22 0.14 0.00 0.25 0.41 0.42 0.40 0.23 0.33 0.30 0.04 0.17 0.01 0.00
*Items with an asterisk appear only in Swierczyna et al. (2009); all others appear in both Swierczyna et al. (2008) and (2009).
tions from the average values. Because large errors could occur in the heat load calculation for specific appliance lines by using a general radiation factor, heat gain values in Table 5 should be applied in the HVAC design. Table 5 lists usage factors, radiation factors, and load factors based on appliance energy consumption rate for typical electrical, steam, and gas appliances under standby (idle or ready-to-cook) and cooking conditions, hooded and unhooded.
Warewashing Applications. Typically, hot-water sanitizing and conveyor-type dish machines have either a dishwasher/condensing hood or direct-connected ductwork. If the ventilation is not operating properly, there are significant sensible and latent gains to the space. Chemical sanitizing and vapor reduction models are typically unhooded; consequently, the dish machines produce internal gains that must be accounted for and managed by the building HVAC system.
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18.10
2017 ASHRAE Handbook—Fundamentals (SI)
Table 5D Recommended Rates of Radiant Heat Gain from Hooded Gas Appliances during Idle (Ready-to-Cook) Conditions
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Appliance
Standby Energy Rate, W
Broiler: batch* Chain (conveyor) Overfired (upright)* Underfired 900 mm Fryer: doughnut Open deep-fat, 1 vat Pressure Griddle: double sided 900 mm, clamshell down* Clamshell up* Flat 900 mm Oven: combi: combi-mode* Convection mode Convection, full-size Conveyor (pizza) Deck Rack mini-rotating* Pasta cooker* Range top: top off/oven on* 3 burners on/oven off 6 burners on/oven off 6 burners on/oven on Range: wok* Rethermalizer* Rice cooker* Salamander* Steam kettle: large (225 L) simmer lid down* Small (38 L) simmer lid down* Medium (150 L) simmer lid down Steamer: compartment: atmospheric* Tilting skillet/braising pan
27 842 38 685 29 307 28 135 12 895 23 446 23 446 31 710 31 710 26 376 22 185 22 185 12 895 49 822 30 772 16 500 23 446 7327 35 169 35 169 42 495 29 014 26 376 10 257 10 257 42 495 15 240 29 307 7620 30 479
Rate of Heat Sensible Gain, W
20 280 28 340 25 761 21 658 3634 1377 2638 2345 4308 5979 1758 1700 3488 20 017 6008 1319 6946 2169 17 614 35 403 36 018 25 614 6829 147 9759 1583 967 1260 2432 3048
2374 3869 733 2638 850 322 234 528 1436 1084 117 293 293 2286 1026 322 0 586 2081 3370 3986 1524 3370 88 1553 0 88 0 0 117
Usage Factor FU
Radiation Factor FR
0.73 0.73 0.88 0.77 0.28 0.06 0.11 0.07 0.14 0.23 0.08 0.08 0.27 0.40 0.20 0.08 0.30 0.30 0.50 1.01 0.85 0.88 0.26 0.01 0.95 0.04 0.06 0.04 0.32 0.10
0.12 0.14 0.03 0.12 0.23 0.23 0.09 0.23 0.33 0.18 0.07 0.17 0.08 0.11 0.17 0.24 0.00 0.27 0.12 0.10 0.11 0.06 0.49 0.60 0.16 0.00 0.09 0.00 0.00 0.04
*Items with an asterisk appear only in Swierczyna et al. (2009); all others appear in both Swierczyna et al. (2008) and (2009).
Table 5E Recommended Rates of Radiant Heat Gain from Hooded Solid-Fuel Appliances during Idle (Ready-to-Cook) Conditions Appliance
Rated
Standby Energy Rate, W
Rate of Sensible Heat Gain, W
Usage Factor FU
Radiation Factor FR
Broiler: solid fuel: charcoal Broiler: solid fuel: wood (mesquite)
18 kg 18 kg
12 309 14 536
1817 2051
N/A N/A
0.15 0.14
Source: Swierczyna (2008).
Table 5F Recommended Rates of Radiant and Convective Heat Gain from Warewashing Equipment during Idle (Standby) or Washing Conditions Rate of Heat Gain, W
Energy Rate, W
Unhooded
Appliance
Standby/ Sensible Sensible Rated Washing Radiant Convective Latent Total
Dishwasher: conveyor type, hot-water sanitizing, washing Standby Dishwasher: conveyor type, chemical sanitizing, washing Standby Dishwasher: door type, hot-water sanitizing, washing With heat recovery and vapor reduction Standby Dishwasher: door type, chemical sanitizing, washing Standby Dishwasher: door type, chemical sanitizing, dump and fill, washing Standby Pot and pan washer: door type, hot-water sanitizing, washing With heat recovery and vapor reduction Dishwasher: under-counter type, hot-water sanitizing, washing With heat recovery and vapor reduction Standby Dishwasher: under-counter type, chemical sanitizing, washing Standby Booster heater
13,712 13,712 13,712 13,712 17,609 15,207 5,391 8,790 5,391 1,787 1,787 15,587 15,587 8,350 7,794 7,794 8,350 7,794 38,090
Sources: PG&E (2010-2016), Swierczyna et al. (2008) and (2009).
N/A 1,670 12,775 1,670 5,420 7,940 352 4,571 352 879 879 10,665 10,314 2,227 6,680 498 2,022 498 0
0 0 0 0 0 0 0 0 0 0 0 0 0 234 0 234 0 0 146
3,545 469 3,252 469 2,227 1,699 668 1,143 264 850 0 1,758 1,611 938 586 146 645 146 0
13,771 17,316 1,201 1,670 10,372 13,624 1,201 1,670 7,384 9,610 3,838 5,538 1,222 1,890 3,868 5,010 88 352 1,231 2,080 0 0 6,885 8,643 5,567 7,178 2,022 3,194 322 908 117 498 1,436 2,080 117 264 0 0
Hooded Usage Radiation Sensible Factor Factor Radiant FU FR 0 0 0 0 0 0 0 0 0 0 0 0 0 800 0 800 0 0 500
N/A 0.12 0.93 0.12 0.31 0.52 0.35 0.52 0.07 0.49 0.49 0.68 0.66 0.27 0.86 0.06 0.24 0.06 0
0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.11 0.00 0.47 0.00 0.00 N/A
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Nonresidential Cooling and Heating Load Calculations Sensible radiant and convective gains are affected by dishwasher insulation, and latent convective gains are affected by door seals. Heat loads may vary. Recirculating Systems. Cooking appliances ventilated by recirculating systems or “ductless” hoods should be treated as unhooded appliances when estimating heat gain. In other words, all energy consumed by the appliance and all moisture produced by cooking is introduced to the kitchen as a sensible or latent cooling load. Recommended Heat Gain Values. Table 5 lists recommended rates of heat gain from typical commercial cooking appliances. Data in the “hooded” columns assume installation under a properly designed exhaust hood connected to a mechanical fan exhaust system operating at an exhaust rate for complete capture and containment of the thermal and effluent plume. Improperly operating hood systems load the space with a significant convective component of the heat gain.
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Hospital and Laboratory Equipment Hospital and laboratory equipment items are major sources of sensible and latent heat gains in conditioned spaces. Care is needed in evaluating the probability and duration of simultaneous usage when many components are concentrated in one area, such as a laboratory, an operating room, etc. Commonly, heat gain from equipment in a laboratory ranges from 50 to 220 W/m2 or, in laboratories with outdoor exposure, as much as four times the heat gain from all other sources combined. Medical Equipment. It is more difficult to provide generalized heat gain recommendations for medical equipment than for general office equipment because medical equipment is much more varied in type and in application. Some heat gain testing has been done, but the equipment included represents only a small sample of the type of equipment that may be encountered. Data presented for medical equipment in Table 6 are relevant for portable and bench-top equipment. Medical equipment is very specific and can vary greatly from application to application. The data are presented to provide guidance in only the most general sense. For large equipment, such as MRI, heat gain must be obtained from the manufacturer. Laboratory Equipment. Equipment in laboratories is similar to medical equipment in that it varies significantly from space to space. Chapter 16 of the 2015 ASHRAE Handbook—HVAC Applications discusses heat gain from equipment, which may range from 50 to 270 W/m2 in highly automated laboratories. Table 7 lists some values for laboratory equipment, but, as with medical equipment, it is for general guidance only. Wilkins and Cook (1999) also examined laboratory equipment heat gains.
Office Equipment Computers, printers, copiers, etc., can generate very significant heat gains, sometimes greater than all other gains combined. ASHRAE research project RP-822 developed a method to measure the actual heat gain from equipment in buildings and the radiant/ convective percentages (Hosni et al. 1998; Jones et al. 1998). This methodology was then incorporated into ASHRAE research project RP-1055 and applied to a wide range of equipment (Hosni et al. 1999) as a follow-up to independent research by Wilkins and McGaffin (1994) and Wilkins et al. (1991). Komor (1997) found similar results. Analysis of measured data showed that results for office equipment could be generalized, but results from laboratory and hospital equipment proved too diverse. The following general guidelines for office equipment are a result of these studies. Nameplate Versus Measured Energy Use. Nameplate data rarely reflect the actual power consumption of office equipment. Actual power consumption is assumed to equal total (radiant plus convective) heat gain, but its ratio to the nameplate value varies widely. ASHRAE research project RP-1055 (Hosni et al. 1999)
18.11 Table 6 Recommended Heat Gain from Typical Medical Equipment Equipment Anesthesia system Blanket warmer Blood pressure meter Blood warmer ECG/RESP Electrosurgery Endoscope Harmonical scalpel Hysteroscopic pump Laser sonics Optical microscope Pulse oximeter Stress treadmill Ultrasound system Vacuum suction X-ray system
Nameplate, W
Peak, W
Average, W
250 500 180 360 1440 1000 1688 230 180 1200 330 72 N/A 1800 621 968 1725 2070
177 504 33 204 54 147 605 60 35 256 65 21 198 1063 337
166 221 29 114 50 109 596 59 34 229 63 20 173 1050 302 82 480 18
534
Source: Hosni et al. (1999).
Table 7 Recommended Heat Gain from Typical Laboratory Equipment Equipment Analytical balance Centrifuge Electrochemical analyzer Flame photometer Fluorescent microscope Function generator Incubator Orbital shaker Oscilloscope Rotary evaporator Spectronics Spectrophotometer Spectro fluorometer Thermocycler Tissue culture
Nameplate, W
Peak, W
7 138 288 5500 50 100 180 150 200 58 515 600 3125 100 72 345 75 94 36 575 200 N/A 340 1840 N/A 475 2346
7 89 136 1176 45 85 107 144 205 29 461 479 1335 16 38 99 74 29 31 106 122 127 405 965 233 132 1178
Average, W 7 87 132 730 44 84 105 143 178 29 451 264 1222 16 38 97 73 28 31 104 121 125 395 641 198 46 1146
Source: Hosni et al. (1999).
found that, for general office equipment with nameplate power consumption of less than 1000 W, the actual ratio of total heat gain to nameplate ranged from 25 to 50%, but when all tested equipment is considered, the range is broader. Generally, if the nameplate value is the only information known and no actual heat gain data are available for similar equipment, it is conservative to use 50% of nameplate as heat gain and more nearly correct if 25% of nameplate is used. Much better results can be obtained, however, by considering heat gain to be predictable based on the type of equipment. However, if the device has a mainly resistive internal electric load (e.g., a space heater), the nameplate rating may be a good estimate of its peak energy dissipation.
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18.12
2017 ASHRAE Handbook—Fundamentals (SI) Table 8A Recommended Heat Gain for Typical Desktop Computers
Description Manufacturer 1 3.0 GHz processor, 4 GB RAM, n = 1 3.3 GHz processor, 8 GB RAM, n = 8 3.5 GHz processor, 8 GB RAM, n = 2 3.6 GHz processor, 16 GB RAM, n = 2 3.3 GHz processor, 16 GB RAM, n = 2 4.0 GHz processor, 16 GB RAM, n = 1 3.3 GHz processor, 8 GB RAM, n = 1 3.7 GHz processor, 32 GB RAM, n = 1
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3.5 GHz processor, 16 GB RAM, n = 3c Manufacturer 2 3.6 GHz processor, 32 GB RAM, n = 8 3.6 GHz processor, 16 GB RAM, n = 1 3.4 GHz processor, 32 GB RAM, n = 1 3.4 GHz processor, 24 GB RAM, n = 1 3.50 GHz processor, 4 GB RAM, n = 1 3.3 GHz processor, 8 GB RAM, n = 1 3.20 GHz processor, 8 GB RAM, n = 1 3.20 GHz processor, 4 GB RAM, n = 1 2.93 GHz processor, 16 GB RAM, n = 1 2.67 GHz processor, 8 GB RAM, n = 1 Average 15-min peak power consumption (range)
Table 8B Recommended Heat Gain for Typical Laptops and Laptop Docking Station
Peak NameHeat plate Power,a Gain,b, d W W NA NA NA NA NA NA NA 750 NA 550 NA
83 50 42 66 52 83 84 116 102 144 93
NA NA NA NA NA NA NA NA NA NA
80 78 72 86 26 78 61 44 151 137
82 (26-151)
Source: Bach and Sarfraz (2017) n = number of tested equipment of same configuration. aNameplate for desktop computer is present on its power supply, which is mounted inside desktop, hence not accessible for most computers, where NA = not available. bFor equipment peak heat gain value, highest 15-min interval of recorded data is listed in tables. cFor tested equipment with same configuration, increasing power supply size does not increase average power consumption. dApproximately 90% convective heat gain and 10% radiative heat gain.
Computers. Based on tests by Hosni et al. (1999) and Wilkins and McGaffin (1994), nameplate values on computers should be ignored when performing cooling load calculations. Tables 8A, 8B, and 8C (Bach and Sarfraz 2017) present typical heat gain values for computers of varying types and models. Monitors. Table 8D shows typical values for various sizes and types. Flat-panel monitors have replaced CRT monitors in almost all workplaces. Power consumption, and thus heat gain, for flat-panel displays are significantly lower than for CRTs. Laser Printers. Hosni et al. (1999) found that power consumption, and therefore the heat gain, of laser printers depended largely on the level of throughput for which the printer was designed. Smaller printers tend to be used more intermittently, and larger printers may run continuously for longer periods. Table 9 presents data on typical printers. These data can be applied by taking the value for continuous operation and then applying an appropriate diversity factor. This would likely be most appropriate for larger open office areas. Another approach, which may be appropriate for a single room or small area, is to take the value that most closely matches the expected operation of the printer with no diversity. Copiers. Bach and Sarfraz (2017) also tested photocopy machines, including desktop and office (freestanding high-volume copiers) models. Larger machines used in production environments were not addressed. Table 9 summarizes the results. Desktop
Equipment Description Laptop computer
Manufacturer 1, 2.6 GHz processor, 8 GB RAM, n = 1 Manufacturer 2, 2.4 GHz processor, 4 GB RAM, n = 1
Average 15-min peak power consumption (range) Laptop with Manufacturer 1, docking 2.7 GHz processor, 8 GB RAM, n = 1 station 1.6 GHz processor, 8 GB RAM, n = 2 2.0 GHz processor, 8 GB RAM, n = 1 2.6 GHz processor, 4 GB RAM, n = 1 2.4 GHz processor, 8 GB RAM, n = 1 2.6 GHz processor, 8 GB RAM, n = 1 2.7 GHz processor, 8 GB RAM, n = 1 3.0 GHz processor, 8 GB RAM, n = 3 2.9 GHz processor, 32 GB RAM, n = 3 3.0 GHz processor, 32 GB RAM, n = 1 3.7 GHz processor, 32 GB RAM, n = 1 3.1 GHz processor, 32 GB RAM, n = 1 Average 15-min peak power consumption (range)
NamePeak plate Heat Power,a Gain,b, c W W NA
46
NA
59
53 (46-59) NA
38
NA NA NA NA NA NA NA NA NA NA NA
45 50 51 40 35 59 70 58 128 63 89
61 (26-151)
Source: Bach and Sarfraz (2017) n = number of tested equipment of same configuration. aVoltage and amperage information for laptop computer and laptop docking station is available on power supply nameplates; however, nameplate does not provide information on power consumption, where NA = not available. bFor equipment peak heat gain value, the highest 15-min interval of recorded data is listed in tables. cApproximately 75% convective heat gain and 25% radiative heat gain.
Table 8C Recommended Heat Gain for Typical Tablet PC
Description 1.7 GHz processor, 4 GB RAM, n = 1 2.2 GHz processor, 16 GB RAM, n = 1 2.3 GHz processor, 8 GB RAM, n = 1 2.5 GHz processor, 8 GB RAM, n = 1 Average 15-min peak power consumption (range)
Peak Nameplate Heat Power,a Gain,b W W NA NA NA NA
42 40 30 31
36 (31-42)
Source: Bach and Sarfraz (2017) n = number of tested equipment of same configuration. aVoltage and amperage information for tablet PC is available on power supply nameplate; however, nameplate does not provide information on power consumption, where NA = not available. bFor equipment peak heat gain value, highest 15-min interval of recorded data is listed in tables.
copiers rarely operate continuously, but office copiers frequently operate continuously for periods of an hour or more. Large, highvolume photocopiers often include provisions for exhausting air outdoors; if so equipped, the direct-to-space or system makeup air heat gain needs to be included in the load calculation. Also, when the air is dry, humidifiers are often operated near copiers to limit static electricity; if this occurs during cooling mode, their load on HVAC systems should be considered. Miscellaneous Office Equipment. Table 10 presents data on miscellaneous office equipment such as vending machines and other equipment tested by Bach and Sarfraz (2017). Diversity. The ratio of measured peak electrical load at equipment panels to the sum of the maximum electrical load of each individual item of equipment is the usage diversity. A small, one- or twoperson office containing equipment listed in Tables 8 to 10 usually
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Nonresidential Cooling and Heating Load Calculations Table 8D Recommended Heat Gain for Typical Monitors
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Descriptiona Manufacturer 1 1397 mm LED flat screen, n = 1 (excluded from average because atypical size) 686 mm LED flat screen, n = 2 546 mm LED flat screen, n = 2 Manufacturer 2 1270 mm 3D LED flat screen, n = 1 (excluded from average because atypical size) Manufacturer 3 864 mm LCD curved screen, n = 1 (excluded from average because atypical size and curved) 584 mm LED flat screen, n = 3 584 mm LED flat screen, n = 1 584 mm LED flat screen, n = 1 Manufacturer 4 610 mm LED flat screen, n = 1 Manufacturer 5 600 mm LED flat screen, n = 1 546 mm LED flat screen, n = 1 Manufacturer 6 546 mm LED flat screen, n = 1 Average 15-min peak power consumption (range)
Nameplate Power, W
Peak Heat Gain,b, c W
240
50
40 29
26 25
94
49
130
48
50 38 38
17 21 14
42
25
26 29
17 22
18.13 Table 9
Equipment
28
24
Recommended Heat Gain for Typical Printers
Description
Multifunction Large, multiuser, office type printer (copy, print, scan)
Average 15-min peak power consumption (range) Multiuser, medium-office type Desktop, small-office type Monochrome Desktop, medium-office type printer Average 15-min peak power consumption (range)
contributes heat gain to the space at the sum of the appropriate listed values. Progressively larger areas with many equipment items always experience some degree of usage diversity resulting from whatever percentage of such equipment is not in operation at any given time. Wilkins and McGaffin (1994) measured diversity in 23 areas within five different buildings totaling over 25 600 m2. Diversity was found to range between 37 and 78%, with the average (normalized based on area) being 46%. Figure 4 shows the relationship between nameplate, sum of peaks, and actual electrical load with diversity accounted for, based on the average of the total area tested. Data on actual diversity can be used as a guide, but diversity varies significantly with occupancy. The proper diversity factor for an office of call center operators is different from that for an office of sales representatives who travel regularly. ASHRAE research project RP-1093 derived diversity profiles for use in energy calculations (Abushakra et al. 2004; Claridge et al. 2004). Those profiles were derived from available measured data sets for a variety of office buildings, and indicated a range of peak weekday diversity factors for lighting ranging from 70 to 85% and for receptacles (appliance load) between 42 and 89%. Heat Gain per Unit Area. Bach and Sarfraz (2017), Wilkins and Hosni (2000, 2011) and Wilkins and McGaffin (1994) summarized research on a heat gain per unit area basis. Diversity testing showed that the actual heat gain per unit area, or load factor, ranged from 4.7 to 11.6 W/m2, with an average (normalized based on area) of 8.7 W/m2. Spaces tested were fully occupied and highly automated, comprising 21 unique areas in five buildings, with a computer and monitor at every workstation. Table 11 presents a range of load factors with a subjective description of the type of space to which they would apply. The medium load density is likely to be appropriate for most standard office spaces. Medium/heavy or heavy load densities
40
1010
30
1300
28
1500
Peak Heat Gain,a W 540 (Idle 29 W) 303 (Idle 116 W) 433 (Idle 28 W)
425 (303-540) 35
900
25
470
55 45
1000 680
732 (Idle 18 W) 56 (Idle 3 W) 222 61
142 (61-222)
Color printer Desktop, medium-office type
40
620
120
Laser printer Desktop, small-office type
14 24 26
310 495 1090
89 67 65
21 (14-26)
Source: Bach and Sarfraz (2017) n = number of tested equipment of same configuration. aScreens with atypical size and shape are excluded for calculating average 15-min peak power consumption. bFor equipment peak heat gain value, highest 15-min interval of recorded data is listed in tables. cApproximately 60% convective heat gain and 40% radiative heat gain.
Max. Printing Speed, NamePages plate per Power, W Minute
Average 15-min peak power consumption (range) Plotter
Manufacturer 1 Manufacturer 2 Average 15-min peak power consumption (range)
Fax machine Medium Small Average 15-min peak power consumption (range)
74 (65-89) 1600 270
571 173
372 (173-571) 1090 600
92 46
69 (46-92)
Source: Bach and Sarfraz (2017) aApproximately 70% convective heat gain and 30% radiative heat gain.
Fig. 4 Office Equipment Load Factor Comparison (Wilkins and McGaffin 1994)
may be encountered but can be considered extremely conservative estimates even for densely populated and highly automated spaces. Table 12 indicates applicable diversity factors. Radiant/Convective Split. ASHRAE research project RP-1482 (Hosni and Beck 2008) examined the radiant/convective split for common office equipment; the most important differentiating feature is whether the equipment had a cooling fan. Footnotes in Tables 8 and 9 summarize those results.
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18.14
2017 ASHRAE Handbook—Fundamentals (SI) Table 10 Recommended Heat Gain for Miscellaneous Equipment Nameplate Power,a W
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Equipment
Vending machine Drinks, 280 to 400 items NA Snacks NA Food (e.g., for sandwiches) NA Thermal binding machine, 2 single 350 documents up to 340 pages Projector, resolution 1024 768 340 Paper shredder, up to 28 sheets 1415 Electric stapler, up to 45 sheets NA Speakers 220 Temperature-controlled electronics 95 soldering station Cell phone charger NA Battery charger 40 V NA AA NA Microwave oven, 25 to 34 L 1000 to 1550 Coffee maker Single cup 1400 Up to 12 cups 950 With grinder 1350 Coffee grinder, up to 12 cups NA Tea kettle, up to 6 cups 1200 Dorm fridge, 88 L NA Freezer, 510 L 130 Fridge, 510 L NA Ice maker and dispenser, 9 kg bin capacity NA Top mounted bottled water cooler NA Cash register 25 Touch screen computer, 380 mm standard NA LCD and 2.2 GHz processor Self-checkout machine NA
Table 11 Recommended Load Factors for Various Types of Offices
Peak Heat Gain,b W Type of Use 940 54 465 28.5 308 265 1.5 15 16 5 19 5.5 713 to 822 385 780 376 73 1200 57 125 387 to 430 658 114 to 350 9 58 15
Source: Bach and Sarfraz (2017) aFor some equipment, nameplate power consumption is not available, where NA = not available. bFor equipment peak heat gain value, highest 15-min interval of recorded data is listed in tables.
3.
INFILTRATION AND MOISTURE MIGRATION HEAT GAINS
Two other load components contribute to space cooling load directly without time delay from building mass: (1) infiltration, and (2) moisture migration through the building envelope.
3.1
INFILTRATION
Principles of estimating infiltration in buildings, with emphasis on the heating season, are discussed in Chapter 16. When economically feasible, somewhat more outdoor air may be introduced to a building than the total of that exhausted, to create a slight overall positive pressure in the building relative to the outdoors. Under these conditions, air usually exfiltrates, rather than infiltrates, through the building envelope and thus effectively eliminates infiltration sensible and latent heat gains. However, there is concern, especially in some climates, that water may condense within the building envelope; actively managing space air pressures to reduce this condensation problem, as well as infiltration, may be needed. When positive air pressure is assumed, most designers do not include infiltration in cooling load calculations for commercial buildings. However, including some infiltration for spaces such entry areas or loading docks may be appropriate, especially when those
Load Factor*, W/m2 Description
100% laptop, docking station light 3.67 15.5 m2/workstation, all laptop docking station use, 1 printer per 10 medium 4.91 11.6 m2/workstation, all laptop docking station use, 1 printer per 10 50% laptop, docking station light 4.75 15.5 m2/workstation, 50% laptop docking station/50% desktop, 1 printer per 10 medium 6.35 11.6 m2/workstation, 50% laptop docking station/50% desktop, 1 printer per 10 100% desktop light 5.83 15.5 m2/workstation, all desktop use, 1 printer per 10 medium 7.79 11.6 m2/workstation, all desktop use, 1 printer per 10 100% laptop, docking station 2 screens 7.44 11.6 m2/workstation, all laptop docking station use, 2 screens, 1 printer per 10 100% desktop 2 screens 9.06 11.6 m2/workstation, all laptop use, 2 screens, 1 printer per 10 3 screens 10.33 11.6 m2/workstation, all desktop use, 3 screens, 1 printer per 10 100% desktop heavy, 2 screens 11.00 7.9 m2/workstation, all desktop use, 2 screens, 1 printer per 8 heavy, 3 screens 12.49 7.9 m2/workstation, all desktop use, 3 screens, 1 printer per 8 100% laptop, docking station full on, 2 screens 12.23 7.9 m2/workstation, all laptop docking use, 2 screens, 1 printer per 8, no diversity 100% desktop full on, 2 screens 14.35 7.9 m2/workstation, all desktop use, 2 screens, 1 printer per 8, no diversity full on, 3 screens 16.48 7.9 m2/workstation, all desktop use, 3 screens, 1 printer per 8, no diversity Source: Bach and Sarfraz (2017) *Medium office type monochrome printer is used for load factor calculator with 15min peak power consumption of 142 W.
Table 12
Diversity Factor for Different Equipment
Equipment Desktop PC Laptop docking station Notebook computer Screen Printer
Diversity Factor, %
Diversity Factor,a %
75 70 75b 70 45
75 NA 75 60 NA
Source: Bach and Sarfraz (2017) a2013 ASHRAE Handbook—Fundamentals bInsufficient data from RP-1742; values based on previous data from 2013 ASHRAE Handbook—Fundamentals and judgment of Bach and Sarfraz (2017).
spaces are on the windward side of buildings. But the downward stack effect, as occurs when indoor air is denser than the outdoor, might eliminate infiltration to these entries on lower floors of tall buildings; infiltration may occur on the upper floors during cooling conditions if makeup air is not sufficient. Infiltration also depends on wind direction and magnitude, temperature differences, construction type and quality, and occupant use of exterior doors and operable windows. As such, it is impossible to accurately predict infiltration rates. Designers usually predict overall rates of infiltration using the number of air changes per
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Nonresidential Cooling and Heating Load Calculations hour (ACH). A common guideline for climates and buildings typical of at least the central United States is to estimate the ACHs for winter heating conditions, and then use half that value for the cooling load calculations.
Standard Air Volumes Because the specific volume of air varies appreciably, calculations are more accurate when made on the basis of air mass instead of volume. However, volumetric flow rates are often required for selecting coils, fans, ducts, etc.; basing volumes on measurement at standard conditions may be used for accurate results. One standard value is 1.2 kgda /m3 (0.833 m3/kg). This density corresponds to about 16°C at saturation and 21°C dry air (at 101.325 kPa). Because air usually passes through the equipment at a density close to standard for locations below about 300 m, the accuracy desired normally requires no correction. When airflow is to be measured at a particular condition or point, such as at a coil entrance or exit, the corresponding specific volume can be read from the sea-level psychrometric chart. For higher elevations, the mass flow rates of air must be adjusted and higher-elevation psychrometric charts or algorithms must be used.
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Heat Gain Calculations Using Standard Air Values
18.15 This latent heat equation can also be expressed as ql = Cl Qs W where Cl = 3010 is the air latent heat factor, in W/(m3 ·s). 4. Elevation correction for total, sensible, and latent heat equations The constants 1.2, 1.23, and 3010 are useful in air-conditioning calculations at sea level (101.325 kPa) and for normal temperatures and moisture ratios. For other conditions, more precise values should be used. For an elevation of 1525 m (84.1 kPa), appropriate values are 1.00, 1.03, and 2500. Equations (8) to (10) can be corrected for elevations other than sea level by multiplying them by the ratio of pressure at sea level divided by the pressure at actual altitude. This can be derived from Equation (3) in Chapter 1 as Cx,alt = Cx,0 P/P0 where Cx,0 is any sea-level C value and P/P0 = [1 – (elevation 2.25577 10–5)]5.2559, where elevation is in metres.
Elevation Correction Examples To correct the C values for El Paso, Texas, the elevation listed in the appendix of Chapter 14 is 1194 m. C values for Equations (7) to (10) can be corrected using Equation (3) in Chapter 1 as follows:
Air-conditioning design often requires the following information:
Ct,1194 = 1.2 × [1 – (1194 × 2.25577 × 10–5)]5.2559 = 1.04
1. Total heat Total heat gain qt corresponding to the change of a given standard flow rate Qs through an enthalpy difference h is
Cs,1194 = 1.23 × [1 – (1194 × 2.25577 × 10–5)]5.2559 = 1.07
qt = 1.2Qs h
(7)
where 1.2 = kgda /m3. This total heat equation can also be expressed as qt = Ct Qs h 2. Sensible heat Sensible heat gain qs corresponding to the change of dry-bulb temperature t for given airflow (standard conditions) Qs is (8)
where 1.006 = specific heat of dry air, kJ/(kg·K) W = humidity ratio, kgw /kgda 1.84 = specific heat of water vapor, kJ/(kg·K)
The specific heats are for a range from about –75 to 90°C. When W = 0, the value of 1.20(1.006 + 1.84W) = 1.21; when W = 0.01, the value is 1.23; when W = 0.02, the value is 1.25; and when W = 0.03, the value is 1.27. Because a value of W = 0.01 approximates conditions found in many air-conditioning problems, the sensible heat change (in kW) has traditionally been found as qs = 1.23Qs t
(9)
This sensible heat equation can also be expressed as qs = Cs Qs t where Cs = 1.23 is the air sensible heat factor, in W/(m3 ·s·K). 3. Latent heat Latent heat gain ql corresponding to the change of humidity ratio W (in kgw/kgda) for given airflow (standard conditions) Qs is ql = 1.20 2500Qs W = 3010Qs W
To correct the C values for Albuquerque, New Mexico, the elevation listed in the appendix of Chapter 14 is 1619 m. C values for Equations (7) to (10) can be corrected as follows: Ct,1619 = 1.2 × [1 – (1619 × 2.25577 × 10–5)]5.2559 = 0.99
where Ct = 1.2 is the air total heat factor, in W/(m3 ·s) per kJ/kg.
qs = 1.2(1.006 + 1.84W )Qs t
Cl,1194 = 3010 × [1 – (1194 × 2.25577 × 10–5)]5.2559 = 2608
Cs,1619 = 1.23 × [1 – (1619 × 2.25577 × 10–5)]5.2559 = 1.01 Cl,1619 = 3010 × [1 – (1619 × 2.25577 × 10–5)]5.2559 = 2475
3.2
Diffusion of moisture through building materials is a natural phenomenon that is always present. Chapters 25 to 27 cover principles, materials, and specific methods used to control moisture. Moisture transfer through walls and roofs is often neglected in comfort air conditioning because the actual rate is quite small and the corresponding latent heat gain is insignificant. Permeability and permeance values for various building materials are given in Chapter 26. Vapor retarders should be specified and installed in the proper location to keep moisture transfer to a minimum, and to minimize condensation within the envelope. Moisture migration up through slabs-on-grade and basement floors has been found to be significant, but has historically not been addressed in cooling load calculations. Under-slab continuous moisture retarders and drainage can reduce upward moisture flow. Some industrial applications require low moisture to be maintained in a conditioned space. In these cases, the latent heat gain accompanying moisture transfer through walls and roofs may be greater than any other latent heat gain. This gain is computed by ql = MApv (hg –hf )
(10)
where 2500 is the approximate heat content of 50% rh vapor at 24°C less the heat content of water at 10°C. A common design condition for the space is 50% rh at 24°C, and 10°C is normal condensate temperature from cooling and dehumidifying coils.
LATENT HEAT GAIN FROM MOISTURE DIFFUSION
m
where ql = latent heat gain from moisture transfer, W m M = permeance of wall or roof assembly, ng/(s·m2 ·Pa) A = area of wall or roof surface, m2
(11)
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18.16 pv hg hf hg – hf
2017 ASHRAE Handbook—Fundamentals (SI) = = = =
vapor pressure difference, Pa enthalpy at room conditions, kJ/kg enthalpy of water condensed at cooling coil, kJ/kg 2500 kJ/kg when room temperature is 24°C and condensate off coil is 10°C
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3.3 OTHER LATENT LOADS Moisture sources within a building (e.g., shower areas, swimming pools or natatoriums, arboretums) can also contribute to latent load. Unlike sensible loads, which correlate to supply air quantities required in a space, latent loads usually only affect cooling coils sizing or refrigeration load. Because air from showers and some other moisture-generating areas is exhausted completely, those airborne latent loads do not reach the cooling coil and thus do not contribute to cooling load. However, system loads associated with ventilation air required to make up exhaust air must be recognized, and any recirculated air’s moisture must be considered when sizing the dehumidification equipment. For natatoriums, occupant comfort and humidity control are critical. In many instances, size, location, and environmental requirements make complete exhaust systems expensive and ineffective. Where recirculating mechanical cooling systems are used, evaporation (latent) loads are significant. Chapter 5 of the 2015 ASHRAE Handbook—HVAC Applications provides guidance on natatorium load calculations.
SHGC() = beam solar heat gain coefficient as a function of incident angle ; may be interpolated between values in Table 10 of Chapter 15 SHGCD = diffuse solar heat gain coefficient (also referred to as hemispherical SHGC); from Table 10 of Chapter 15 Tin = indoor temperature, °C Tout = outdoor temperature, °C U = overall U-factor, including frame and mounting orientation from Table 4 of Chapter 15, W/(m2 ·K) IAC(.) = indoor solar attenuation coefficient for beam solar heat gain coefficient; = 1.0 if no indoor shading device. IAC(.) is a function of shade type and, depending on type, may also be a function of beam solar angle of incidence and shade geometry IACD = indoor solar attenuation coefficient for diffuse solar heat gain coefficient; = 1.0 if not indoor shading device. IACD is a function of shade type and, depending on type, may also be a function of shade geometry
If specific window manufacturer’s SHGC and U-factor data are available, those should be used. For fenestration equipped with indoor shading (blinds, drapes, or shades), the indoor solar attenuation coefficients IAC(.) and IACD are listed in Tables 14A to 14G of Chapter 15. Note that, as discussed in Chapter 15, fenestration ratings (Ufactor and SHGC) are based on the entire product area, including frames. Thus, for load calculations, fenestration area is the area of the entire opening in the wall or roof.
4.2
4. FENESTRATION HEAT GAIN For spaces with neutral or positive air pressurization, the primary weather-related variable affecting cooling load is solar radiation. The effect of solar radiation is more pronounced and immediate on exposed, nonopaque surfaces. Chapter 14 includes procedures for calculating clear-sky solar radiation intensity and incidence angles for weather conditions encountered at specific locations. That chapter also includes some useful solar equations. Calculation of solar heat gain and conductive heat transfer through various glazing materials and associated mounting frames, with or without interior and/or exterior shading devices, is discussed in Chapter 15. This chapter covers application of such data to overall heat gain evaluation, and conversion of calculated heat gain into a composite cooling load for the conditioned space.
4.1
FENESTRATION DIRECT SOLAR, DIFFUSE SOLAR, AND CONDUCTIVE HEAT GAINS
For fenestration heat gain, use the following equations: Direct beam solar heat gain qb: qb = AEt,b SHGC()IAC(,)
(12)
Diffuse solar heat gain qd: qd = A(Et,d + Et,r)SHGCD IACD
(13)
Conductive heat gain qc: qc = UA(Tout – Tin)
(14)
Total fenestration heat gain Q: Q = qb + qd + qc
(15)
where A = window area, m2 Et,b, Et,d, and Et,r = beam, sky diffuse, and ground-reflected diffuse irradiance, calculated using equations in Chapter 14
EXTERIOR SHADING
Nonuniform exterior shading, caused by roof overhangs, side fins, or building projections, requires separate hourly calculations for the externally shaded and unshaded areas of the window in question, with the indoor shading SHGC still used to account for any internal shading devices. The areas, shaded and unshaded, depend on the location of the shadow line on a surface in the plane of the glass. Sun (1968) developed fundamental algorithms for analysis of shade patterns. McQuiston and Spitler (1992) provide graphical data to facilitate shadow line calculation. Equations for calculating shade angles [Chapter 15, Equations (34) to (37)] can be used to determine the shape and area of a moving shadow falling across a given window from external shading elements during the course of a design day. Thus, a subprofile of heat gain for that window can be created by separating its sunlit and shaded areas for each hour.
5.
HEAT BALANCE METHOD
Cooling load estimation involves calculating a surface-bysurface conductive, convective, and radiative heat balance for each room surface and a convective heat balance for the room air. These principles form the foundation for all methods described in this chapter. The heat balance (HB) method solves the problem directly instead of introducing transformation-based procedures. The advantages are that it contains no arbitrarily set parameters, and no processes are hidden from view. Some computations required by this rigorous approach require the use of computers. The heat balance procedure is not new. Many energy calculation programs have used it in some form for many years. The first implementation that incorporated all the elements to form a complete method was NBSLD (Kusuda 1967). The heat balance procedure is also implemented in both the BLAST and TARP energy analysis programs (Walton 1983). Before ASHRAE research project RP-875, the method had never been described completely or in a form applicable to cooling load calculations. The papers resulting from RP-875 describe the heat balance procedure in detail (Liesen and Pedersen 1997; McClellan and Pedersen 1997; Pedersen et al. 1997).
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Nonresidential Cooling and Heating Load Calculations The HB method is codified in the software called Hbfort that accompanies Cooling and Heating Load Calculation Principles (Pedersen et al. 1998). ASHRAE research project RP-1117 constructed two model rooms for which cooling loads were physically measured using extensive instrumentation (Chantrasrisalai et al. 2003; Eldridge et al. 2003; Iu et al. 2003). HB calculations closely approximated measured cooling loads when provided with detailed data for the test rooms.
5.1
ASSUMPTIONS
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All calculation procedures involve some kind of model; all models require simplifying assumptions and, therefore, are approximate. The most fundamental assumption is that air in the thermal zone can be modeled as well mixed, meaning its temperature is uniform throughout the zone. ASHRAE research project RP-664 (Fisher and Pedersen 1997) established that this assumption is valid over a wide range of conditions. The next major assumption is that the surfaces of the room (walls, windows, floor, etc.) can be treated as having • • • •
18.17 Wall Conduction Process The wall conduction process has been formulated in more ways than any of the other processes. Techniques include • • • •
Numerical finite difference Numerical finite element Transform methods Time series methods
This process introduces part of the time dependence inherent in load calculation. Figure 6 shows surface temperatures on the indoor and outdoor faces of the wall element, and corresponding conductive heat fluxes away from the outer face and toward the indoor face. All four quantities are functions of time. Direct formulation of the
Uniform surface temperatures Uniform long-wave (LW) and short-wave (SW) irradiation Diffuse radiating surfaces One-dimensional heat conduction within
The resulting formulation is called the heat balance (HB) model. Note that the assumptions, although common, are quite restrictive and set certain limits on the information that can be obtained from the model.
5.2
ELEMENTS
Within the framework of the assumptions, the HB can be viewed as four distinct processes: 1. 2. 3. 4.
Outdoor-face heat balance Wall conduction process Indoor-face heat balance Air heat balance
Figure 5 shows the relationship between these processes for a single opaque surface. The top part of the figure, inside the shaded box, is repeated for each surface enclosing the zone. The process for transparent surfaces is similar, but the absorbed solar component appears in the conduction process block instead of at the outdoor face, and the absorbed component splits into inward- and outwardflowing fractions. These components participate in the surface heat balances.
Fig. 5 Schematic of Heat Balance Processes in Zone
Outdoor-Face Heat Balance The heat balance on the outdoor face of each surface is qsol + qLWR + qconv – qko = 0
(16)
where qsol = absorbed direct and diffuse solar radiation flux (q/A), W/m2 qLWR = net long-wave radiation flux exchange with air and surroundings, W/m2 qconv = convective exchange flux with outdoor air, W/m2 qko = conductive flux (q/A) into wall, W/m2
All terms are positive for net flux to the face except qko, which is traditionally taken to be positive from outdoors to inside the wall. Each term in Equation (16) has been modeled in several ways, and in simplified methods the first three terms are combined by using the sol-air temperature.
Fig. 6 Schematic of Wall Conduction Process
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18.18 process uses temperature functions as input or known quantities, and heat fluxes as outputs or resultant quantities. In some models, surface heat transfer coefficients are included as part of the wall element, making the temperatures in question the indoor and outdoor air temperatures. This is not a desirable formulation, because it hides the heat transfer coefficients and prohibits changing them as airflow conditions change. It also prohibits treating the internal long-wave radiation exchange appropriately. Because heat balances on both sides of the element induce both the temperature and heat flux, the solution must deal with this simultaneous condition. Two computational methods that have been used widely are finite difference and conduction transfer function methods. Because of the computational time advantage, the conduction transfer function formulation has been selected for presentation here.
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Indoor-Face Heat Balance The heart of the HB method is the internal heat balance involving the inner faces of the zone surfaces. This heat balance has many heat transfer components, and they are all coupled. Both long-wave (LW) and short-wave (SW) radiation are important, as well as wall conduction and convection to the air. The indoor-face heat balance for each surface can be written as follows: qLWX + qSW + qLWS + qki + qsol + qconv = 0 (17) where qLWX = net long-wave radiant flux exchange between zone surfaces, W/m2 qSW = net short-wave radiation flux to surface from lights, W/m2 qLWS = long-wave radiation flux from equipment in zone, W/m2 qki = conductive flux through wall, W/m2 qsol = transmitted solar radiative flux absorbed at surface, W/m2 qconv = convective heat flux to zone air, W/m2
These terms are explained in the following paragraphs. LW Radiation Exchange Among Zone Surfaces. The limiting cases for modeling internal LW radiation exchange are • Zone air is completely transparent to LW radiation • Zone air completely absorbs LW radiation from surfaces in the zone Most HB models treat air as completely transparent and not participating in LW radiation exchange among surfaces in the zone. The second model is attractive because it can be formulated simply using a combined radiative and convective heat transfer coefficient from each surface to the zone air and thus decouples radiant exchange among surfaces in the zone. However, because the transparent air model allows radiant exchange and is more realistic, the second model is inferior. Furniture in a zone increases the amount of surface area that can participate in radiative and convective heat exchanges. It also adds thermal mass to the zone. These two changes can affect the time response of the zone cooling load. SW Radiation from Lights. The short-wavelength radiation from lights is usually assumed to be distributed over the surfaces in the zone in some manner. The HB procedure retains this approach but allows the distribution function to be changed. LW Radiation from Internal Sources. The traditional model for this source defines a radiative/convective split for heat introduced into a zone from equipment. The radiative part is then distributed over the zone’s surfaces in some manner. This model is not completely realistic, and it departs from HB principles. In a true HB model, equipment surfaces are treated just as other LW radiant sources in the zone. However, because information about the surface temperature of equipment is rarely known, it is reasonable to keep the radiative/convective split concept even though it ignores the true nature of the radiant exchange. ASHRAE research project
2017 ASHRAE Handbook—Fundamentals (SI) RP-1055 (Hosni et al. 1999) determined radiative/convective splits for many additional equipment types, as listed in footnotes for Tables 8 and 9. Transmitted Solar Heat Gain. Chapter 15’s calculation procedure for determining transmitted solar energy through fenestration uses the solar heat gain coefficient (SHGC) directly rather than relating it to double-strength glass, as is done when using a shading coefficient (SC). The difficulty with this plan is that the SHGC includes both transmitted solar and inward-flowing fraction of the solar radiation absorbed in the window. With the HB method, this latter part should be added to the conduction component so it can be included in the indoor-face heat balance. Transmitted solar radiation is also distributed over surfaces in the zone in a prescribed manner. It is possible to calculate the actual position of beam solar radiation, but this involves partial surface irradiation, which is inconsistent with the rest of the zone model, which assumes uniform conditions over an entire surface.
Using SHGC to Calculate Solar Heat Gain The total solar heat gain through fenestration consists of directly transmitted solar radiation plus the inward-flowing fraction of solar radiation that is absorbed in the glazing system. Both parts contain beam and diffuse contributions. Transmitted radiation goes directly onto surfaces in the zone and is accounted for in the surface indoor heat balance. The zone heat balance model accommodates the resulting heat fluxes without difficulty. The second part, the inwardflowing fraction of the absorbed solar radiation, interacts with other surfaces of the enclosure through long-wave radiant exchange and with zone air through convective heat transfer. As such, it depends both on geometric and radiative properties of the zone enclosure and convection characteristics inside and outside the zone. The solar heat gain coefficient (SHGC) combines the transmitted solar radiation and the inward-flowing fraction of the absorbed radiation. The SHGC is defined as SHGC = + N k k n
(18)
k=1
where k n Nk
= = = =
solar transmittance of glazing solar absorptance of the kth layer of the glazing system number of layers inward-flowing fraction of absorbed radiation in the kth layer
Note that Equation (18) is written generically. It can be written for a specific incidence angle and/or radiation wavelength and integrated over the wavelength and/or angle, but the principle is the same in each case. Refer to Chapter 15 for the specific expressions. Unfortunately, the inward-flowing fraction N interacts with the zone in many ways. This interaction can be expressed as N = f (indoor convection coefficient, outdoor convection coefficient, glazing system overall heat transfer coefficient, zone geometry, zone radiation properties)
The only way to model these interactions correctly is to combine the window model with the zone heat balance model and solve both simultaneously. This has been done recently in some energy analysis programs, but is not generally available in load calculation procedures. In addition, the SHGC used for rating glazing systems is based on specific values of the indoor, outdoor, and overall heat transfer coefficients and does not include any zonal long-wavelength radiation considerations. So, the challenge is to devise a way to use SHGC values within the framework of heat balance calculation in the most accurate way possible, as discussed in the following paragraphs. Using SHGC Data. The normal incidence SHGC used to rate and characterize glazing systems is not sufficient for determining solar heat gain for load calculations. These calculations require
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Nonresidential Cooling and Heating Load Calculations
18.19
Table 13 Single-Layer Glazing Data Produced by WINDOW 7.4.6 Incident Angle Parameter Vt c Rfv
0
10
20
30
40
50
60
70
80
90
Diffuse (Hemis.)
0.899 0.083
0.899 0.083
0.898 0.083
0.896 0.085
0.889 0.091
0.870 0.109
0.822 0.156
0.705 0.272
0.441 0.536
0 1
0.822 0.148
Rbv
0.083
0.083
0.083
0.085
0.091
0.109
0.156
0.272
0.536
1
0.148
Tsol
0.834
0.833
0.831
0.827
0.818
0.797
0.749
0.637
0.389
0
0.753
Rf
0.075
0.075
0.075
0.077
0.082
0.099
0.143
0.253
0.506
1
0.136
Rb
0.075
0.075
0.075
0.077
0.082
0.099
0.143
0.253
0.506
1
0.136
Abs1 SHGC
0.091 0.861
0.092 0.860
0.094 0.859
0.096 0.855
0.100 0.847
0.104 0.827
0.108 0.781
0.110 0.669
0.105 0.424
0 0
0.101 0.783
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Source: LBL (2015).
solar heat gain as a function of the incident solar angle to determine the hour-by-hour gain profile. Thus, it is necessary to use angular SHGC values and also diffuse SHGC values. These can be obtained from the WINDOW 7.4.6 program (LBL 2015). This program does a detailed optical and thermal simulation of a glazing system and, when applied to a single clear layer, produces the information shown in Table 13. Table 13 shows the parameters as a function of incident solar angle and also the diffuse values. The specific parameters shown are Vtc = transmittance in visible spectrum Rf v and Rbv = front and back surface visible reflectances Tsol = solar transmittance [ in Equations (18), (19), and (20)] Rf and Rb = front and back surface solar reflectances Abs1 = solar absorptance for layer 1, which is the only layer in this case [ in Equations (18), (19), and (20)] SHGC = solar heat gain coefficient at center of glazing
The parameters used for heat gain calculations are Tsol , Abs , and SHGC. For the specific convective conditions assumed in WINDOW 7.4.6 program, the inward-flowing fraction of the absorbed solar can be obtained by rearranging Equation (18) to give Nk k = SHGC –
(19)
This quantity, when multiplied by the appropriate incident solar intensity, provides the amount of absorbed solar radiation that flows inward. In the heat balance formulation for zone loads, this heat flux is combined with that caused by conduction through glazing and included in the surface heat balance. The outward-flowing fraction of absorbed solar radiation is used in the heat balance on the outdoor face of the glazing and is determined from (1 – Nk) k = k – Nk k = k – (SHGC – )
(20)
If there is more than one layer, the appropriate summation of absorptances must be done. There is some potential inaccuracy in using the WINDOW 7.4.6 SHGC values because the inward-flowing fraction part was determined under specific conditions for the indoor and outdoor heat transfer coefficients. However, the program can be run with indoor and outdoor coefficients of one’s own choosing. Normally, however, this effect is not large, and only in highly absorptive glazing systems might cause significant error. For solar heat gain calculations, then, it seems reasonable to use the generic window property data that comes from WINDOW 7.4.6. Considering Table 13, the procedure is as follows: 1. Determine angle of incidence for the glazing. 2. Determine corresponding SHGC. 3. Evaluate Nk k using Equation (18).
4. Multiply Tsol by incident beam radiation intensity to get transmitted beam solar radiation. 5. Multiply Nk k by incident beam radiation intensity to get inward-flowing absorbed heat. 6. Repeat steps 2 to 5 with diffuse parameters and diffuse radiation. 7. Add beam and diffuse components of transmitted and inwardflowing absorbed heat. This procedure is incorporated into the HB method so the solar gain is calculated accurately for each hour. Table 10 in Chapter 15 contains SHGC information for many additional glazing systems. That table is similar to Table 13 but is slightly abbreviated. Again, the information needed for heat gain calculations is Tsol , SHGC, and Abs . The same caution about the indoor and outdoor heat transfer coefficients applies to the information in Table 10 in Chapter 15. Those values were also obtained with specific indoor and outdoor heat transfer coefficients, and the inward-flowing fraction N is dependent upon those values. Convection to Zone Air. Indoor convection coefficients presented in past editions of this chapter and used in most load calculation procedures and energy programs are based on very old, natural convection experiments and do not accurately describe heat transfer coefficients in a mechanically ventilated zone. In previous load calculation procedures, these coefficients were buried in the procedures and could not be changed. A heat balance formulation keeps them as working parameters. In this way, research results such as those from ASHRAE research project RP-664 (Fisher 1998) can be incorporated into the procedures. It also allows determining the sensitivity of the load calculation to these parameters.
Air Heat Balance In HB formulations aimed at determining cooling loads, the capacitance of air in the zone is neglected and the air heat balance is done as a quasisteady balance in each time period. Four factors contribute to the air heat balance: qconv + qCE + qIV + qsys = 0
(21)
where qconv qCE qIV qsys
= = = =
convective heat transfer from surfaces, W convective parts of internal loads, W sensible load caused by infiltration and ventilation air, W heat transfer to/from HVAC system, W
Convection from zone surfaces qconv is the sum of all the convective heat transfer quantities from the indoor-surface heat balance. This comes to the air through the convective heat transfer coefficient on the surfaces. The convective parts of the internal loads qCE is the companion to qL WS , the radiant contribution from internal loads [Equation (17)]. It is added directly to the air heat balance. This also violates the tenets of the HB approach, because surfaces producing internal
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18.20
2017 ASHRAE Handbook—Fundamentals (SI)
loads exchange heat with zone air through normal convective processes. However, once again, this level of detail is generally not included in the heat balance, so it is included directly into the air heat balance instead. In keeping with the well-mixed model for zone air, any air that enters directly to a space through infiltration or ventilation qIV is immediately mixed with the zone’s air. The amount of infiltration or natural ventilation air is uncertain. Sometimes it is related to the indoor/outdoor temperature difference and wind speed; however it is determined, it is added directly to the air heat balance. Conditioned air that enters the zone from the HVAC system and provides qsys is also mixed directly with the zone air. For commercial HVAC systems, ventilation air is most often provided using outdoor air as part of this mixed-in conditioned air; ventilation air is thus normally a system load rather than a direct-to-space load. An exception is where infiltration or natural ventilation is used to provide all or part of the ventilation air, as discussed in Chapter 16.
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5.3
GENERAL ZONE FOR LOAD CALCULATION
The HB procedure is tailored to a single thermal zone, shown in Figure 7. The definition of a thermal zone depends on how the fixed temperature is controlled. If air circulated through an entire building or an entire floor is uniformly well stirred, the entire building or floor could be considered a thermal zone. On the other hand, if each room has a different control scheme, each room may need to be considered as a separate thermal zone. The framework needs to be flexible enough to accommodate any zone arrangement, but the heat balance aspect of the procedure also requires that a complete zone be described. This zone consists of four walls, a roof or ceiling, a floor, and a “thermal mass surface” (described in the section on Input Required). Each wall and the roof can include a window (or skylight in the case of the roof). This makes a total of 12 surfaces, any of which may have zero area if it is not present in the zone to be modeled. The heat balance processes for this general zone are formulated for a 24 h steady-periodic condition. The variables are the indoor and outdoor temperatures of the 12 surfaces plus either the HVAC system energy required to maintain a specified air temperature or the air temperature, if system capacity is specified. This makes a total of 25 24 = 600 variables. Although it is possible to set up the problem for a simultaneous solution of these variables, the relatively weak coupling of the problem from one hour to the next allows a double iterative approach. One iteration is through all the surfaces in each hour, and the other is through the 24 h of a day. This procedure automatically reconciles nonlinear aspects of surface radiative exchange and other heat flux terms.
5.4
MATHEMATICAL DESCRIPTION
Conduction Process Because it links the outdoor and indoor heat balances, the wall conduction process regulates the cooling load’s time dependence. For the HB procedure presented here, wall conduction is formulated using conduction transfer functions (CTFs), which relate conductive heat fluxes to current and past surface temperatures and past heat fluxes. The general form for the indoor heat flux is qki t = – Z o Tsi, – Z j Tsi, – j nz
j=1
+ Y o Tso, + Y j T so, – j + j q ki,–j nz
nq
j=1
j=1
(22)
For outdoor heat flux, the form is qko t = – Y o T si, – Y j T si, – j nz
j=1
+ X o T so, + X j T so, – j + j q ko,–j nz
nq
j=1
j=1
(23)
where Xj Yj Zj j Tsi Tso qki qko
= = = = =
= = = = =
outdoor CTF, j = 0,1,…nz cross CTF, j = 0,1,…nz indoor CTF, j = 0,1,…nz flux CTF, j = 1,2,…nq time time step indoor-face temperature, °C outdoor-face temperature, °C conductive heat flux on indoor face, W/m2 conductive heat flux on outdoor face, W/m2
The subscript following the comma indicates the time period for the quantity in terms of time step . Also, the first terms in the series have been separated from the rest to facilitate solving for the current temperature in the solution scheme. The two summation limits nz and nq depend on wall construction and also somewhat on the scheme used for calculating the CTFs. If nq = 0, the CTFs are generally referred to as response factors, but then theoretically nz is infinite. Values for nz and nq are generally set to minimize the amount of computation. A development of CTFs can be found in Hittle and Pedersen (1981).
Heat Balance Equations The primary variables in the heat balance for the general zone are the 12 indoor face temperatures and the 12 outdoor face temperatures at each of the 24 h, assigning i as the surface index and j as the hour index, or, in the case of CTFs, the sequence index. Thus, the primary variables are Tsoi,j = outdoor face temperature, i = 1,2,…,12; j = 1,2,…, 24 Tsii,j = indoor face temperature, i = 1,2,…,12; j = 1,2,…, 24
Fig. 7 Schematic View of General Heat Balance Zone
In addition, qsysj = cooling load, j = 1,2,…, 24. Equations (16) and (23) are combined and solved for Tso to produce 12 equations applicable in each time step:
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Nonresidential Cooling and Heating Load Calculations
T so
i j
nz nq nz = T si Y i k – T so Z i , k – i k qko i j – k i j – k i j – k k=1 k=1 k=1
+ qsol
i j
+ qLWR
i j
+ T si Y i 0 + T o h co Z i 0 + h co i j j i j i j
18.21 For Day = 1 to Maxdays For j = 1 to 24 {hours in the day} For SurfaceIter = 1 to MaxIter For i = 1 to 12 {The twelve zone surfaces} Evaluate Equations (33) and (34) Next i Next SurfaceIter Evaluate Equation (35) Next j If not converged, Next Day
(24)
where To = outdoor air temperature hco = outdoor convection coefficient, introduced by using qconv = hco (To – Tso)
8. Display results.
Equation (24) shows the need to separate Zi,0, because the contribution of current surface temperature to conductive flux can be collected with the other terms involving that temperature. Equations (17) and (22) are combined and solved for Tsi to produce the next 12 equations: = T Y + T so Y i, k i, j – k si i, j i, 0 k–1
Generally, four or six surface iterations are sufficient to provide convergence. The convergence check on the day iteration should be based on the difference between the indoor and outdoor conductive heat flux terms qk . A limit, such as requiring the difference between all indoor and outdoor flux terms to be less than 1% of either flux, works well.
nz
T si
i, j
– T si
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nz
i, j – k
k=1
Z i, k +
5.5
i, k qki nq
k=1
i, j – k
+ T a h ci + q LWS j
(25)
j
+ qLWX + qSW + qsol e Z i, 0 + h ci i, j where Ta = zone air temperature hci = convective heat transfer coefficient indoors, obtained from qconv = hci (Ta – Tsi )
Note that in Equations (24) and (25), the opposite surface temperature at the current time appears on the right-hand side. The two equations could be solved simultaneously to eliminate those variables. Depending on the order of updating the other terms in the equations, this can have a beneficial effect on solution stability. The remaining equation comes from the air heat balance, Equation (21). This provides the cooling load qsys at each time step: q sys = j
Ai hci Tsi 12
i, j
– T a + q CE + q IV j
(26)
i=1
In Equation (26), the convective heat transfer term is expanded to show the interconnection between the surface temperatures and the cooling load.
Overall HB Iterative Solution The iterative HB procedure consists of a series of initial calculations that proceed sequentially, followed by a double iteration loop, as shown in the following steps: 1. Initialize areas, properties, and face temperatures for all surfaces, 24 h. 2. Calculate incident and transmitted solar flux for all surfaces and hours. 3. Distribute transmitted solar energy to all indoor faces, 24 h. 4. Calculate internal load quantities for all 24 h. 5. Distribute LW, SW, and convective energy from internal loads to all surfaces for all hours. 6. Calculate infiltration and direct-to-space ventilation loads for all hours. 7. Iterate the heat balance according to the following scheme:
INPUT REQUIRED
Previous methods for calculating cooling loads attempted to simplify the procedure by precalculating representative cases and grouping the results with various correlating parameters. This generally tended to reduce the amount of information required to apply the procedure. With heat balance, no precalculations are made, so the procedure requires a fairly complete description of the zone. Global Information. Because the procedure incorporates a solar calculation, some global information is required, including latitude, longitude, time zone, month, day of month, directional orientation of the zone, and zone height (floor to floor). Additionally, to take full advantage of the flexibility of the method to incorporate, for example, variable outdoor heat transfer coefficients, things such as wind speed, wind direction, and terrain roughness may be specified. Normally, these variables and others default to some reasonable set of values, but the flexibility remains. Wall Information (Each Wall). Because the walls are involved in three of the fundamental processes (external and internal heat balance and wall conduction), each wall of the zone requires a fairly large set of variables. They include • • • • • • • • • •
Facing angle with respect to solar exposure Tilt (degrees from horizontal) Area Solar absorptivity outdoors Long-wave emissivity outdoors Short-wave absorptivity indoors Long-wave emissivity indoors Exterior boundary temperature condition (solar versus nonsolar) External roughness Layer-by-layer construction information
Again, some of these parameters can be defaulted, but they are changeable, and they indicate the more fundamental character of the HB method because they are related to true heat transfer processes. Window Information (Each Window). The situation for windows is similar to that for walls, but the windows require some additional information because of their role in the solar load. Necessary parameters include • • • • • • •
Area Normal solar transmissivity Normal SHGC Normal total absorptivity Long-wave emissivity outdoors Long-wave emissivity indoor Surface-to-surface thermal conductance
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18.22
Licensed for single user. © 2017 ASHRAE, Inc.
• Reveal (for solar shading) • Overhang width (for solar shading) • Distance from overhang to window (for solar shading) Roof and Floor Details. The roof and floor surfaces are specified similarly to walls. The main difference is that the ground outdoor boundary condition will probably be specified more often for a floor. Thermal Mass Surface Details. An “extra” surface, called a thermal mass surface, can serve several functions. It is included in radiant heat exchange with the other surfaces in the space but is only exposed to the indoor air convective boundary condition. As an example, this surface would be used to account for movable partitions in a space. Partition construction is specified layer by layer, similar to specification for walls, and those layers store and release heat by the same conduction mechanism as walls. As a general definition, the extra thermal mass surface should be sized to represent all surfaces in the space that are exposed to the air mass, except the walls, roof, floor, and windows. In the formulation, both sides of the thermal mass participate in the exchange. Internal Heat Gain Details. The space can be subjected to several internal heat sources: people, lights, electrical equipment, and infiltration. Infiltration energy is assumed to go immediately into the air heat balance, so it is the least complicated of the heat gains. For the others, several parameters must be specified. These include the following fractions: • • • • • • •
Sensible heat gain Latent heat gain Short-wave radiation Long-wave radiation Energy that enters the air immediately as convection Activity level of people Lighting heat gain that goes directly to the return air
Radiant Distribution Functions. As mentioned previously, the generally accepted assumptions for the HB method include specifying the distribution of radiant energy from several sources to surfaces that enclose the space. This requires a distribution function that specifies the fraction of total radiant input absorbed by each surface. The types of radiation that require distribution functions are • Long-wave, from equipment and lights • Short-wave, from lights • Transmitted solar Other Required Information. Additional flexibility is included in the model so that results of research can be incorporated easily. This includes the capability to specify such things as • Heat transfer coefficients/convection models • Solar coefficients • Sky models The amount of input information required may seem extensive, but many parameters can be set to default values in most routine applications. However, all parameters listed can be changed when necessary to fit unusual circumstances or when additional information is obtained.
6. RADIANT TIME SERIES (RTS) METHOD The radiant time series (RTS) method is a simplified method for performing design cooling load calculations that is derived from the heat balance (HB) method. It effectively replaces all other simplified (non-heat-balance) methods, such as the transfer function method (TFM), the cooling load temperature difference/cooling
2017 ASHRAE Handbook—Fundamentals (SI) load factor (CLTD/CLF) method, and the total equivalent temperature difference/time averaging (TETD/TA) method. This method was developed to offer an approach that is rigorous, yet does not require iterative calculations, and that quantifies each component’s contribution to the total cooling load. In addition, it is desirable for the user to be able to inspect and compare the coefficients for different construction and zone types in a form showing their relative effect on the result. These characteristics of the RTS method make it easier to apply engineering judgment during cooling load calculation. The RTS method is suitable for peak design load calculations, but it should not be used for annual energy simulations because of its inherent limiting assumptions. Although simple in concept, RTS involves too many calculations for practical use as a manual method, although it can easily be implemented in a simple computerized spreadsheet, as shown in the examples. For a manual cooling load calculation method, refer to the CLTD/CLF method in Chapter 28 of the 1997 ASHRAE Handbook—Fundamentals.
6.1
ASSUMPTIONS AND PRINCIPLES
Design cooling loads are based on the assumption of steadyperiodic conditions (i.e., the design day’s weather, occupancy, and heat gain conditions are identical to those for preceding days such that the loads repeat on an identical 24 h cyclical basis). Thus, the heat gain for a particular component at a particular hour is the same as 24 h prior, which is the same as 48 h prior, etc. This assumption is the basis for the RTS derivation from the HB method. Cooling load calculations must address two time-delay effects inherent in building heat transfer processes: • Delay of conductive heat gain through opaque massive exterior surfaces (walls, roofs, or floors) • Delay of radiative heat gain conversion to cooling loads. Exterior walls and roofs conduct heat because of temperature differences between outdoor and indoor air. In addition, solar energy on exterior surfaces is absorbed, then transferred by conduction to the building interior. Because of the mass and thermal capacity of the wall or roof construction materials, there is a substantial time delay in heat input at the exterior surface becoming heat gain at the interior surface. As described in the section on Cooling Load Principles, most heat sources transfer energy to a room by a combination of convection and radiation. The convective part of heat gain immediately becomes cooling load. The radiative part must first be absorbed by the finishes and mass of the interior room surfaces, and becomes cooling load only when it is later transferred by convection from those surfaces to the room air. Thus, radiant heat gains become cooling loads over a delayed period of time.
6.2
OVERVIEW
Figure 8 gives an overview of the RTS method. When calculating solar radiation, transmitted solar heat gain through windows, sol-air temperature, and infiltration, RTS is exactly the same as previous simplified methods (TFM and TETD/TA). Important areas that differ from previous simplified methods include • Computation of conductive heat gain • Splitting of all heat gains into radiant and convective portions • Conversion of radiant heat gains into cooling loads The RTS method accounts for both conduction time delay and radiant time delay effects by multiplying hourly heat gains by 24 h time series. The time series multiplication, in effect, distributes heat gains over time. Series coefficients, which are called radiant time factors and conduction time factors, are derived using the HB method. Radiant time factors reflect the percentage of an earlier
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Licensed for single user. © 2017 ASHRAE, Inc.
Nonresidential Cooling and Heating Load Calculations
18.23
Fig. 8 Overview of Radiant Time Series Method
Fig. 9 CTS for Light to Heavy Walls radiant heat gain that becomes cooling load during the current hour. Likewise, conduction time factors reflect the percentage of an earlier heat gain at the exterior of a wall or roof that becomes heat gain indoors during the current hour. By definition, each radiant or conduction time series must total 100%. These series can be used to easily compare the time-delay effect of one construction versus another. This ability to compare choices is of particular benefit during design, when all construction details may not have been decided. Comparison can show the magnitude of difference between the choices, allowing the engineer to apply judgment and make more informed assumptions in estimating the load. Figure 9 shows conduction time series (CTS) values for three walls with similar U-factors but with light to heavy construction. Figure 10 shows CTS for three walls with similar construction but with different amounts of insulation, thus with significantly different U-factors. Figure 11 shows RTS values for zones varying from light to heavy construction.
Fig. 10
CTS for Walls with Similar Mass and Increasing Insulation
Fig. 11 RTS for Light to Heavy Construction
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18.24
2017 ASHRAE Handbook—Fundamentals (SI) Table 14
Recommended Radiative/Convective Splits for Internal Heat Gains
Recommended Recommended Radiative Fraction Convective Fraction Comments
Heat Gain Type Occupants, typical office conditions Equipment Office, with fan Without fan Lighting Conduction heat gain Through walls and floors Through roof Through windows Solar heat gain through fenestration Without interior shading With interior shading Infiltration
0.60 0.1 to 0.8 0.10 0.30
0.40 0.9 to 0.2 0.90 0.70
See Table 1 for other conditions. See Tables 6 to 12 for details of equipment heat gain and recommended radiative/convective splits for motors, cooking appliances, laboratory equipment, medical equipment, office equipment, etc. Varies; see Table 3.
0.46 0.60 0.33 (SHGC > 0.5) 0.46 (SHGC < 0.5)
0.54 0.40 0.67 (SHGC > 0.5) 0.54 (SHGC < 0.5)
1.00
0.00
0.00
1.00
Varies; see Tables 14A to 14G in Chapter 15.
Source: Nigusse (2007).
Licensed for single user. © 2017 ASHRAE, Inc.
6.3
RTS PROCEDURE
The general procedure for calculating cooling load for each load component (lights, people, walls, roofs, windows, appliances, etc.) with RTS is as follows: 1. Calculate 24 h profile of component heat gains for design day (for conduction, first account for conduction time delay by applying conduction time series). 2. Split heat gains into radiant and convective parts (see Table 14 for radiant and convective fractions). 3. Apply appropriate radiant time series to radiant part of heat gains to account for time delay in conversion to cooling load. 4. Sum convective part of heat gain and delayed radiant part of heat gain to determine cooling load for each hour for each cooling load component. After calculating cooling loads for each component for each hour, sum those to determine the total cooling load for each hour and select the hour with the peak load for design of the air-conditioning system. Repeat this process for multiple design months to determine the month when the peak load occurs, especially with windows on southern exposures (northern exposure in southern latitudes), which can result in higher peak room cooling loads in winter months than in summer.
6.4
HEAT GAIN THROUGH EXTERIOR SURFACES
Heat gain through exterior opaque surfaces is derived from the same elements of solar radiation and thermal gradient as that for fenestration areas. It differs primarily as a function of the mass and nature of the wall or roof construction, because those elements affect the rate of conductive heat transfer through the composite assembly to the interior surface.
Sol-Air Temperature Sol-air temperature is the outdoor air temperature that, in the absence of all radiation changes gives the same rate of heat entry into the surface as would the combination of incident solar radiation, radiant energy exchange with the sky and other outdoor surroundings, and convective heat exchange with outdoor air. Heat Flux into Exterior Sunlit Surfaces. The heat balance at a sunlit surface gives the heat flux into the surface q/A as
q --- = Et + ho(to – ts) – R A
(27)
where = absorptance of surface for solar radiation Et = total solar radiation incident on surface, W/m2 ho = coefficient of heat transfer by long-wave radiation and convection at outer surface, W/(m2 ·K) to = outdoor air temperature, °C ts = surface temperature, °C = hemispherical emittance of surface R = difference between long-wave radiation incident on surface from sky and surroundings and radiation emitted by blackbody at outdoor air temperature, W/m2
Assuming the rate of heat transfer can be expressed in terms of the sol-air temperature te , q --- = ho(te – ts) A
(28)
and from Equations (27) and (28), E t R te = to + ---------- – -----------ho ho
(29)
For horizontal surfaces that receive long-wave radiation from the sky only, an appropriate value of R is about 63 W/m2, so that if = 1 and ho = 17 W/(m2·K), the long-wave correction term is about 4 K (Bliss 1961). Because vertical surfaces receive long-wave radiation from the ground and surrounding buildings as well as from the sky, accurate R values are difficult to determine. When solar radiation intensity is high, surfaces of terrestrial objects usually have a higher temperature than the outdoor air; thus, their long-wave radiation compensates to some extent for the sky’s low emittance. Therefore, it is common practice to assume R = 0 for vertical surfaces. Tabulated Temperature Values. The sol-air temperatures in Example Cooling and Heating Load Calculations section have been calculated based on R/ho values of 4 K for horizontal surfaces and 0 K for vertical surfaces; total solar intensity values used for the calculations were calculated using equations in Chapter 14. Surface Colors. Sol-air temperature values are given in the Example Cooling and Heating Load Calculations section for two values of the parameter /ho; the value of 0.026 is appropriate for a
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Nonresidential Cooling and Heating Load Calculations Table 15
Solar Absorptance Values of Various Surfaces
Surface
Absorptance
Brick, red (Purdue) a Paint Redb Black, matteb Sandstoneb White acrylica Sheet metal, galvanized Newa Weathereda Shingles Grayb Brownb Blackb Whiteb Concretea,c
Licensed for single user. © 2017 ASHRAE, Inc.
aIncropera
and DeWitt (1990).
0.63 0.63 0.94 0.50 0.26 0.65 0.80 0.82 0.91 0.97 0.75 0.60 to 0.83 bParker
et al. (2000).
cMiller
(1971).
light-colored surface, whereas 0.052 represents the usual maximum value for this parameter (i.e., for a dark-colored surface or any surface for which the permanent lightness cannot reliably be anticipated). Solar absorptance values of various surfaces are included in Table 15. This procedure was used to calculate the sol-air temperatures included in the Examples section. Because of the tedious solar angle and intensity calculations, using a simple computer spreadsheet or other software for these calculations can reduce the effort involved.
Calculating Conductive Heat Gain Using Conduction Time Series In the RTS method, conduction through exterior walls and roofs is calculated using CTS values. Wall and roof conductive heat input at the exterior is defined by the familiar conduction equation as qi,-n = UA(te,-n – trc)
(30)
where qi,n U A te,-n trc
= = = = =
conductive heat input for surface n hours ago, W overall heat transfer coefficient for surface, W/(m2 ·K) surface area, m2 sol-air temperature n hours ago, °C presumed constant room air temperature, °C
Conductive heat gain through walls or roofs can be calculated using conductive heat inputs for the current hours and past 23 h and conduction time series: q = c0qi, + c1qi,-1 + c2qi,-2 + c3qi,-3 + … + c23qi,-23
(31)
18.25 material data used in the calculations for walls and roofs in Tables 16 and 17 are listed in Table 18. Heat gains calculated for walls or roofs using periodic response factors (and thus CTS) are identical to those calculated using conduction transfer functions for the steady periodic conditions assumed in design cooling load calculations. The methodology for calculating periodic response factors from conduction transfer functions was originally developed as part of ASHRAE research project RP-875 (Spitler and Fisher 1999b; Spitler et al. 1997). For walls and roofs that are not reasonably close to the representative constructions in Tables 16 and 17, CTS coefficients may be computed with a computer program such as that described by Iu and Fisher (2004). For walls and roofs with thermal bridges, the procedure described by Karambakkam et al. (2005) may be used to determine an equivalent wall construction, which can then be used as the basis for finding the CTS coefficients. When considering the level of detail needed to make an adequate approximation, remember that, for buildings with windows and internal heat gains, the conduction heat gains make up a relatively small part of the cooling load. For heating load calculations, the conduction heat loss may be more significant. The tedious calculations involved make a simple computer spreadsheet or other computer software a useful labor saver.
6.5
HEAT GAIN THROUGH INTERIOR SURFACES
Whenever a conditioned space is adjacent to a space with a different temperature, heat transfer through the separating physical section must be considered. The heat transfer rate is given by q = UA(tb – ti) (32) where q = heat transfer rate, W U = coefficient of overall heat transfer between adjacent and conditioned space, W/(m2 ·K) A = area of separating section concerned, m2 tb = average air temperature in adjacent space, °C ti = air temperature in conditioned space, °C
U-values can be obtained from Chapter 27. Temperature tb may differ greatly from ti. The temperature in a kitchen or boiler room, for example, may be as much as 8 to 28 K above the outdoor air temperature. Actual temperatures in adjoining spaces should be measured, when possible. Where nothing is known except that the adjacent space is of conventional construction, contains no heat sources, and itself receives no significant solar heat gain, tb – ti may be considered the difference between the outdoor air and conditioned space design dry-bulb temperatures minus 3 K. In some cases, air temperature in the adjacent space corresponds to the outdoor air temperature or higher.
where
Floors
q = hourly conductive heat gain for surface, W qi, = heat input for current hour qi,-n = heat input n hours ago c0, c1, etc.=conduction time factors
For floors directly in contact with the ground or over an underground basement that is neither ventilated nor conditioned, sensible heat transfer may be neglected for cooling load estimates because usually there is a heat loss rather than a gain. An exception is in hot climates (i.e., where average outdoor air temperature exceeds indoor design condition), where the positive soil-to-indoor temperature difference causes sensible heat gains (Rock 2005). In many climates and for various temperatures and local soil conditions, moisture transport up through slabs-on-grade and basement floors is also significant, and contributes to the latent heat portion of the cooling load.
Conduction time factors for representative wall and roof types are included in Tables 16 and 17. Those values were derived by first calculating conduction transfer functions for each example wall and roof construction. Assuming steady-periodic heat input conditions for design load calculations allows conduction transfer functions to be reformulated into periodic response factors, as demonstrated by Spitler and Fisher (1999a). The periodic response factors were further simplified by dividing the 24 periodic response factors by the respective overall wall or roof U-factor to form the conduction time series. The conduction time factors can then be used in Equation (31) and provide a way to compare time delay characteristics between different wall and roof constructions. Construction
6.6
CALCULATING COOLING LOAD
The instantaneous cooling load is the rate at which heat energy is convected to the zone air at a given point in time. Computation of cooling load is complicated by the radiant exchange between
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18.26
2017 ASHRAE Handbook—Fundamentals (SI) Table 16
Wall Conduction Time Series (CTS)
Curtainwalls
Studwalls
25 mm Metal Metal 25 mm Spandrel Spandrel Stone, Metal Metal 25 mm 25 mm Wall Panel, Wall Panel, Stone, Glass, Glass, Wall Panel, Wall Panel, Stone, Stone, Sheathing, Sheathing, Sheathing, Sheathing, R-1.8 R-3.9 R-3.5 R-1.8 R-3.5 R-1.8 R-3.5 R-.9 R-3.9 R-1.9 Insulation Insulation Insulation Insulation Insulation Insulation Batt Batt Batt Batt Board, Board, Board, Board, Board, Board, Insulation, Insulation, Insulation, Insulation, Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Wall Number
1
2
3
4
5
6
7
8
9
10
U, W/(m2 ·K)
0.430
0.245
0.431
0.246
0.429
0.245
0.418
0.231
0.416
0.230
Total R
2.33
4.08
2.32
4.07
2.33
4.08
2.39
4.34
2.40
4.35
Licensed for single user. © 2017 ASHRAE, Inc.
Hour
Conduction Time Factors, %
Conduction Time Factors, %
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
18.0 57.1 19.8 4.0 0.8 0.2 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
3.4 35.9 36.8 15.9 5.5 1.8 0.6 0.2 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
25.0 56.1 15.2 3.0 0.6 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
5.4 40.9 33.8 13.4 4.5 1.4 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
8.3 44.0 31.2 11.6 3.5 1.0 0.3 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
1.4 22.3 35.9 23.2 10.7 4.2 1.5 0.5 0.2 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
19.3 57.5 18.7 3.7 0.7 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
5.6 45.0 34.4 11.1 2.9 0.7 0.2 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
6.5 41.1 32.7 13.3 4.5 1.4 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
1.6 24.9 37.3 21.9 9.2 3.4 1.2 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
Layer ID from outdoors to indoors (See Table 18)
F01 F09 F04 I02 F04 G01 F02 0 0 0
F01 F09 F04 I02 I02 F04 G01 F02 0 0
F01 F08 F04 I02 F04 G01 F02 0 0 0
F01 F08 F04 I02 I02 F04 G01 F02 0 0
F01 F10 F04 I02 F04 I02 F02 0 0 0
F01 F10 F04 I02 I02 F04 G01 F02 0 0
F01 F08 G03 I04 G01 F02 0 0 0 0
F01 F08 G03 I04 I04 G01 F02 0 0 0
F01 F10 G03 I04 G01 F02 0 0 0 0
F01 F10 G03 I04 I04 G01 F02 0 0 0
surfaces, furniture, partitions, and other mass in the zone. Most heat gain sources transfer energy by both convection and radiation. Radiative heat transfer introduces a time dependency to the process that is not easily quantified. Radiation is absorbed by thermal masses in the zone and then later transferred by convection into the space. This process creates a time lag and dampening effect. The convective portion, on the other hand, is assumed to immediately become cooling load in the hour in which that heat gain occurs. Heat balance procedures calculate the radiant exchange between surfaces based on their surface temperatures and emissivities, but they typically rely on estimated “radiative/convective splits” to determine the contribution of internal loads, including people, lighting, appliances, and equipment, to the radiant exchange. RTS
further simplifies the HB procedure by also relying on an estimated radiative/convective split of wall and roof conductive heat gain instead of simultaneously solving for the instantaneous convective and radiative heat transfer from each surface, as in the HB procedure. Thus, the cooling load for each load component (lights, people, walls, roofs, windows, appliances, etc.) for a particular hour is the sum of the convective portion of the heat gain for that hour plus the time-delayed portion of radiant heat gains for that hour and the previous 23 h. Table 14 contains recommendations for splitting each of the heat gain components into convective and radiant portions. RTS converts the radiant portion of hourly heat gains to hourly cooling loads using radiant time factors, the coefficients of the
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Nonresidential Cooling and Heating Load Calculations Table 16
18.27
Wall Conduction Time Series (CTS) (Continued)
Studwalls
EIFS
Wood Wood 25 mm Siding, Siding, 25 mm Stucco, EIFS, EIFS, Sheathing, Sheathing, Stucco, R-0.9 R-1.8 R-1.9 R-3.9 Sheathing, Sheathing, R-3.9 Insulation Insulation Batt Batt R-1.9 Board, Board, Batt Insulation, Insulation, Batt 12.5 mm 12.5 mm Insulation, Insulation, Sheathing, Sheathing, Wood Wood Gyp. Board Gyp. Board Gyp. Board Gyp. Board
EIFS, R-0.9 Insulation Board, Sheathing, R-3.9 Batt Insulation, Gyp. Board
EIFS, R0.9 Insulation Board, Sheathing, 200 mm LW CMU, Gyp. Board
EIFS, R-1.8 Insulation Board, Sheathing, 200 mm LW CMU, Gyp. Board
Wall Number
11
12
13
14
15
16
17
18
19
20
U, W/(m2 ·K)
0.402
0.226
0.412
0.229
0.670
0.422
0.305
0.191
0.526
0.360
Total R
2.49
4.43
2.43
4.37
1.49
2.37
3.28
5.23
1.90
2.78
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
EIFS, R-0.9 Insulation Board, Sheathing, R-1.9 Batt Insulation, Gyp. Board
Conduction Time Factors, %
Conduction Time Factors, %
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
6.4 40.7 32.4 13.5 4.7 1.6 0.5 0.2 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
1.5 24.3 36.3 21.9 9.8 3.9 1.5 0.5 0.2 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
5.7 40.4 33.6 13.7 4.6 1.4 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
1.3 23.8 37.6 22.5 9.6 3.5 1.2 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
11.9 48.8 26.3 8.8 2.8 0.9 0.3 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
6.0 40.7 31.6 13.2 5.1 2.0 0.8 0.3 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
2.6 25.2 30.7 19.5 10.6 5.5 2.9 1.5 0.8 0.4 0.2 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
0.5 11.9 25.9 22.9 15.4 9.5 5.7 3.4 2.0 1.2 0.7 0.4 0.2 0.1 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 100
1.0 2.0 5.8 8.7 9.3 8.9 8.1 7.2 6.5 5.7 5.1 4.5 4.0 3.6 3.2 2.8 2.5 2.2 2.0 1.8 1.6 1.4 1.2 1.1 100
1.3 1.8 4.5 7.3 8.3 8.2 7.7 7.0 6.4 5.8 5.3 4.8 4.3 3.9 3.5 3.2 2.9 2.6 2.4 2.1 1.9 1.8 1.6 1.4 100
Layer ID from outdoors to indoors (See Table 18)
F01 F11 G02 I04 G01 F02 0 0 0 0
F01 F11 G02 I04 I04 G01 F02 0 0 0
F01 F07 G03 I04 G01 F02 0 0 0 0
F01 F07 G03 I04 I04 G01 F02 0 0 0
F01 F06 I01 G03 F04 G01 F02 0 0 0
F01 F06 I01 I01 G03 F04 G01 F02 0 0
F01 F06 I01 G03 I04 G01 F02 0 0 0
F01 F06 I01 G03 I04 I04 G01 F02 0 0
F01 F06 I01 G03 M03 F04 G01 F02 0 0
F01 F06 I01 I01 G03 M03 F04 G01 F02 0
radiant time series. Radiant time factors are used to calculate the cooling load for the current hour on the basis of current and past heat gains. The radiant time series for a particular zone gives the timedependent response of the zone to a single pulse of radiant energy. The series shows the portion of the radiant pulse that is convected to zone air for each hour. Thus, r0 represents the fraction of the radiant pulse convected to the zone air in the current hour r1 in the previous hour, and so on. The radiant time series thus generated is used to convert the radiant portion of hourly heat gains to hourly cooling loads according to the following equation: Qr, = r0qr, + r1qr, –1 + r2 qr, –2 + r3qr, –3 (33) + … + r23qr, –23 where
Qr, = radiant cooling load Qr for current hour , W qr, = radiant heat gain for current hour, W qr,n = radiant heat gain n hours ago, W r0, r1, etc. = radiant time factors
The radiant cooling load for the current hour, which is calculated using RTS and Equation (33), is added to the convective portion to determine the total cooling load for that component for that hour. Radiant time factors are generated by a heat-balance-based procedure. A separate series of radiant time factors is theoretically required for each unique zone and for each unique radiant energy distribution function assumption. For most common design applications, RTS variation depends primarily on the overall massiveness of the construction and the thermal responsiveness of the surfaces the radiant heat gains strike.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
18.28
2017 ASHRAE Handbook—Fundamentals (SI) Table 16
Wall Conduction Time Series (CTS) (Continued) Brick Walls
Brick, Brick, R-0.9 R-0.9 Brick, Brick, Brick, Brick, Insulation Insulation R-0.9 Board, R-1.8 Sheathing, Sheathing, Board, Sheathing, Sheathing, Insulation Insulation R-1.9 R-3.9 R-1.9 Batt R-3.9 Batt Board, Board, Batt Batt Sheathing, Sheathing, Insulation, Insulation, Insulation, Insulation, Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board
Brick, R-1.8 Insulation Board, 200 mm LW CMU
Brick, 200 mm LW CMU, R-1.9 Batt Insulation, Gyp. Board
Wall Number
21
22
23
24
25
26
27
28
29
U, W/(m2 ·K)
0.573
0.381
0.376
0.217
0.283
0.157
0.583
0.386
0.347
Total R
1.75
2.62
2.66
4.60
3.54
6.36
1.75
2.59
2.88
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
Brick, R-0.9 Insulation Board, 200 mm LW CMU
Conduction Time Factors, %
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
0.2 4.8 13.9 16.7 14.9 12.0 9.2 7.0 5.3 4.0 3.0 2.3 1.7 1.3 1.0 0.7 0.5 0.4 0.3 0.2 0.2 0.1 0.1 0.1 100
0.1 3.0 11.1 15.5 15.0 12.7 10.1 7.8 6.0 4.6 3.5 2.6 2.0 1.5 1.1 0.9 0.7 0.5 0.4 0.3 0.2 0.2 0.1 0.1 100
0.2 4.1 13.3 16.6 14.8 11.8 9.2 7.1 5.4 4.2 3.2 2.4 1.9 1.4 1.1 0.8 0.6 0.5 0.4 0.3 0.2 0.2 0.1 0.1 100
0.1 1.6 8.5 14.5 15.2 13.1 10.6 8.3 6.5 5.0 3.9 3.0 2.3 1.8 1.4 1.1 0.8 0.6 0.5 0.4 0.3 0.2 0.2 0.1 100
0.1 1.5 6.8 11.7 13.3 12.7 11.1 9.2 7.5 5.9 4.7 3.6 2.8 2.2 1.7 1.3 1.0 0.8 0.6 0.5 0.4 0.3 0.2 0.2 100
0.4 0.5 2.0 5.3 8.2 9.7 10.1 9.6 8.8 7.8 6.8 5.8 4.9 4.1 3.4 2.8 2.3 1.9 1.5 1.2 1.0 0.8 0.6 0.5 100
0.6 0.8 2.6 5.5 7.6 8.7 9.0 8.7 8.2 7.4 6.6 5.8 5.0 4.3 3.7 3.1 2.6 2.2 1.9 1.6 1.3 1.1 0.9 0.7 100
0.8 0.8 2.1 4.5 6.6 7.9 8.4 8.4 8.0 7.4 6.7 6.0 5.3 4.7 4.1 3.5 3.0 2.6 2.2 1.9 1.6 1.4 1.1 1.0 100
1.6 1.5 1.9 3.3 5.0 6.2 6.9 7.1 7.0 6.7 6.3 5.9 5.4 5.0 4.5 4.1 3.7 3.4 3.0 2.7 2.5 2.2 2.0 1.8 100
Layer ID from outdoors to indoors (See Table 18)
F01 M01 F04 I01 G03 F04 G01 F02 0 0
F01 M01 F04 I01 I01 G03 F04 G01 F02 0
F01 M01 F04 G03 I04 G01 F02 0 0 0
F01 M01 F04 G03 I04 I04 G01 F02 0 0
F01 M01 F04 I01 G03 I04 G01 F02 0 0
F01 M01 F04 I01 I01 G03 I04 I04 G01 F02
F01 M01 F04 I01 M03 F02 0 0 0 0
F01 M01 F04 I01 I01 M03 F02 0 0 0
F01 M01 F04 M03 I04 G01 F02 0 0 0
One goal in developing RTS was to provide a simplified method based directly on the HB method; thus, it was deemed desirable to generate RTS coefficients directly from a heat balance. A heat balance computer program was developed to do this: Hbfort, which is included as part of Cooling and Heating Load Calculation Principles (Pedersen et al. 1998). The RTS procedure is described by Spitler et al. (1997). The procedure for generating RTS coefficients may be thought of as analogous to the custom weighting factor generation procedure used by DOE 2.1 (Kerrisk et al. 1981; Sowell 1988a, 1988b). In both cases, a zone model is pulsed with a heat gain. With DOE 2.1, the resulting loads are used to estimate the best values of the transfer function method weighting factors to most closely match the load profile. In the procedure described
here, a unit periodic heat gain pulse is used to generate loads for a 24 h period. As long as the heat gain pulse is a unit pulse, the resulting loads are equivalent to the RTS coefficients. Two different radiant time series are used: solar, for direct transmitted solar heat gain (radiant energy assumed to be distributed to the floor and furnishings only) and nonsolar, for all other types of heat gains (radiant energy assumed to be uniformly distributed on all internal surfaces). Nonsolar RTS apply to radiant heat gains from people, lights, appliances, walls, roofs, and floors. Also, for diffuse solar heat gain and direct solar heat gain from fenestration with indoor shading (blinds, drapes, etc.), the nonsolar RTS should be used. Radiation from those sources is assumed to be more uniformly
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Nonresidential Cooling and Heating Load Calculations Table 16
18.29
Wall Conduction Time Series (CTS) (Continued) Brick Walls
Brick, Brick, Brick, 200 mm R-0.9 R-1.8 LW CMU, Insulation Insulation R-3.9 Board, Board, Batt 200 mm 200 mm Insulation, HW CMU, HW CMU, Gyp. Board Gyp. Board Gyp. Board
Brick, R-0.9 Insulation Board, Brick
Wall Number
30
31
32
33
34
35
36
37
38
U, W/(m2 ·K)
0.207
0.630
0.406
0.704
0.436
0.515
0.355
0.549
0.355
Total R
4.83
1.59
2.46
1.42
2.30
1.94
2.82
1.82
2.82
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
1.9 1.8 1.8 2.7 4.0 5.4 6.2 6.7 6.8 6.6 6.4 6.0 5.6 5.2 4.8 4.4 4.0 3.7 3.4 3.1 2.8 2.5 2.3 2.1 100
1.8 1.7 2.4 3.8 5.1 6.0 6.5 6.6 6.6 6.4 6.1 5.7 5.3 4.9 4.6 4.2 3.8 3.5 3.2 2.9 2.6 2.4 2.1 1.9 100
2.0 1.9 2.3 3.4 4.6 5.5 6.1 6.3 6.3 6.2 6.0 5.7 5.4 5.0 4.7 4.3 4.0 3.7 3.4 3.1 2.9 2.6 2.4 2.2 100
0.9 1.3 3.3 5.8 7.3 8.0 8.2 7.9 7.5 6.9 6.2 5.6 5.0 4.4 3.8 3.3 2.9 2.5 2.2 1.9 1.6 1.4 1.2 1.0 100
1.0 1.2 2.8 5.0 6.6 7.5 7.8 7.7 7.4 6.9 6.4 5.8 5.2 4.6 4.1 3.6 3.2 2.8 2.4 2.1 1.8 1.6 1.4 1.2 100
3.3 3.1 3.0 3.1 3.4 3.8 4.2 4.6 4.8 5.0 5.1 5.1 5.1 5.0 4.9 4.7 4.6 4.4 4.2 4.1 3.9 3.7 3.6 3.4 100
3.4 3.3 3.2 3.2 3.4 3.7 4.1 4.4 4.6 4.8 4.9 5.0 4.9 4.9 4.8 4.7 4.6 4.4 4.3 4.1 4.0 3.9 3.7 3.6 100
3.8 3.8 3.7 3.7 3.8 3.9 4.1 4.2 4.3 4.4 4.5 4.5 4.6 4.6 4.5 4.5 4.3 4.3 4.2 4.2 4.1 4.0 4.0 3.9 100
3.9 3.8 3.8 3.8 3.8 3.9 4.0 4.2 4.3 4.4 4.5 4.5 4.5 4.5 4.5 4.5 4.3 4.3 4.2 4.2 4.1 4.1 4.0 3.9 100
Layer ID from outdoors to indoors (See Table 18)
F01 M01 F04 M03 I04 I04 G01 F02 0 0
F01 M01 F04 I01 M05 G01 F02 0 0 0
F01 M01 F04 I01 I01 M05 G01 F02 0 0
F01 M01 F04 I01 M01 F02 0 0 0 0
F01 M01 F04 I01 I01 M01 F02 0 0 0
F01 M01 F04 I01 M13 F04 G01 F02 0 0
F01 M01 F04 I01 I01 M13 F04 G01 F02 0
F01 M01 F04 I01 M16 F04 G01 F02 0 0
F01 M01 F04 I01 I01 M16 F04 G01 F02 0
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
Brick, R-1.8 Insulation Board, Brick
Brick, Brick, Brick, Brick, R-0.9 R-1.8 R-0.9 R-1.8 Insulation Insulation Insulation Insulation Board, Board, Board, Board, 200 mm 200 mm 300 mm 300 mm LW LW HW HW Concrete, Concrete, Concrete, Concrete, Gyp. Board Gyp. Board Gyp. Board Gyp. Board
Conduction Time Factors, %
distributed onto all room surfaces. Effect of beam solar radiation distribution assumptions is addressed by Hittle (1999). Representative solar and nonsolar RTS data for light, medium, and heavyweight constructions are provided in Tables 19 and 20. Those were calculated using the Hbfort computer program (Pedersen et al. 1998) with zone characteristics listed in Table 21. Customized RTS values may be calculated using the HB method where the zone is not reasonably similar to these typical zones or where more precision is desired. ASHRAE research project RP-942 compared HB and RTS results over a wide range of zone types and input variables (Rees et al. 2000; Spitler et al. 1998). In general, total cooling loads calculated using RTS closely agreed with or were slightly higher than those of the HB
method with the same inputs. The project examined more than 5000 test cases of varying zone parameters. The dominating variable was overall thermal mass, and results were grouped into lightweight, U.S. medium-weight, U.K. medium-weight, and heavyweight construction. Best agreement between RTS and HB results was obtained for light- and medium-weight construction. Greater differences occurred in heavyweight cases, with RTS generally predicting slightly higher peak cooling loads than HB. Greater differences also were observed in zones with extremely high internal radiant loads and large glazing areas or with a very lightweight exterior envelope. In this case, heat balance calculations predict that some of the internal radiant load will be transmitted to the outdoor
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
18.30
2017 ASHRAE Handbook—Fundamentals (SI) Table 16
Wall Conduction Time Series (CTS) (Continued)
Brick Walls Brick, 200 mm HW Concrete, R-1.9 Batt Insulation, Gyp. Board
Concrete Block Walls
200 mm 200 mm Brick, 25 mm 25 mm LW CMU LW CMU 200 mm Stucco, Stucco, 200 mm w/Fill w/Fill HW 200 mm 200 mm 200 mm Concrete, LW CMU, LW CMU, Insulation, Insulation, HW CMU, HW CMU, R-3.9 R-1.9 R-3.9 R-3.9 R-1.9 R-1.9 R-3.9 Batt Batt Batt Batt Batt Batt Batt Insulation, Insulation, Insulation, Insulation, Insulation, Insulation, Insulation, Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board
Wall Number
39
40
41
42
43
44
45
46
47
U, W/(m2 ·K)
0.383
0.217
0.382
0.219
0.335
0.203
0.412
0.229
1.058
Total R
2.61
4.60
2.62
4.56
2.99
4.93
2.42
4.37
0.95
Hour Conduction Time Factors, %
Licensed for single user. © 2017 ASHRAE, Inc.
200 mm LW CMU w/Fill Insulation
Conduction Time Factors, %
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
3.4 3.3 3.3 3.6 4.0 4.4 4.7 4.8 4.9 4.9 4.9 4.8 4.7 4.6 4.5 4.4 4.2 4.1 4.0 3.9 3.8 3.7 3.6 3.5 100
3.5 3.4 3.3 3.5 3.8 4.2 4.5 4.7 4.8 4.9 4.9 4.8 4.7 4.6 4.5 4.4 4.3 4.2 4.1 4.0 3.9 3.8 3.7 3.6 100
0.2 4.6 13.3 15.8 14.0 11.4 9.0 7.0 5.5 4.3 3.4 2.6 2.0 1.6 1.3 1.0 0.8 0.6 0.5 0.4 0.3 0.2 0.2 0.1 100
0.2 1.9 8.8 13.9 14.1 12.3 10.0 8.1 6.4 5.1 4.1 3.2 2.6 2.1 1.6 1.3 1.0 0.8 0.7 0.5 0.4 0.3 0.3 0.2 100
0.6 1.6 5.7 9.5 10.8 10.3 9.3 8.1 7.0 6.0 5.1 4.4 3.7 3.2 2.7 2.3 2.0 1.7 1.5 1.2 1.1 0.9 0.8 0.7 100
0.8 1.0 3.4 7.1 9.4 9.8 9.3 8.3 7.4 6.5 5.6 4.9 4.3 3.7 3.2 2.8 2.4 2.1 1.8 1.6 1.4 1.2 1.1 0.9 100
0.5 2.3 8.0 11.6 11.7 10.5 9.1 7.7 6.5 5.5 4.6 3.9 3.3 2.8 2.3 2.0 1.6 1.4 1.2 1.0 0.8 0.7 0.6 0.5 100
0.5 1.2 5.1 9.6 11.3 10.8 9.6 8.3 7.1 6.0 5.1 4.3 3.7 3.1 2.7 2.3 1.9 1.6 1.4 1.2 1.0 0.8 0.7 0.6 100
0.7 10.4 20.6 19.5 14.8 10.5 7.3 5.0 3.5 2.4 1.6 1.1 0.8 0.5 0.4 0.2 0.2 0.1 0.1 0.1 0.0 0.0 0.0 0.0 100
Layer ID from outdoors to indoors (See Table 18)
F01 M01 F04 M15 I04 G01 F02 0 0 0
F01 M01 F04 M15 I04 I04 G01 F02 0 0
F01 M03 I04 G01 F02 0 0 0 0 0
F01 M03 I04 I04 G01 F02 0 0 0 0
F01 M08 I04 G01 F02 0 0 0 0 0
F01 M08 I04 I04 G01 F02 0 0 0 0
F01 F07 M05 I04 G01 F02 0 0 0 0
F01 F07 M05 I04 I04 G01 F02 0 0 0
F01 M08 F02 0 0 0 0 0 0 0
environment and never becomes cooling load in the space. RTS does not account for energy transfer out of the space to the environment, and thus predicted higher cooling loads. ASHRAE research project RP-1117 built two model rooms for which cooling loads were physically measured using extensive instrumentation. The results agreed with previous simulations (Chantrasrisalai et al. 2003; Eldridge et al. 2003; Iu et al. 2003). HB calculations closely approximated measured cooling loads when provided with detailed data for the test rooms. RTS overpredicted measured cooling loads in tests with large, clear, singleglazed window areas with bare concrete floor and no furnishings or internal loads. Tests under more typical conditions (venetian blinds, carpeted floor, office-type furnishings, and normal internal
loads) provided good agreement between HB, RTS, and measured loads.
7. HEATING LOAD CALCULATIONS Techniques for estimating design heating load for commercial, institutional, and industrial applications are essentially the same as for those estimating design cooling loads for such uses, with the following exceptions: • Temperatures outdoor conditioned spaces are generally lower than maintained space temperatures. • Credit for solar or internal heat gains is not included • Thermal storage effect of building structure or content is ignored.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Nonresidential Cooling and Heating Load Calculations Table 16
Wall Conduction Time Series (CTS) (Continued)
Concrete Block Walls
200 mm LW CMU w/Fill Insulation, Gyp. Board
18.31
Precast and Cast-In-Place Block Walls
100 mm 100 mm 100 mm 100 mm 100 mm 100 mm LW LW LW LW LW LW EIFS, EIFS, 300 mm Concrete. Concrete. Concrete. Concrete. Concrete. Concrete. R-0.9 R-1.8 LW CMU R-0.9 R-1.8 R-1.9 R-3.9 R-1.8 R-3.5 Insulation Insulation w/Fill Board Board Batt Batt Board Board Board, Board, Insulation, Insulation, Insulation, Insulation, Insulation, Insulation, Insulation, 200 mm LW 200 mm LW Gyp. Gyp. Gyp. Gyp. Gyp. 100 mm LW 100 mm LW Concrete, Concrete, Board Board Board Board Board Concrete Concrete Gyp. Board Gyp. Board
Wall Number
48
49
50
51
52
53
54
55
56
57
U, W/(m2 ·K)
0.835
0.688
0.675
0.424
0.417
0.230
0.435
0.247
0.652
0.412
Total R
1.20
1.45
1.48
2.36
2.40
4.34
2.30
4.05
1.53
2.41
Licensed for single user. © 2017 ASHRAE, Inc.
Hour Conduction Time Factors, %
Conduction Time Factors, %
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
0.2 3.6 11.8 15.5 14.6 12.2 9.7 7.5 5.8 4.5 3.5 2.7 2.0 1.6 1.2 0.9 0.7 0.6 0.4 0.3 0.3 0.2 0.1 0.1 100
1.0 1.1 2.6 5.0 7.1 8.3 8.5 8.3 7.7 7.0 6.3 5.6 4.9 4.3 3.8 3.3 2.9 2.5 2.2 1.9 1.7 1.5 1.3 1.1 100
0.7 10.4 19.7 18.1 13.9 10.2 7.4 5.4 3.9 2.8 2.1 1.5 1.1 0.8 0.6 0.4 0.3 0.2 0.2 0.1 0.1 0.1 0.0 0.0 100
0.3 7.1 17.4 18.1 14.6 11.1 8.2 6.1 4.5 3.3 2.5 1.8 1.3 1.0 0.7 0.5 0.4 0.3 0.2 0.2 0.1 0.1 0.1 0.0 100
0.4 8.4 18.2 17.9 14.2 10.7 7.9 5.9 4.3 3.2 2.4 1.7 1.3 0.9 0.7 0.5 0.4 0.3 0.2 0.2 0.1 0.1 0.1 0.0 100
0.1 3.8 13.6 17.5 15.6 12.3 9.4 7.0 5.3 3.9 3.0 2.2 1.6 1.2 0.9 0.7 0.5 0.4 0.3 0.2 0.2 0.1 0.1 0.1 100
0.7 0.9 2.8 5.6 7.7 8.7 8.9 8.6 8.0 7.3 6.5 5.7 5.0 4.3 3.7 3.2 2.7 2.3 1.9 1.6 1.4 1.1 0.9 0.8 100
0.9 0.8 1.6 3.7 5.9 7.4 8.2 8.3 8.1 7.6 6.9 6.3 5.6 4.9 4.3 3.7 3.2 2.8 2.4 2.0 1.7 1.5 1.2 1.0 100
2.2 2.2 3.2 4.6 5.7 6.2 6.3 6.2 6.0 5.7 5.4 5.1 4.8 4.5 4.2 3.9 3.7 3.5 3.2 3.0 2.8 2.6 2.5 2.3 100
2.4 2.4 3.1 4.2 5.2 5.7 5.9 5.9 5.8 5.6 5.3 5.1 4.8 4.6 4.3 4.1 3.9 3.6 3.4 3.3 3.1 2.9 2.7 2.6 100
Layer ID from outdoors to indoors (See Table 18)
F01 M08 F04 G01 F02 0 0 0 0 0
F01 M09 F04 G01 F02 0 0 0 0 0
F01 M11 I01 F04 G01 F02 0 0 0 0
F01 M11 I01 I01 F04 G01 F02 0 0 0
F01 M11 I04 G01 F02 0 0 0 0 0
F01 M11 I04 I04 G01 F02 0 0 0 0
F01 M11 I02 M11 F02 0 0 0 0 0
F01 M11 I02 I02 M11 F02 0 0 0 0
F01 F06 I01 M13 G01 F02 0 0 0 0
F01 F06 I01 I01 M13 G01 F02 0 0 0
Thermal bridging effects on wall and roof conduction are greater for heating loads than for cooling loads, and greater care must be taken to account for bridging effects on U-factors used in heating load calculations. Heat losses (negative heat gains) are thus considered to be instantaneous, heat transfer essentially conductive, and latent heat treated only as a function of replacing space humidity lost to the exterior environment. This simplified approach is justified because it evaluates worstcase conditions that can reasonably occur during a heating season. Therefore, the near-worst-case load is based on the following: • Design interior and exterior conditions • Including infiltration and/or ventilation
• No solar effect (at night or on cloudy winter days) • Before the periodic presence of people, lights, and appliances has an offsetting effect Typical commercial and retail spaces have nighttime unoccupied periods at a setback temperature where little to no ventilation is required, building lights and equipment are off, and heat loss is primarily through conduction and infiltration. Before being occupied, buildings are warmed to the occupied temperature (see the following discussion). During occupied time, building lights, equipment, and people cooling loads can offset conduction heat loss, although some perimeter heat may be required, leaving infiltration and ventilation as the primary heating loads. Ventilation heat load may be
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
18.32
2017 ASHRAE Handbook—Fundamentals (SI) Table 16 Wall Conduction Time Series (CTS) (Concluded) Precast and Cast-In-Place Block Walls EIFS EIFS 200 mm 200 mm Finish, Finish, 200 mm 200 mm 300 mm 300 mm LW LW R-1.8 R-3.5 HW HW HW HW Concrete. Concrete. Insulation Insulation Concrete, Concrete, Concrete, Concrete, R-11 R-22 Board, Board, R-11 R-22 R-3.3 R-6.7 Batt Batt 200 mm HW 200 mm HW Batt Batt Batt Batt Insulation, Insulation, Concrete, Concrete, Insulation, Insulation, Insulation, Insulation, Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Gyp. Board Wall Number
58
59
60
61
62
63
64
65
66
U, W/(m2 ·K)
0.068
0.039
0.082
0.045
0.076
0.041
0.047
0.025
0.549
Total R
14.7
25.7
12.1
22.1
13.1
24.2
21.4
40.5
1.8
0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 Total Percentage
1.4 1.6 3.2 5.6 7.2 7.7 7.7 7.3 6.8 6.2 5.6 5.1 4.7 4.2 3.8 3.5 3.2 2.9 2.6 2.4 2.1 1.9 1.8 1.6 100
1.6 1.6 2.4 4.3 6.2 7.2 7.4 7.3 6.9 6.4 5.9 5.4 4.9 4.5 4.1 3.7 3.4 3.1 2.8 2.6 2.4 2.2 2.0 1.8 100
2.8 3.0 4.2 5.2 5.6 5.6 5.5 5.3 5.2 5.0 4.8 4.6 4.4 4.3 4.1 3.9 3.8 3.6 3.5 3.4 3.2 3.1 3.0 2.9 100
2.9 2.9 3.5 4.5 5.2 5.5 5.5 5.4 5.2 5.0 4.9 4.7 4.5 4.4 4.2 4.1 3.9 3.8 3.6 3.5 3.4 3.3 3.1 3.0 100
1.1 2.1 5.5 8.2 8.9 8.6 7.9 7.1 6.4 5.7 5.1 4.6 4.1 3.7 3.3 2.9 2.6 2.3 2.1 1.9 1.7 1.5 1.3 1.2 100
1.2 1.5 3.8 6.9 8.4 8.6 8.1 7.4 6.7 6.1 5.4 4.9 4.4 3.9 3.5 3.2 2.8 2.5 2.3 2.0 1.8 1.6 1.5 1.3 100
2.5 2.4 2.7 3.6 4.7 5.5 5.9 6.0 5.9 5.7 5.5 5.2 5.0 4.7 4.4 4.2 4.0 3.7 3.5 3.3 3.1 3.0 2.8 2.6 100
2.7 2.6 2.5 2.8 3.5 4.3 5.1 5.5 5.8 5.8 5.7 5.5 5.3 5.1 4.8 4.6 4.3 4.1 3.9 3.6 3.4 3.2 3.1 2.9 100
1.2 1.9 4.3 6.6 7.8 8.1 7.9 7.4 6.8 6.2 5.6 5.0 4.5 4.0 3.6 3.2 2.9 2.6 2.3 2.0 1.8 1.6 1.5 1.3 100
Layer ID from outdoors to indoors (See Table 18)
F01 M13 I04 G01 F02 0 0 0 0 0
F01 M13 I04 I04 G01 F02 0 0 0 0
F01 F06 I02 M15 G01 F02 0 0 0 0
F01 F06 I02 I02 M15 G01 F02 0 0 0
F01 M15 I04 G01 F02 0 0 0 0 0
F01 M15 I04 I04 G01 F02 0 0 0 0
F01 M16 I05 G01 F02 0 0 0 0 0
F01 M16 I05 I05 G01 F02 0 0 0 0
F01 M16 F02 0 0 0 0 0 0 0
Conduction Time Factors, %
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
300 mm HW Concrete
offset with heat recovery equipment. These loads (conduction loss, warm-up load, and ventilation load) may not be additive when sizing building heating equipment, and it is prudent to analyze each load and their interactions to arrive at final equipment sizing for heating.
7.1
HEAT LOSS CALCULATIONS
The general procedure for calculation of design heat losses of a structure is as follows: 1. Select outdoor design conditions: temperature, humidity, and wind direction and speed. 2. Select indoor design conditions to be maintained.
3. Estimate temperature in any adjacent unheated spaces. 4. Select transmission coefficients and compute heat losses for walls, floors, ceilings, windows, doors, and foundation elements. 5. Compute heat load through infiltration and any other outdoor air introduced directly to the space. 6. Sum the losses caused by transmission and infiltration.
Outdoor Design Conditions The ideal heating system provides enough heat to match the structure’s heat loss. However, weather conditions vary considerably from year to year, and heating systems designed for the worst weather conditions on record would have a great excess of capacity most of the time. A system’s failure to maintain design conditions
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Nonresidential Cooling and Heating Load Calculations
18.33
Table 17 Roof Conduction Time Series (CTS) Sloped Frame Roofs Metal Metal Roof, Roof, Metal Metal R-3.3 R-6.7 Roof, Roof, Batt Batt R-3.3 R-6.7 Insulation, Insulation, Batt Batt Suspended Suspended Insulation, Insulation, Acoustical Acoustical Gyp. Board Gyp. Board Ceiling Ceiling Roof Number
1
2
3
4
5
6
7
8
9
U, W/(m2 ·K)
0.2485
0.1355
0.2265
0.1287
0.2547
0.0242
0.2348
0.1313
0.2388
4.02
7.38
4.42
7.77
3.93
41.35
4.26
7.62
4.19
Total R
Conduction Time Factors, %
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
Metal Roof, R-3.3 Batt Insulation
Asphalt Asphalt Slate or Shingles, Shingles, Tile, Wood Wood Wood Metal Sheathing, Sheathing, Sheathing, Roof, R-3.3 R-6.7 R-3.3 R-6.7 Batt Batt Batt Batt Insulation, Insulation, Insulation, Insulation Gyp. Board Gyp. Board Gyp. Board
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
6.4 44.2 32.7 11.6 3.6 1.1 0.3 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.3 10.9 28.5 25.9 16.2 8.9 4.6 2.3 1.2 0.6 0.3 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
10.1 55.6 27.3 5.7 1.0 0.2 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.5 14.8 32.1 24.3 13.6 7.1 3.7 1.9 1.0 0.5 0.3 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
26.6 61.0 11.2 1.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
1.9 27.5 34.7 19.1 9.0 4.2 1.9 0.9 0.4 0.2 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.9 16.5 30.1 23.5 14.0 7.5 3.8 1.9 0.9 0.5 0.2 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.0 2.6 13.3 21.0 20.2 15.5 10.6 6.7 4.1 2.5 1.4 0.8 0.5 0.3 0.2 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.8 16.6 32.8 25.0 13.6 6.4 2.8 1.2 0.5 0.2 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
Total Percentage
100
100
100
100
100
100
100
100
100
Layer ID from outdoors to indoors (See Table 18)
F01 F08 G03 F05 I05 G01 F03 0 0 0 0 0 0 0
F01 F08 G03 F05 I05 I05 G01 F03 0 0 0 0 0 0
F01 F08 G03 F05 I05 F05 F16 F03 0 0 0 0 0 0
F01 F08 G03 F05 I05 I05 F05 F16 F03 0 0 0 0 0
F01 F08 G03 F05 I05 F03 0 0 0 0 0 0 0 0
F01 F08 G03 F05 I05 I05 F03 0 0 0 0 0 0 0
F01 F12 G05 F05 I05 F05 G01 F03 0 0 0 0 0 0
F01 F12 G05 F05 I05 I05 F05 G01 F03 0 0 0 0 0
F01 F14 G05 F05 I05 F05 G01 F03 0 0 0 0 0 0
during brief periods of severe weather usually is not critical. However, close regulation of indoor temperature may be critical for some occupancies or industrial processes. Design temperature data and discussion of their application are given in Chapter 14. Generally, the 99% temperature values given in the tabulated weather data are used. However, caution is needed, and local conditions should always be investigated. In some locations, outdoor temperatures are
commonly much lower and wind velocities higher than those given in the tabulated weather data.
Indoor Design Conditions The main purpose of the heating system is to maintain indoor conditions that make most of the occupants comfortable. Keep in mind, however, that the purpose of heating load calculations is to
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
18.34
2017 ASHRAE Handbook—Fundamentals (SI) Table 17 Roof Conduction Time Series (CTS) (Continued) Sloped Frame Roofs
Wood Deck
Wood Wood Slate or Shingles, Shingles, Tile, Wood Wood Wood Membrane, Sheathing, Sheathing, Sheathing, Sheathing, R-6.7 R-3.3 R-6.7 R-1.8 Batt Batt Batt Insulation Insulation, Insulation, Insulation, Board, Gyp. Board Gyp. Board Gyp. Board Wood Deck
Metal Deck Roofs Membrane, Sheathing, R-3.5 Insulation Board, Wood Deck, Suspended Acoustical Ceiling
Membrane, Sheathing, R-1.8 Insulation Board, Metal Deck
Membrane, Sheathing, R-3.5 Insulation Board, Metal Deck
Roof Number
10
11
12
13
14
15
16
17
18
U, W/(m2 ·K)
0.1325
0.2300
0.1298
0.3944
0.2333
0.3306
0.2094
0.4539
0.2528
7.54
4.35
7.70
2.54
4.29
3.02
4.78
2.20
3.96
Total R Hour
Licensed for single user. © 2017 ASHRAE, Inc.
Membrane, Sheathing, R-3.5 Insulation Board, Wood Deck
Membrane, Sheathing, R-1.8 Insulation Board, Wood Deck, Suspended Acoustical Ceiling
Conduction Time Factors, %
Conduction Time Factors, %
Conduction Time Factors, %
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
0.0 2.5 14.0 22.8 21.4 15.7 10.1 6.0 3.4 1.9 1.0 0.6 0.3 0.2 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.6 11.6 24.2 22.1 15.6 10.0 6.2 3.8 2.3 1.4 0.8 0.5 0.3 0.2 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.0 1.7 9.7 17.0 18.1 15.5 11.9 8.5 5.9 4.0 2.6 1.7 1.1 0.7 0.5 0.3 0.2 0.1 0.1 0.1 0.0 0.0 0.0 0.0
0.3 6.9 17.2 17.7 14.3 10.9 8.2 6.2 4.6 3.5 2.6 2.0 1.5 1.1 0.8 0.6 0.5 0.3 0.3 0.2 0.1 0.1 0.1 0.1
0.1 2.1 10.0 15.4 15.2 12.7 10.1 7.9 6.1 4.7 3.6 2.8 2.2 1.7 1.3 1.0 0.8 0.6 0.5 0.4 0.3 0.2 0.2 0.1
0.9 2.7 7.8 10.1 9.8 8.8 7.8 6.9 6.1 5.4 4.8 4.2 3.7 3.3 2.9 2.6 2.3 2.0 1.8 1.5 1.4 1.2 1.1 0.9
1.2 1.5 4.3 7.6 8.8 8.6 8.0 7.2 6.5 5.9 5.3 4.7 4.3 3.8 3.4 3.1 2.8 2.5 2.2 2.0 1.8 1.6 1.5 1.3
18.0 60.0 18.4 3.0 0.5 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
3.3 38.1 37.6 14.6 4.5 1.3 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
Total Percentage
100
100
100
100
100
100
100
100
100
Layer ID from outdoors to indoors (See Table 18)
F01 F14 G05 F05 I05 I05 F05 G01 F03 0 0 0 0 0
F01 F15 G05 F05 I05 F05 G01 F03 0 0 0 0 0 0
F01 F15 G05 F05 I05 I05 F05 G01 F03 0 0 0 0 0
F01 F13 G03 I02 G06 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I02 I02 G06 F03 0 0 0 0 0 0 0
F01 F13 G03 I02 G06 F05 F16 F03 0 0 0 0 0 0
F01 F13 G03 I02 I02 G06 F05 F16 F03 0 0 0 0 0
F01 F13 G03 I02 F08 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I02 I02 F08 F03 0 0 0 0 0 0 0
obtain data for sizing the heating system components. In many cases, the system will rarely be called upon to operate at the design conditions. Therefore, the use and occupancy of the space are general considerations from the design temperature point of view. Later, when the building’s energy requirements are computed, the actual conditions in the space and outdoor environment, including internal heat gains, must be considered.
The indoor design temperature should be selected at the lower end of the acceptable temperature range, so that the heating equipment will not be oversized. Even properly sized equipment operates under partial load, at reduced efficiency, most of the time; therefore, any oversizing aggravates this condition and lowers overall system efficiency. A maximum design dry-bulb temperature of 21°C is recommended for most occupancies. The indoor design value of relative
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Nonresidential Cooling and Heating Load Calculations
18.35
Table 17 Roof Conduction Time Series (CTS) (Continued) Metal Deck Roofs Membrane, Sheathing, R-1.8 Insulation Board, Metal Deck, Suspended Acoustical Ceiling
Membrane, Sheathing, R-3.5 Insulation Board, Metal Deck, Suspended Acoustical Ceiling
Membrane, Sheathing, R-5.3 Insulation Board, Metal Deck
Membrane, Membrane, Sheathing, Sheathing, R-2.6 R-5.3 Insulation Insulation Board, Board, 100 mm 100 mm LW LW Concrete Concrete
Roof Number
19
20
21
22
23
24
25
26
27
U, W/(m2 ·K)
0.3714
0.1880
0.3248
0.1752
0.2485
0.2984
0.1673
0.3059
0.1696
2.69
5.32
3.08
5.71
4.02
3.35
5.98
3.27
5.90
Total R
Conduction Time Factors, %
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
Membrane, Sheathing, R-2.6 Insulation Board, Metal Deck
Concrete Roofs
50 mm 50 mm Concrete Concrete Roof Ballast, Roof Ballast, Membrane, Membrane, Membrane, Sheathing, Sheathing, Sheathing, R-4.4 R-2.6 R-5.3 Insulation Insulation Insulation Board, Board, Board, Metal Deck Metal Deck Metal Deck
Conduction Time Factors, %
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
4.8 40.0 34.7 13.8 4.6 1.4 0.4 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.2 8.8 26.6 26.3 17.3 9.8 5.2 2.7 1.4 0.7 0.4 0.2 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
8.6 52.5 29.8 7.3 1.5 0.3 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.3 12.8 31.1 25.5 14.7 7.7 3.9 2.0 1.0 0.5 0.2 0.1 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
6.4 44.2 32.7 11.6 3.6 1.1 0.3 0.1 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.4 10.1 21.9 19.5 14.2 10.1 7.1 5.0 3.5 2.5 1.7 1.2 0.9 0.6 0.4 0.3 0.2 0.2 0.1 0.1 0.1 0.0 0.0 0.0
0.1 1.3 8.1 14.7 15.8 14.0 11.4 8.8 6.7 5.0 3.7 2.8 2.0 1.5 1.1 0.8 0.6 0.4 0.3 0.2 0.2 0.1 0.1 0.1
0.6 2.2 7.9 11.2 11.2 10.0 8.7 7.5 6.4 5.5 4.7 4.0 3.4 2.9 2.5 2.2 1.8 1.6 1.4 1.2 1.0 0.9 0.7 0.6
0.8 0.9 2.5 5.9 8.6 9.6 9.4 8.7 7.8 6.9 6.0 5.2 4.5 3.9 3.4 2.9 2.5 2.2 1.9 1.6 1.4 1.2 1.1 0.9
Total Percentage
100
100
100
100
100
100
100
100
100
Layer ID from outdoors to indoors (See Table 18)
F01 F13 G03 I02 F08 F05 F16 F03 0 0 0 0 0 0
F01 F13 G03 I02 I02 F08 F05 F16 F03 0 0 0 0 0
F01 F13 G03 I03 F08 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I03 I03 F08 F03 0 0 0 0 0 0 0
F01 F08 G03 F05 I05 G01 F03 0 0 0 0 0 0 0
F01 M17 F13 G03 I03 F08 F03 0 0 0 0 0 0 0
F01 M17 F13 G03 I03 I03 F08 F03 0 0 0 0 0 0
F01 F13 G03 I03 M11 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I03 I03 M11 F03 0 0 0 0 0 0 0
humidity should be compatible with a healthful environment and the thermal and moisture integrity of the building envelope. A minimum relative humidity of 30% is recommended for most situations.
q = A HF
(34)
HF = U t
(35)
Calculation of Transmission Heat Losses
where HF is the heating load factor in W/m2.
Exterior Surface Above Grade. All above-grade surfaces exposed to outdoor conditions (walls, doors, ceilings, fenestration, and raised floors) are treated identically, as follows:
Below-Grade Surfaces. An approximate method for estimating below-grade heat loss [based on the work of Latta and Boileau (1969)] assumes that the heat flow paths shown in Figure 12 can be
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
18.36
2017 ASHRAE Handbook—Fundamentals (SI) Table 17 Roof Conduction Time Series (CTS) (Concluded) Concrete Roofs Membrane, Sheathing, R-2.6 Insulation Board, 150 mm LW Concrete
Membrane, Sheathing, R-5.3 Insulation Board, 150 mm LW Concrete
Membrane, Sheathing, R-2.6 Insulation Board, 200 mm LW Concrete
Membrane, Sheathing, R-5.3 Insulation Board, 200 mm LW Concrete
Membrane, Sheathing, R-5.3 Insulation Board, 150 mm HW Concrete
Membrane, Sheathing, R-2.6 Insulation Board, 200 mm HW Concrete
Membrane, Membrane, Membrane, 150 mm HW 150 mm HW Sheathing, Concrete, Concrete, R-5.3 R-3.3 R-6.7 Insulation Batt Batt Board, Insulation, Insulation, 200 mm Suspended Suspended HW Acoustical Acoustical Concrete Ceiling Ceiling
Roof Number
28
29
30
31
32
33
34
35
36
37
U, W/(m2 ·K)
0.2972
0.1669
0.2890
0.1688
0.3167
0.1729
0.3139
0.1719
0.2387
0.1325
3.36
5.99
3.46
5.92
3.16
5.79
3.19
5.82
4.19
7.54
Total R
Conduction Time Factors, %
Hour
Licensed for single user. © 2017 ASHRAE, Inc.
Membrane, Sheathing, R-2.6 Insulation Board, 150 mm HW Concrete
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
1.5 1.7 3.4 6.0 7.5 7.8 7.6 7.1 6.5 6.0 5.5 5.0 4.6 4.2 3.8 3.5 3.2 2.9 2.6 2.4 2.2 2.0 1.8 1.7
1.9 1.7 2.0 3.2 4.9 6.2 6.9 7.0 6.9 6.5 6.1 5.7 5.2 4.8 4.4 4.1 3.7 3.4 3.1 2.9 2.6 2.4 2.2 2.0
2.4 2.3 2.6 3.7 4.9 5.7 6.1 6.1 6.0 5.8 5.5 5.2 5.0 4.7 4.4 4.1 3.9 3.7 3.4 3.2 3.0 2.9 2.7 2.5
1.5 1.4 1.5 2.6 4.6 6.4 7.4 7.8 7.6 7.2 6.7 6.1 5.5 5.0 4.5 4.0 3.6 3.2 2.9 2.6 2.3 2.1 1.9 1.7
2.0 2.4 4.6 6.5 7.0 6.8 6.5 6.1 5.7 5.3 5.0 4.7 4.4 4.1 3.8 3.6 3.4 3.1 2.9 2.8 2.6 2.4 2.3 2.1
2.3 2.2 2.7 4.1 5.4 6.2 6.4 6.3 6.1 5.8 5.5 5.2 4.8 4.5 4.3 4.0 3.8 3.5 3.3 3.1 2.9 2.7 2.6 2.4
2.6 2.6 3.5 4.8 5.7 5.9 5.9 5.7 5.5 5.3 5.0 4.8 4.6 4.4 4.2 4.0 3.8 3.6 3.4 3.3 3.1 3.0 2.8 2.7
2.8 2.7 2.8 3.4 4.3 5.0 5.5 5.6 5.6 5.5 5.3 5.1 4.9 4.7 4.5 4.3 4.1 3.9 3.7 3.6 3.4 3.2 3.1 3.0
1.4 2.3 5.7 8.0 8.2 7.8 7.2 6.6 6.0 5.5 5.0 4.6 4.2 3.8 3.5 3.2 2.9 2.6 2.4 2.2 2.0 1.8 1.7 1.5
1.6 1.6 2.6 4.8 6.5 7.3 7.4 7.1 6.7 6.2 5.7 5.3 4.8 4.4 4.0 3.7 3.4 3.1 2.8 2.6 2.4 2.2 2.0 1.8
Total Percentage
100
100
100
100
100
100
100
100
100
100
Layer ID from outdoors to indoors (See Table 18)
F01 F13 G03 I03 M12 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I03 I03 M12 F03 0 0 0 0 0 0 0
F01 F13 G03 I03 M13 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I03 I03 M13 F03 0 0 0 0 0 0 0
F01 F13 G03 I03 M14 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I03 I03 M14 F03 0 0 0 0 0 0 0
F01 F13 G03 I03 M15 F03 0 0 0 0 0 0 0 0
F01 F13 G03 I03 I03 M15 F03 0 0 0 0 0 0 0
F01 F13 M14 F05 I05 F16 F03 0 0 0 0 0 0 0
F01 F13 M14 F05 I05 I05 F16 F03 0 0 0 0 0 0
used to find the steady-state heat loss to the ground surface, as follows: HF = Uavg (tin – tgr) (36) where Uavg = average U-factor for below-grade surface from Equation (38) or (39), W/(m2 ·K) tin = below-grade space air temperature, °C tgr = design ground surface temperature from Equation (37), °C
The effect of soil heat capacity means that none of the usual external design air temperatures are suitable values for tgr . Ground surface temperature fluctuates about an annual mean value by amplitude A, which varies with geographic location and surface cover. The minimum ground surface temperature, suitable for heat loss estimates, is therefore tgr = tgr – A
(37)
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Nonresidential Cooling and Heating Load Calculations
18.37
Table 18 Thermal Properties and Code Numbers of Layers Used in Wall and Roof Descriptions for Tables 16 and 17
Licensed for single user. © 2017 ASHRAE, Inc.
Layer ID F01 F02 F03 F04 F05 F06 F07 F08 F09 F10 F11 F12 F13 F14 F15 F16 F17 F18 G01 G02 G03 G04 G05 G06 G07 I01 I02 I03 I04 I05 I06 M01 M02 M03 M04 M05 M06 M07 M08 M09 M10 M11 M12 M13 M14 M15 M16 M17
Description Outdoor surface resistance Indoor vertical surface resistance Indoor horizontal surface resistance Wall air space resistance Ceiling air space resistance EIFS finish 25 mm stucco Metal surface Opaque spandrel glass 25 mm stone Wood siding Asphalt shingles Built-up roofing Slate or tile Wood shingles Acoustic tile Carpet Terrazzo 16 mm gyp board 16 mm plywood 13 mm fiberboard sheathing 13 mm wood 25 mm wood 50 mm wood 100 mm wood 25 mm insulation board 50 mm insulation board 75 mm insulation board 89 mm batt insulation 154 mm batt insulation 244 mm batt insulation 100 mm brick 150 mm LW concrete block 200 mm LW concrete block 300 mm LW concrete block 200 mm concrete block 300 mm concrete block 150 mm LW concrete block (filled) 200 mm LW concrete block (filled) 300 mm LW concrete block (filled) 200 mm concrete block (filled) 100 mm lightweight concrete 150 mm lightweight concrete 200 mm lightweight concrete 150 mm heavyweight concrete 200 mm heavyweight concrete 300 mm heavyweight concrete 50 mm LW concrete roof ballast
Thickness, mm
Conductivity, W/(m·K)
— — — — — 9.5 25.4 v0.8 6.4 25.4 12.7 3.2 9.5 12.7 6.4 19.1 12.7 25.4 15.9 15.9 12.7 12.7 25.4 50.8 101.6 25.4 50.8 76.2 89.4 154.4 243.8 101.6 152.4 203.2 304.8 203.2 304.8 152.4 203.2 304.8 203.2 101.6 152.4 203.2 152.4 203.2 304.8 50.8
— — — — — 0.72 0.72 45.28 0.99 3.17 0.09 0.04 0.16 1.59 0.04 0.06 0.06 1.80 0.16 0.12 0.07 0.15 0.15 0.15 0.15 0.03 0.03 0.03 0.05 0.05 0.05 0.89 0.49 0.50 0.71 1.11 1.40 0.29 0.26 0.29 0.72 0.53 0.53 0.53 1.95 1.95 1.95 0.19
Notes: The following notes give sources for the data in this table. 1. Chapter 26, Table 1 for 3.4 m/s wind 2. Chapter 26, Table 1 for still air, horizontal heat flow 3. Chapter 26, Table 1 for still air, downward heat flow 4. Chapter 26, Table 3 for 40 mm space, 32.2°C, horizontal heat flow, 0.82 emittance 5. Chapter 26, Table 3 for 90 mm space, 32.2°C, downward heat flow, 0.82 emittance 6. EIFS finish layers approximated by Chapter 26, Table 4 for 10 mm cement plaster, sand aggregate 7. Chapter 33, Table 3 for steel (mild), 22 gage 8. Chapter 26, Table 4 for architectural glass 9. Chapter 26, Table 4 for marble and granite 10. Chapter 26, Table 4, density assumed same as Southern pine 11. Chapter 26, Table 4 for mineral fiberboard, wet molded, acoustical tile 12. Chapter 26, Table 4 for carpet and rubber pad, density assumed same as fiberboard 13. Chapter 26, Table 4, density assumed same as stone
Specific Resistance Density, Heat, R, kg/m3 kJ/(kg·K) (m2 ·K)/W — — — — — 1856 1856 7824 2528 2560 592 1120 1120 1920 592 368 288 2560 800 544 400 608 608 608 608 43 43 43 19 19 19 1920 512 464 512 800 800 512 464 512 800 1280 1280 1280 2240 2240 2240 640 14. 15. 16. 17. 18. 19. 20. 21. 22. 23. 24. 25. 26. 27. 28.
— — — — — 0.84 0.84 0.50 0.88 0.79 1.17 1.26 1.46 1.26 1.30 0.59 1.38 0.79 1.09 1.21 1.30 1.63 1.63 1.63 1.63 1.21 1.21 1.21 0.96 0.96 0.96 0.79 0.88 0.88 0.88 0.92 0.92 0.88 0.88 0.88 0.92 0.84 0.84 0.84 0.90 0.90 0.90 0.84
0.04 0.12 0.16 0.15 0.18 — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — — —
R
Mass, kg/m2
0.04 0.12 0.16 0.15 0.18 0.01 0.04 0.00 0.01 0.01 0.14 0.08 0.06 0.01 0.17 0.31 0.22 0.01 0.10 0.14 0.19 0.08 0.17 0.33 0.66 0.88 1.76 2.64 1.94 3.34 5.28 0.11 0.31 0.41 0.43 0.18 0.22 0.53 0.78 1.04 0.28 0.19 0.29 0.38 0.08 0.10 0.16 0.27
— — — — — 17.7 47.2 6.0 16.1 65.1 7.5 3.6 10.7 24.4 3.8 7.0 3.7 65.1 12.7 8.6 5.1 7.7 15.5 30.9 61.8 1.1 2.2 3.3 1.7 3.0 4.7 195.2 78.1 94.3 156.2 162.7 244.0 78.1 94.3 156.2 162.7 130.1 195.2 260.3 341.6 455.5 683.2 32.5
Thermal Capacity, kJ/(m2·K) Notes — — — — — 14.92 39.45 3.07 14.10 51.71 8.79 4.50 15.74 30.67 4.91 4.09 5.11 51.71 13.90 10.42 6.54 12.67 25.35 50.49 100.97 1.43 2.66 4.09 1.64 2.86 4.50 155.34 68.68 82.99 137.36 149.83 224.84 68.68 82.99 137.36 149.83 108.95 163.52 218.10 307.62 410.23 615.24 2 7.19
1 2 3 4 5 6 6 7 8 9 10
11 12 13
14 15 15 15 15 16 16 16 17 17 17 18 19 20 21 22 23 24 25 26 27
28
Chapter 26, Table 4 for nail-base sheathing Chapter 26, Table 4 for Southern pine Chapter 26, Table 4 for expanded polystyrene Chapter 26, Table 4 for glass fiber batt, specific heat per glass fiber board Chapter 26, Table 4 for clay fired brick Chapter 26, Table 4, 7.3 kg block, 200 400 mm face Chapter 26, Table 4, 8.6 kg block, 200 400 mm face Chapter 26, Table 4, 14.5 kg block, 200 400 mm face Chapter 26, Table 4, 15 kg normal weight block, 200 400 mm face Chapter 26, Table 4, 22.7 kg normal weight block, 200 400 mm face Chapter 26, Table 4, 7.3 kg block, vermiculite fill Chapter 26, Table 4, 8.6 kg block, 200 400 mm face, vermiculite fill Chapter 26, Table 4, 14.5 kg block, 200 400 mm face, vermiculite fill Chapter 26, Table 4, 15 kg normal weight block, 200 400 mm face, vermiculite fill Chapter 26, Table 4 for 640 kg/m3 LW concrete
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18.38
2017 ASHRAE Handbook—Fundamentals (SI) Table 19 Representative Nonsolar RTS Values for Light to Heavy Construction Interior Zones With Carpet
% Glass
Medium No Carpet
With Carpet
No Carpet
Licensed for single user. © 2017 ASHRAE, Inc.
With Carpet
Light No Carpet
10% 50% 90% 10% 50% 90% 10% 50% 90% 10% 50% 90% 10% 50% 90% 10% 50% 90%
Hour 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23
Heavy
Medium
Heavy
With Carpet No Carpet With Carpet No Carpet With Carpet No Carpet
Light
Radiant Time Factor, % 47 19 11 6 4 3 2 2 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0
50 18 10 6 4 3 2 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0
53 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0
41 20 12 8 5 4 3 2 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0 0
43 19 11 7 5 3 3 2 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0
46 19 11 7 5 3 2 2 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0 0
100 100 100 100 100 100
46 18 10 6 4 2 2 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0 0
49 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0
52 16 8 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0 0
31 17 11 8 6 4 4 3 3 2 2 2 1 1 1 1 1 1 1 0 0 0 0 0
33 16 10 7 5 4 3 3 2 2 2 2 1 1 1 1 1 1 1 1 1 1 1 0
35 15 10 7 5 4 3 3 2 2 2 2 1 1 1 1 1 1 1 1 1 1 0 0
100 100 100 100 100 100
34 9 6 4 4 4 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 2 1 1
38 9 6 4 4 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 1 1 1 1
42 9 5 4 4 3 3 3 3 2 2 2 2 2 2 2 2 2 1 1 1 1 1 1
22 10 6 5 5 4 4 4 4 3 3 3 3 3 3 2 2 2 2 2 2 2 2 2
25 9 6 5 5 4 4 4 3 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2
28 9 6 5 4 4 4 4 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 1
100 100 100 100 100 100
46 19 11 6 4 3 2 2 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0
40 20 12 8 5 4 3 2 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0
46 18 10 6 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0
31 17 11 8 6 4 4 3 3 2 2 2 1 1 1 1 1 1 1 0 0 0 0 0
33 9 6 5 4 4 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 2 1 1
21 9 6 5 5 4 4 4 4 3 3 3 3 3 3 3 3 2 2 2 2 2 2 2
100 100 100 100 100 100
Table 20 Representative Solar RTS Values for Light to Heavy Construction Light % Glass
With Carpet 10%
50%
Medium No Carpet
90%
10%
50%
With Carpet
90%
10%
Hour 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23
50%
Heavy No Carpet
90%
10%
50%
With Carpet
No Carpet
90%
10%
50%
90%
10%
50%
90%
Radiant Time Factor, % 53 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0
55 17 9 5 3 2 2 1 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0
56 17 9 5 3 2 2 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0 0
44 19 11 7 5 3 3 2 1 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0
45 20 11 7 5 3 2 2 1 1 1 1 1 0 0 0 0 0 0 0 0 0 0 0
46 20 11 7 5 3 2 2 1 1 1 1 0 0 0 0 0 0 0 0 0 0 0 0
52 16 8 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0 0
54 16 8 4 3 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0 0
55 15 8 4 3 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0 0
28 15 10 7 6 5 4 4 3 3 2 2 2 2 1 1 1 1 1 1 1 0 0 0
29 15 10 7 6 5 4 3 3 3 2 2 2 2 1 1 1 1 1 1 1 0 0 0
29 15 10 7 6 5 4 3 3 3 2 2 2 2 1 1 1 1 1 1 1 0 0 0
47 11 6 4 3 2 2 2 2 2 2 2 2 2 2 1 1 1 1 1 1 1 1 1
49 12 6 4 3 2 2 2 2 2 2 2 1 1 1 1 1 1 1 1 1 1 1 1
51 12 6 3 3 2 2 2 2 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1
26 12 7 5 4 4 3 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 2 2
27 13 7 5 4 4 3 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 1 1
28 13 7 5 4 4 3 3 3 3 3 2 2 2 2 2 2 2 2 2 2 2 1 1
100
100
100
100
100
100
100
100
100
100
100
100
100
100
100
100
100
100
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Nonresidential Cooling and Heating Load Calculations Table 21 Construction Class Exterior Wall Light Medium
Licensed for single user. © 2017 ASHRAE, Inc.
Heavy
18.39
RTS Representative Zone Construction for Tables 19 and 20 Roof/Ceiling
Partitions
Steel siding, 50 mm insulation, 100 mm LW concrete, ceil- 19 mm gyp., air space, air space, 19 mm gyp. ing air space, acoustic tile 19 mm gyp. 100 mm face brick, 50 mm insu- 100 mm HW concrete, ceil- 19 mm gyp., air space, lation, air space, 19 mm gyp. ing air space, acoustic tile 19 mm gyp. 100 mm face brick, 200 mm 200 mm HW concrete, ceil- 19 mm gyp., 200 mm HW concrete air space, 50 mm ing air space, acoustic tile HW concrete block, insulation, 19 mm gyp. 19 mm gyp.
Floor
Furnishings
Acoustic tile, ceiling air space, 100 mm LW concrete Acoustic tile, ceiling air space, 100 mm HW concrete Acoustic tile, ceiling air space, 200 mm HW concrete
25 mm wood @ 50% of floor area 25 mm wood @ 50% of floor area 25 mm wood @ 50% of floor area
Fig. 13 Ground Temperature Amplitude
Fig. 12 Heat Flow from Below-Grade Surface
where tgr = mean ground temperature, °C, estimated from the annual average air temperature or from well-water temperatures, shown in Figure 18 of Chapter 34 in the 2011 ASHRAE Handbook— HVAC Applications A = ground surface temperature amplitude, K, from Figure 13 for North America
Figure 14 shows depth parameters used in determining Uavg. For walls, the region defined by z1 and z2 may be the entire wall or any portion of it, allowing partially insulated configurations to be analyzed piecewise. The below-grade wall average U-factor is given by 2k soil U avg,bw = ----------------------- z2 – z1 (38) 2k soil R other 2k soil R other × ln z 2 + ----------------------------- – ln z 1 + ----------------------------- where Uavg,bw = average U-factor for wall region defined by z1 and z2, W/(m2 ·K) ksoil = soil thermal conductivity, W/(m·K) Rother = total resistance of wall, insulation, and indoor surface resistance, (m2 ·K)/W z1, z2 = depths of top and bottom of wall segment under consideration, m (Figure 14)
The value of soil thermal conductivity k varies widely with soil type and moisture content. A typical value of 1.4 W/(m · K) has been used previously to tabulate U-factors, and Rother is approximately 0.259 (m2 · K)/W for uninsulated concrete walls. For these parameters, representative values for Uavg,bw are shown in Table 22. The average below-grade floor U-factor (where the entire basement floor is uninsulated or has uniform insulation) is given by
Fig. 14 Below-Grade Parameters Table 22 Average U-Factor for Basement Walls with Uniform Insulation Uavg,bw from Grade to Depth, W/(m2·K)
Depth, m
Uninsulated
R-0.88
R-1.76
R-2.64
0.3 0.6 0.9 1.2 1.5 1.8 2.1 2.4
2.468 1.898 1.571 1.353 1.195 1.075 0.980 0.902
0.769 0.689 0.628 0.579 0.539 0.505 0.476 0.450
0.458 0.427 0.401 0.379 0.360 0.343 0.328 0.315
0.326 0.310 0.296 0.283 0.272 0.262 0.252 0.244
Soil conductivity = 1.4 W/(m·K); insulation is over entire depth. For other soil conductivities and partial insulation, use Equation (39).
2k soil U avg ,bf = -------------w b
(39)
z f k soil R other z f k soil R other w × ln -----b- + --- + --------------------------- – ln --- + --------------------------- 2 2 2
where wb = basement width (shortest dimension), m
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18.40
2017 ASHRAE Handbook—Fundamentals (SI)
Table 23 Average U-Factor for Basement Floors Uavg,bf ,
W/(m2·K)
zf (Depth of Floor Below Grade), m
6
7
8
9
0.3 0.6 0.9 1.2 1.5 1.8 2.1
0.370 0.310 0.271 0.242 0.220 0.202 0.187
0.335 0.283 0.249 0.224 0.204 0.188 0.175
0.307 0.261 0.230 0.208 0.190 0.176 0.164
0.283 0.242 0.215 0.195 0.179 0.166 0.155
wb (Shortest Width of Basement), m
Soil conductivity is 1.4 W/(m·K); floor is uninsulated. For other soil conductivities and insulation, use Equation (39).
Table 24
Heat Loss Coefficient Fp of Slab Floor Construction
Construction
Insulation
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200 mm block wall, brick facing 100 mm block wall, brick facing Metal stud wall, stucco
Fp , W/(m·K)
Uninsulated R-0.95 (m2 ·K)/W from edge to footer Uninsulated R-0.95 (m2 ·K)/W from edge to footer Uninsulated R-0.95 (m2 ·K)/W from edge to footer Poured concrete wall with Uninsulated duct near perimeter* R-0.95 (m2 ·K)/W from edge to footer
1.17 0.86 1.45 0.85 2.07 0.92 3.67 1.24
*Weighted average temperature of heating duct was assumed at 43°C during heating season (outdoor air temperature less than 18°C).
factor and infiltration is reduced to a simple sensible component. Under cooling conditions, both sensible and latent components are added to the space load to be treated by the air conditioning system.Procedures for estimating the infiltration rate are discussed in Chapter 16. The infiltration rate is reduced to a volumetric flow rate at a known dry bulb/wet bulb condition. Along with indoor air condition, the following equations define the infiltration sensible and latent loads. qs (kW) = [(m3/s)/v)]cp (tin – to)
(43)
where m3/s = volume flow rate of infiltrating air cp = specific heat capacity of air, kJ/(kg·K) v = specific volume of infiltrating air, m3/kg
Assuming standard air conditions (15°C and sea-level conditions) for v and cp , Equation (43) may be written as qs (kW) = 1.23(m3/s)(tin – to)
(44)
The infiltrating air also introduces a latent heating load given by ql (kW) = [(m3/s)/v](Win – Wo)Dh
(45)
where Win = humidity ratio for indoor space air, kgw /kga Wo = humidity ratio for outdoor air, kgw /kga Dh = change in enthalpy to convert 1 kg water from vapor to liquid, kJ/kg
For standard air and nominal indoor comfort conditions, the latent load may be expressed as
zf = floor depth below grade, m (see Figure 14)
Representative values of Uavg,bf for uninsulated basement floors are shown in Table 23. At-Grade Surfaces. Concrete slab floors may be (1) unheated, relying for warmth on heat delivered above floor level by the heating system, or (2) heated, containing heated pipes or ducts that constitute a radiant slab or portion of it for complete or partial heating of the house. The simplified approach that treats heat loss as proportional to slab perimeter allows slab heat loss to be estimated for both unheated and heated slab floors:
ql = 3010(m3/s)(Win – Wo)
(46)
The coefficients 1.23 in Equation (44) and 3010 in Equation (46) are given for standard conditions. They depend on temperature and altitude (and, consequently, pressure).
7.2
HEATING SAFETY FACTORS AND LOAD ALLOWANCES
Infiltration
Before mechanical cooling became common in the second half of the 1900s, and when energy was less expensive, buildings included much less insulation; large, operable windows; and generally more infiltration-prone assemblies than the energy-efficient and much tighter buildings typical of today. Allowances of 10 to 20% of the net calculated heating load for piping losses to unheated spaces, and 10 to 20% more for a warm-up load, were common practice, along with other occasional safety factors reflecting the experience and/or concern of the individual designer. Such measures are less conservatively applied today with newer construction. A combined warm-up/safety allowance of 20 to 25% is fairly common but varies depending on the particular climate, building use, and type of construction. Engineering judgment must be applied for the particular project. Armstrong et al. (1992a, 1992b) provide a design method to deal with warm-up and cooldown load.
Infiltration of outdoor air through openings into a structure is caused by thermal forces, wind pressure, and negative pressure (planned or unplanned) with respect to the outdoors created by mechanical systems. Typically, in building design, if the mechanical systems are designed to maintain positive building pressure, infiltration need not be considered except in ancillary spaces such as entryways and loading areas. Infiltration is treated as a room load and has both sensible and latent components. During winter, this means heat and humidity loss because cold, dry air must be heated to design temperature and moisture must be added to increase the humidity to design condition. Typically, during winter, controlling indoor humidity is not a
Calculation of design heating load estimates has essentially become a subset of the more involved and complex estimation of cooling loads for such spaces. Chapter 19 discusses using the heating load estimate to predict or analyze energy consumption over time. Special provisions to deal with particular applications are covered in the 2015 ASHRAE Handbook—HVAC Applications and the 2016 ASHRAE Handbook—HVAC Systems and Equipment. The 1989 ASHRAE Handbook—Fundamentals was the last edition to contain a chapter dedicated only to heating load. Its contents were incorporated into this volume’s Chapter 17, which describes
q = p HF
(40)
HF = Fp t
(41)
where q = heat loss through perimeter, W Fp = heat loss coefficient per metre of perimeter, W/(m·K), Table 24 p = perimeter (exposed edge) of floor, m
Surfaces Adjacent to Buffer Space. Heat loss to adjacent unconditioned or semiconditioned spaces can be calculated using a heating factor based on the partition temperature difference: HF = U (tin – tb)
(42)
7.3
OTHER HEATING CONSIDERATIONS
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Nonresidential Cooling and Heating Load Calculations Table 25 Common Sizing Calculations in Other Chapters Subject
Volume/Chapter
Duct heat transfer Piping heat transfer Pump power Moist-air sensible heating and cooling Moist-air cooling and dehumidification Air mixing Space heat absorption and moist-air moisture gains Adiabatic mixing of water injected into moist air
ASTM Standard C680 Fundamentals Ch. 4 Table 2 Systems Ch. 44 (3), (4) Fundamentals Ch. 1 (43) Fundamentals Ch. 1 (45) Fundamentals Ch. 1 (46) Fundamentals Ch. 1 (48) Fundamentals Ch. 1
Equation(s)
(47)
steady-state conduction and convection heat transfer and provides, among other data, information on losses through basement floors and slabs.
18.41 load calculations. Chapter 16 includes more information on ventilating commercial buildings. Depending on ventilation requirements and local climate conditions, peak cooling coil loads may occur at peak dehumidification or enthalpy conditions instead of design dry-bulb conditions. Coil loads should be checked against all those peak conditions.
8.3
AIR HEAT TRANSPORT SYSTEMS
Heat transport equipment is usually selected to provide adequate heating or cooling for the peak load condition. However, selection must also consider maintaining desired indoor conditions during all occupied hours, which requires matching the rate of heat transport to room peak heating and cooling loads. Automatic control systems normally vary the heating and cooling system capacity during these off-peak hours of operation.
On/Off Control Systems
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8.
SYSTEM HEATING AND COOLING LOAD EFFECTS
The heat balance (HB) or radiant time series (RTS) methods are used to determine cooling loads of rooms within a building, but they do not address the plant size necessary to reject the heat. Principal factors to consider in determining the plant size are ventilation, heat transport equipment, and air distribution systems. Some of these factors vary as a function of room load, ambient temperature, and control strategies, so it is often necessary to evaluate the factors and strategies dynamically and simultaneously with the heat loss or gain calculations. Detailed analysis of system components and methods calculating their contribution to equipment sizing are beyond the scope of this chapter, which is general in nature. Table 25 lists the most frequently used calculations in other chapters and volumes.
8.1
ZONING
Organization of building rooms into zones as defined for load calculations and air-handling units has no effect on room cooling loads. However, specific grouping and ungrouping of rooms into zones may cause peak system loads to occur at different times during the day or year, and may significantly affect heat removal equipment sizes. For example, if each room is cooled by a separate heat removal system, the total capacity of the heat transport systems equals the sum of peak room loads. Conditioning all rooms by a single heat transport system (e.g., a variable-volume air handler) requires less capacity (equal to the simultaneous peak of the combined rooms load, which includes some rooms at off-peak loads). This may significantly reduce equipment capacity, depending on the configuration of the building.
8.2
VENTILATION
Consult ASHRAE Standard 62.1 and building codes to determine the required quantity of ventilation air for an application, and the various methods of achieving acceptable indoor air quality. The following discussion is confined to the effect of mechanical ventilation on sizing heat removal equipment. Where natural ventilation is used, through operable windows or other means, it is considered as infiltration and is part of the direct-to-room heat gain. Where ventilation air is conditioned and supplied through the mechanical system, its sensible and latent loads are applied directly to heat transport and central equipment, and do not affect room heating and cooling loads. If the mechanical ventilation rate sufficiently exceeds exhaust airflows, air pressure may be positive and infiltration from envelope openings and outdoor wind may not be included in the
On/off control systems, common in residential and light commercial applications, cycle equipment on and off to match room load. They are adaptable to heating or cooling because they can cycle both heating and cooling equipment. In their purest form, their heat transport matches the combined room and ventilation load over a series of cycles.
Variable-Air-Volume Systems Variable-air-volume (VAV) systems have airflow controls that adjust cooling airflow to match the room cooling load. Damper leakage or minimum airflow settings may cause overcooling, so most VAV systems are used in conjunction with separate heating systems. These may be duct-mounted heating coils, or separate radiant or convective heating systems. The amount of heat added by the heating systems during cooling becomes part of the room cooling load. Calculations must determine the minimum airflow relative to off-peak cooling loads. The quantity of heat added to the cooling load can be determined for each terminal by Equation (8) using the minimum required supply airflow rate and the difference between supply air temperature and the room indoor heating design temperature.
Constant-Air-Volume Reheat Systems In constant-air-volume (CAV) reheat systems, all supply air is cooled to remove moisture and then heated to avoid overcooling rooms. Reheat refers to the amount of heat added to cooling supply air to raise the supply air temperature to the temperature necessary for picking up the sensible load. The quantity of heat added can be determined by Equation (8). With a constant-volume reheat system, heat transport system load does not vary with changes in room load, unless the cooling coil discharge temperature is allowed to vary. Where a minimum circulation rate requires a supply air temperature greater than the available design supply air temperature, reheat adds to the cooling load on the heat transport system. This makes the cooling load on the heat transport system larger than the room peak load.
Mixed Air Systems Mixed air systems change the supply air temperature to match the cooling capacity by mixing airstreams of different temperatures; examples include multizone and dual-duct systems. Systems that cool the entire airstream to remove moisture and to reheat some of the air before mixing with the cooling airstream influence load on the heat transport system in the same way a reheat system does. Other systems separate the air paths so that mixing of hot- and colddeck airstreams does not occur. For systems that mix hot and cold airstreams, the contribution to the heat transport system load is determined as follows.
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18.42
2017 ASHRAE Handbook—Fundamentals (SI)
1. Determine the ratio of cold-deck flow to hot-deck flow from Qh ------ = (Tc – Tr)/(Tr – Th) Qc 2. From Equation (9), the hot-deck contribution to room load during off-peak cooling is qrh = 1.23Qh (Th – Tr) where Qh = heating airflow, L/s Qc = cooling airflow, L/s Tc = cooling air temperature, °C Th = heating air temperature, °C Tr = room or return air temperature, °C qrh = heating airflow contribution to room load at off-peak hours, W
Heat Gain from Fans
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Fans that circulate air through HVAC systems add energy to the system through the following processes: • Increasing velocity and static pressure adds kinetic and potential energy • Fan inefficiency in producing airflow and static pressure adds sensible heat (fan heat) to the airflow • Inefficiency of motor and drive dissipates sensible heat The power required to provide airflow and static pressure can be determined from the first law of thermodynamics with the following equation: PA = 0.009804Vp where
at standard air conditions with air density = 1.2 kg/m3 built into the multiplier 0.009804. The power necessary at the fan shaft must account for fan inefficiencies, which may vary from 50 to 70%. This may be determined from PF = PA /F where PF = power required at fan shaft, kW F = fan efficiency, dimensionless
The power necessary at the input to the fan motor must account for fan motor inefficiencies and drive losses. Fan motor efficiencies generally vary from 80 to 95%, and drive losses for a belt drive are 3% of the fan power. This may be determined from PM = (1 + DL) PF /EM ED where = = = = =
If motor and drive are outside the airstream, qf s = PF qfr = (PM – PF ) If motor and drive are inside the airstream, qf s = PM qfr = 0.0 where PF PM qf s qfr
= = = =
power required at fan shaft, kW power required at input to motor, kW heat release to airstream, kW heat release to room housing motor and drive, kW
Supply airstream temperature rise may be determined from psychrometric formulas or Equation (8). Variable- or adjustable-frequency drives (VFDs or AFDs) often drive fan motors in VAV air-handling units. These devices release heat to the surrounding space. Refer to manufacturers’ data for heat released or efficiencies. The disposition of heat released is determined by the drive’s location: in the conditioned space, in the return air path, or in a nonconditioned equipment room. These drives, and other electronic equipment such as building control, data processing, and communications devices, are temperature sensitive, so the rooms in which they are housed require cooling, frequently year round.
Duct Surface Heat Transfer
PA = air power, kW V = flow rate, m3/s p = pressure, kPa
PM ED EM PF DL
housing the motor and drive. Therefore, the following equations may be used to calculate heat generated by fans and motors:
power required at input to motor, kW belt drive efficiency, dimensionless fan motor efficiency, dimensionless power required at fan shaft, kW drive loss, dimensionless
Almost all the energy required to generate airflow and static pressure is ultimately dissipated as heat in the building and HVAC system; a small portion is discharged with any exhaust air. Generally, it is assumed that all the heat is released at the fan rather than dispersed to the remainder of the system. The portion of fan heat released to the airstream depends on the location of the fan motor and drive: if they are within the airstream, all the energy input to the fan motor is released to the airstream. If the fan motor and drive are outdoor the airstream, the energy is split between the airstream and the room
Heat transfer across the duct surface is one mechanism for energy transfer to or from air inside a duct. It involves conduction through the duct wall and insulation, convection at inner and outer surfaces, and radiation between the duct and its surroundings. Chapter 4 presents a rigorous analysis of duct heat loss and gain, and Chapter 23 addresses application of analysis to insulated duct systems. The effect of duct heat loss or gain depends on the duct routing, duct insulation, and its surrounding environment. Consider the following conditions: • For duct run within the area cooled or heated by air in the duct, heat transfer from the space to the duct has no effect on heating or cooling load, but beware of the potential for condensation on cold ducts. • For duct run through unconditioned spaces or outdoors, heat transfer adds to the cooling or heating load for the air transport system but not for the conditioned space. • For duct run through conditioned space not served by the duct, heat transfer affects the conditioned space as well as the air transport system serving the duct. • For an extensive duct system, heat transfer reduces the effective supply air differential temperature, requiring adjustment through air balancing to increase airflow to extremities of the distribution system.
Duct Leakage Air leakage from supply ducts can considerably affect HVAC system energy use. Leakage reduces cooling and/or dehumidifying capacity for the conditioned space, and must be offset by increased airflow (sometimes reduced supply air temperatures), unless leaked air enters the conditioned space directly. Supply air leakage into a ceiling return plenum or leakage from unconditioned spaces into return ducts also affects return air temperature and/or humidity.
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Nonresidential Cooling and Heating Load Calculations
18.43
Determining leakage from a duct system is complex because of the variables in paths, fabrication, and installation methods. Refer to Chapter 21 and publications from the Sheet Metal and Air Conditioning Contractors’ National Association (SMACNA) for methods of determining leakage. In general, good-quality ducts and postinstallation duct sealing provide highly cost-effective energy savings, with improved thermal comfort and delivery of ventilation air.
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Ceiling Return Air Plenum Temperatures The space above a ceiling, when used as a return air path, is a ceiling return air plenum, or simply a return plenum. Unlike a traditional ducted return, the plenum may have multiple heat sources in the air path. These heat sources may be radiant and convective loads from lighting and transformers; conduction loads from adjacent walls, roofs, or glazing; or duct and piping systems within the plenum. As heat from these sources is picked up by the unducted return air, the temperature differential between the ceiling cavity and conditioned space is small. Most return plenum temperatures do not rise more than 0.6 to 1.7 K above space temperature, thus generating only a relatively small thermal gradient for heat transfer through plenum surfaces, except to the outdoors. This yields a relatively large-percentage reduction in space cooling load by shifting plenum loads to the system. Another reason plenum temperatures do not rise more is leakage into the plenum from supply air ducts, and, if exposed to the roof, increasing levels of insulation. Where the ceiling space is used as a return air plenum, energy balance requires that heat picked up from the lights into the return air (1) become part of the cooling load to the return air (represented by a temperature rise of return air as it passes through the ceiling space), (2) be partially transferred back into the conditioned space through the ceiling material below, and/or (3) be partially lost from the space through floor surfaces above the plenum. If the plenum has one or more exterior surfaces, heat gains through them must be considered; if adjacent to spaces with different indoor temperatures, partition loads must be considered, too. In a multistory building, the conditioned space frequently gains heat through its floor from a similar plenum below, offsetting the floor loss. The radiant component of heat leaving the ceiling or floor surface of a plenum is normally so small, because of relatively small temperature differences, that all such heat transfer is considered convective for calculation purposes (Rock and Wolfe 1997). Figure 15 shows a schematic of a typical return air plenum. The following equations, using the heat flow directions shown in Figure 15, represent the heat balance of a return air plenum design for a typical interior room in a multifloor building: q1 = Uc Ac(tp – tr)
(47)
q2 = Uf Af (tp – tfa )
(48)
q3 = 1.1Q(tp – tr)
(49)
qlp – q2 – q1 – q3 = 0
(50)
qr + q1 Q = ----------------------------1.23 t r – t s
(51)
where q1 q2 q3 Q qlp qlr qf qw qr
= = = = = = = = =
tp tr tfa ts
= = = =
heat gain to space from plenum through ceiling, kW heat loss from plenum through floor above, kW heat gain “pickup” by return air, kW return airflow, L/s light heat gain to plenum via return air, kW light heat gain to space, kW heat gain from plenum below, through floor, kW heat gain from exterior wall, kW space cooling load, including appropriate treatment of qlr , qf , and/or qw , kW plenum air temperature, °C space air temperature, °C space air temperature of floor above, °C supply air temperature, °C
By substituting Equations (47), (48), (49), and (51) into heat balance Equation (50), tp can be found as the resultant return air temperature or plenum temperature. The results, although rigorous and best solved by computer, are important in determining the cooling load, which affects equipment size selection, future energy consumption, and other factors. Equations (47) to (51) are simplified to illustrate the heat balance relationship. Heat gain into a return air plenum is not limited to heat from lights. Exterior walls directly exposed to the ceiling space can transfer heat directly to or from return air. For single-story buildings or the top floor of a multistory building, roof heat gain or loss enters or leaves the ceiling plenum rather than the conditioned space directly. The supply air quantity calculated by Equation (51) is only for the conditioned space under consideration, and is assumed to equal the return air quantity. The amount of airflow through a return plenum above a conditioned space may not be limited to that supplied into the space; it will, however, have no noticeable effect on plenum temperature if the surplus comes from an adjacent plenum operating under similar conditions. Where special conditions exist, Equations (47) to (51) must be modified appropriately. Finally, although the building’s thermal storage has some effect, the amount of heat entering the return air is small and may be considered as convective for calculation purposes.
Ceiling Plenums with Ducted Returns
Fig. 15 Schematic Diagram of Typical Return Air Plenum
Compared to those in unducted plenum returns, temperatures in ceiling plenums that have well-sealed return or exhaust air ducts float considerably. In cooling mode, heat from lights and other equipment raises the ceiling plenum’s temperature considerably. Solar heat gain through a poorly insulated roof can drive the ceiling plenum temperature to extreme levels, so much so that heat gains to uninsulated supply air ducts in the plenum can dramatically decrease available cooling capacity to the rooms below. In cold weather, much heat is lost from warm supply ducts. Thus, insulating supply air ducts and sealing them well to minimize air leaks are highly desirable, if not essential. Appropriately insulating roofs and plenums’ exterior walls and minimizing infiltration are also key to lowering total building loads and improving HVAC system performance.
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18.44
2017 ASHRAE Handbook—Fundamentals (SI)
Underfloor Air Distribution Systems Room cooling loads determined by methods in this chapter cannot model two distinguishing aspects of the thermal performance of underfloor air distribution (UFAD) systems under cooling operation:
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• Room air stratification: UFAD systems supply cool air at the floor and extract warmer air at the ceiling, thus creating vertical thermal stratification. Cooling load models assume a well-mixed uniform space temperature. • Underfloor air supply plenums: cool supply air flowing through the underfloor plenum is exposed to heat gain from both the concrete slab (conducted from the warm return air on the adjacent floor below in a multistory building) and the raised floor panels (conducted from the warmer room above). Extensive simulation and experimental research led to the development of a whole-building energy simulation program capable of modeling energy performance and load calculations for UFAD systems (Bauman et al. 2007; Webster et al. 2008). Previously, it was thought that cooling loads for UFAD and overhead (OH) mixing systems were nearly identical. However, energy modeling studies show that the UFAD cooling load is generally higher than that calculated in the same building for a well-mixed system (Schiavon et al. 2010a). The difference is primarily caused by the thermal storage effect of the lower-mass raised-floor panels compared to the greater mass of a structural floor slab. Schiavon et al. (2010b) showed that the presence of the raised floor reduces the slab’s ability to store heat, thereby producing higher peak cooling loads for a raised-floor system than for one without a raised floor. A second contributing factor is that the raised-floor surface above the underfloor plenum tends to be cooler (except when illuminated by the sun) than most other room surfaces, producing a room surface temperature distribution resembling a chilled radiant floor system, which has a different peak cooling load than an all-air system (Feng et al. 2012). The precise magnitude of difference in design cooling loads between OH and UFAD systems is still under investigation, but mainly depends on zone orientation and floor level, and possibly the effects of furniture. Methods for determining UFAD cooling loads will be updated as additional research results become available. For more information about simplified approaches to UFAD cooling load calculations, see the ASHRAE Underfloor Air Distribution (UFAD) Design Guide (ASHRAE 2013), Bauman et al. (2010), and Schiavon et al. (2010c).
Consider a heating hot-water pipe. If the pipe serves a room heater and is routed through the heated space, any heat loss from the pipe adds heat to the room. Heat transfer to the heated space and heat loss from the piping system is null. If the piping is exposed to ambient conditions en route to the heater, the loss must be considered when selecting the heating equipment; if the pipe is routed through a space requiring cooling, heat loss from the piping also becomes a load on the cooling system. In summary, the designer must evaluate both the magnitude of the pipe heat transfer and the routing of the piping.
Pumps Calculating heat gain from pumps is addressed in the section on Electric Motors. For pumps serving hydronic systems, disposition of heat from the pumps depends on the service. For chilledwater systems, energy applied to the fluid to generate flow and pressure becomes a chiller load. For condenser water pumps, pumping energy must be rejected through the cooling tower. The magnitude of pumping energy relative to cooling load is generally small.
9. EXAMPLE COOLING AND HEATING LOAD CALCULATIONS To illustrate the cooling and heating load calculation procedures discussed in this chapter, an example problem has been developed based on the ASHRAE headquarters building located in Atlanta, Georgia. This example is a two-story office building of approximately 3250 m2, including a variety of common office functions and occupancies. In addition to demonstrating calculation procedures, a hypothetical design/construction process is discussed to illustrate (1) application of load calculations and (2) the need to develop reasonable assumptions when specific data are not yet available, as often occurs in everyday design processes. Table 26 summarizes RTS load calculation procedures.
9.1
SINGLE-ROOM EXAMPLE
Calculate the peak heating and cooling loads for the office room shown in Figure 16, for Atlanta, Georgia. The room is on the second floor of a two-story building and has two vertical exterior exposures, with a flat roof above.
Plenums in Load Calculations Currently, most designers include ceiling and floor plenums within neighboring occupied spaces when thermally zoning a building. However, temperatures in these plenums, and the way that they behave, are significantly different from those of occupied spaces. Thus, they should be defined as a separate thermal zone. Most hand and computer-based load calculation routines, though, currently do not allow floating air temperatures or humidities; assuming a constant air temperature in plenums, attics, and other unconditioned spaces is a poor, but often necessary, assumption. The heat balance method does allow floating space conditions, and when fully implemented in design load software, should allow more accurate modeling of plenums and other complex spaces.
8.4
CENTRAL PLANT
Piping Losses must be considered for piping systems that transport heat. For water or hydronic piping systems, heat is transferred through the piping and insulation (see Chapter 23 for ways to determine this transfer). However, distribution of this transferred heat depends on the fluid in the pipe and the surrounding environment.
Fig. 16
Single-Room Example Office
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Nonresidential Cooling and Heating Load Calculations
18.45
Table 26 Summary of RTS Load Calculation Procedures Equation No. in Chapter
Equation External Heat Gain Sol-Air Temperature E t R te = to + --------- – ---------ho ho where te to a Et ho
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R
(29)
= sol-air temperature, °C = outdoor air temperature, °C = absorptance of surface for solar radiation = total solar radiation incident on surface, W/m2 = coefficient of heat transfer by long-wave radiation and convection at outer surface, W/(m2·K) = hemispherical emittance of surface = difference between long-wave radiation incident on surface from sky and surroundings and radiation emitted by blackbody at outdoor air temperature, W/m2; 20 for horizontal surfaces; 0 for vertical surfaces
Wall and Roof Transmission q = c0qi, + c1qi,-1 + c2qi,-2 + … + c23qi,-23 (31) qi,-n = UA(te,-n – trc) (30) where = hourly conductive heat gain for surface, W q qi, = heat input for current hour qi,-n = conductive heat input for surface n hours ago, W c0 , c1, etc. = conduction time factors U = overall heat transfer coefficient for surface, W/(m2·K) A = surface area, m2 Fenestration Transmission qc = UA(Tout – Tin) (14) where q = fenestration transmission heat gain, W U = overall U-factor, including frame and mounting orientation from Table 4 of Chapter 15, W/(m2·K) A = window area, m2 = indoor temperature, °C Tin = outdoor temperature, °C Tout Fenestration Solar = outdoor temperature, °C Tout qb = AEt,b SHGC()IAC(,) (12) qd = A(Et,d + Et,r)SHGCD IACD (13) where = beam solar heat gain, W qb = diffuse solar heat gain, W qd A = window area, m2 Et,b, Et,d , = beam, sky diffuse, and ground-reflected diffuse irradiance, calculated using equations in Chapter 14 and Et,r = beam solar heat gain coefficient as a function of incident angle ; may be interpolated between values in Table 10 SHGC() of Chapter 15 = indoor solar attenuation coefficient for beam solar heat gain coefficient; = 1.0 if no indoor shading device. IAC(.) is a function of shade type and, depending on IAC(.) type, may also be a function of beam solar angle of incidence and shade geometry IACD
= indoor solar attenuation coefficient for diffuse solar heat gain coefficient; = 1.0 if not indoor shading device. IACD is a function of shade type and, depending on type, may also be a function of shade geometry
Equation No. in Chapter
Equation
Partitions, Ceilings, Floors Transmission q = UA(tb – ti) (32) where q = heat transfer rate, W U = coefficient of overall heat transfer between adjacent and conditioned space, W/(m2·K) A = area of separating section concerned, m2 tb = average air temperature in adjacent space, °C ti = air temperature in conditioned space, °C Internal Heat Gain Occupants qs = qs,per N ql = ql,per N where qs = occupant sensible heat gain, W = occupant latent heat gain, W ql = latent heat gain per person, W/person; see Table 1 ql,per = number of occupants N Lighting qel = WFul Fsa where qel = heat gain, W W = total light wattage, W Ful = lighting use factor Fsa = lighting special allowance factor
(1)
Electric Motors qem = (P/EM)FUM FLM where qem P EM FUM FLM
(2)
= heat equivalent of equipment operation, W = motor power rating, W = motor efficiency, decimal fraction 1.0 = motor use factor, 1.0 or decimal fraction 1.0 = motor load factor, 1.0 or decimal fraction 1.0
Hooded Cooking Appliances qs = qinput FU FR where qs = sensible heat gain, W qinput = nameplate or rated energy input, W FU = usage factor; see Tables 5B, 5C, 5D FR = radiation factor; see Tables 5B, 5C, 5D For other appliances and equipment, find qs for Unhooded cooking appliances: Table 5A Other kitchen equipment: Table 5E Hospital and laboratory equipment: Tables 6 and 7 Computers, printers, scanners, etc.: Tables 8 and 9 Miscellaneous office equipment: Table 10 Find ql for Unhooded cooking appliances: Table 5A Other kitchen equipment: Table 5E Ventilation and Infiltration Air Heat Gain qs = 1.23Qs t ql = 1.20 2500Qs W = 3010Qs W where = sensible heat gain due to infiltration, W qs
(9) (10)
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18.46
2017 ASHRAE Handbook—Fundamentals (SI) Table 26 Summary of RTS Load Calculation Procedures (Concluded) Equation No. in Chapter
Equation ql Qs to ti Wo Wi 1.23
= latent heat gain due to infiltration, W = infiltration airflow at standard air conditions, m3/s = outdoor air temperature, °C = indoor air temperature, °C = outdoor air humidity ratio, kg/kg = indoor air humidity ratio, kg/kg = air sensible heat factor at standard air conditions, W/(m3·s)
3010
= air latent heat factor at standard air conditions, W/(m3·s)
Instantaneous Room Cooling Load Qs = Qi,r + Qi,c Ql = qi,l
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where Qs Qi,r
= room sensible cooling load, W = radiant portion of sensible cooling load for current hour, resulting from heat gain element i, W
= convective portion of sensible cooling load, resulting from heat gain element i, W = room latent cooling load, W Ql = latent heat gain for heat gain element i, W qi,l Radiant Portion of Sensible Cooling Load Qi,c
Qi,r = Qr, (33) Qr, = r0qr, + r1qr,–1 + r2qr,–2 + r3qr,–3 + … + r23qr,–23 where Qr, = radiant cooling load Qr for current hour , W = radiant heat gain for current hour, W qr, qr,n = radiant heat gain n hours ago, W r0, r1, etc. = radiant time factors; see Table 19 for radiant time factors for nonsolar heat gains: wall, roof, partition, ceiling, floor, fenestration transmission heat gains, and occupant, lighting, motor, appliance heat gain. Also used for fenestration diffuse solar heat gain; see Table 20 for radiant time factors for fenestration beam solar heat gain.
Room Characteristics Area: 12 m2. Floor: Carpeted 127 mm concrete slab on metal deck above a conditioned space. Roof: Flat metal deck topped with rigid closed-cell polyisocyanurate foam core insulation (R = 5.3), and light-colored membrane roofing. Space above 2.75 m suspended acoustical tile ceiling is used as a return air plenum. Assume 30% of cooling load from the roof is directly absorbed in the return airstream without becoming room load. Use roof U = 0.18 W/(m2·K). Spandrel wall: Spandrel bronze-tinted glass, opaque, backed with air space, rigid mineral fiber insulation (R = 0.88), mineral fiber batt insulation (R = 2.3), and 16 mm gypsum wall board. Use spandrel wall U = 0.44 W/(m2/K). Brick wall: Light-brown-colored face brick (102 mm), low-mass concrete block (152 mm), rigid continuous insulation (R = 0.88), mineral fiber batt insulation (R = 2.3), and gypsum wall board (16 mm). Use brick wall U = 0.44 W/(m2·K). Windows: Double glazed, 6 mm bronze-tinted outdoor pane, 13 mm air space and 6 mm clear indoor pane with light-colored interior miniblinds. Window normal solar heat gain coefficient (SHGC) = 0.49. Windows are nonoperable and mounted in aluminum frames with thermal breaks having overall combined U =
Equation No. in Chapter
Equation
qr, = qi,sFr where qi,s = sensible heat gain from heat gain element i, W Fr = fraction of heat gain that is radiant. Data Sources: Wall transmission: see Table 14 Roof transmission: see Table 14 Floor transmission: see Table 14 Fenestration transmission: see Table 14 Fenestration solar heat gain: see Table 14, Chapter 18 and Tables 14A to 14G, Chapter 15 Lighting: see Table 3 Occupants: see Tables 1 and 14 Hooded cooking appliances: see Tables 5B, 5C, and 5D Unhooded cooking appliances: see Table 5A Other appliances and equipment: see Tables 5E, 8, 9, 10, and 14 Infiltration: see Table 14 Lighting: see Table 3 Convective Portion of Sensible Cooling Load Qi,c = qi,c where qi,c is convective portion of heat gain from heat gain element i, W. qi,c = qi,s(1 – Fr) where qi,s Fr
= sensible heat gain from heat gain element i, W fraction of heat gain that is radiant; see row for radiant portion = for sources of radiant fraction data for individual heat gain elements
3.18 W/(m2·K) (based on Type 5d from Tables 4 and 10 of Chapter 15). Indoor attenuation coefficients (IACs) for indoor miniblinds are based on light venetian blinds (assumed louver reflectance = 0.8 and louvers positioned at 45° angle) with heat-absorbing double glazing (Type 5d from Table 14B of Chapter 15), IAC(0) = 0.74, IAC(60) = 0.65, IAD(diff) = 0.79, and radiant fraction = 0.54. Each window is 1.91 m wide by 1.95 m tall for an area per window = 3.72 m2. South exposure: Orientation Window area Spandrel wall area Brick wall area West exposure: Orientation Window area Spandrel wall area Brick wall area
= = = = = = = =
30° east of true south 3.72 m2 5.57 m2 5.57 m2 60° west of south 3.72 m2 5.57 m2 3.72 m2
Occupancy: 1 person from 8:00 AM to 5:00 PM. Lighting: One 4-lamp pendant fluorescent 2440 mm type. The fixture has four 32 W T-8 lamps plus electronic ballasts (special allowance factor 0.85 per manufacturer’s data), for a total of 110 W for the room. Operation is from 7:00 AM to 7:00 PM. Assume 0% of cooling load from lighting is directly absorbed in the return airstream without becoming room load, per Table 3.
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Nonresidential Cooling and Heating Load Calculations Table 27 January
Licensed for single user. © 2017 ASHRAE, Inc.
Hour db 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
7.7 7.3 6.9 6.6 6.4 6.6 7.4 9.2 11.3 13.1 14.7 15.8 16.7 17.2 17.2 16.6 15.7 14.6 13.0 11.8 10.8 9.8 9.1 8.3
February
18.47
Monthly/Hourly 5% Design Temperatures for Hartsfield-Jackson Atlanta International Airport, °C March
April
May
June
July
August
September October
November December
wb
db
wb
db
wb
db
wb
db
wb
db
wb
db
wb
db
wb
db
wb
db
wb
db
wb
db
wb
7.3 7.1 6.8 6.6 6.4 6.6 7.1 8.4 9.8 11.1 12.3 13.0 13.6 14.0 14.0 13.6 12.9 12.2 11.1 10.2 9.6 8.9 8.3 7.8
8.6 8.2 7.8 7.4 7.2 7.4 8.3 10.2 12.4 14.4 16.2 17.3 18.3 18.8 18.8 18.1 17.2 16.1 14.3 13.1 12.0 10.9 10.1 9.3
7.3 7.1 6.8 6.6 6.4 6.6 7.1 8.4 9.8 11.0 12.1 12.9 13.4 13.8 13.8 13.4 12.8 12.1 10.9 10.2 9.5 8.8 8.3 7.8
12.1 11.6 11.2 10.8 10.6 10.8 11.7 13.8 16.2 18.3 20.2 21.4 22.4 23.1 23.1 22.3 21.3 20.1 18.2 16.8 15.7 14.6 13.7 12.8
9.7 9.4 9.3 9.1 8.9 9.1 9.5 10.6 11.7 12.7 13.7 14.3 14.8 15.1 15.1 14.7 14.2 13.6 12.7 12.0 11.4 10.9 10.5 10.1
15.4 15.0 14.6 14.2 14.0 14.2 15.1 17.2 19.5 21.6 23.4 24.6 25.6 26.2 26.2 25.5 24.5 23.3 21.4 20.1 19.0 17.9 17.1 16.2
12.8 12.6 12.4 12.3 12.2 12.3 12.7 13.6 14.6 15.4 16.2 16.8 17.2 17.4 17.4 17.1 16.7 16.2 15.4 14.8 14.3 13.8 13.5 13.1
19.4 18.9 18.6 18.3 18.1 18.3 19.1 20.9 23.1 25.0 26.7 27.8 28.7 29.2 29.2 28.6 27.7 26.6 24.9 23.7 22.6 21.6 20.8 20.1
16.8 16.6 16.5 16.4 16.3 16.4 16.7 17.4 18.2 18.9 19.5 19.9 20.2 20.4 20.4 20.2 19.9 19.4 18.8 18.4 18.0 17.6 17.3 17.1
22.3 21.8 21.5 21.2 20.9 21.2 21.9 23.8 25.9 27.9 29.6 30.7 31.6 32.1 32.1 31.4 30.6 29.4 27.8 26.6 25.5 24.5 23.7 22.9
19.3 19.1 19.0 18.9 18.8 18.9 19.2 19.8 20.5 21.1 21.7 22.0 22.3 22.5 22.5 22.3 22.0 21.6 21.1 20.7 20.3 20.0 19.8 19.5
23.2 22.8 22.4 22.1 21.9 22.1 22.9 24.8 26.9 28.8 30.6 31.7 32.6 33.1 33.1 32.4 31.6 30.4 28.7 27.5 26.5 25.5 24.7 23.9
20.5 20.4 20.3 20.2 20.1 20.2 20.4 21.0 21.6 22.2 22.7 23.1 23.3 23.5 23.5 23.3 23.0 22.7 22.2 21.8 21.5 21.2 20.9 20.7
23.2 22.8 22.4 22.1 21.9 22.1 22.9 24.7 26.8 28.6 30.2 31.3 32.2 32.7 32.7 32.1 31.2 30.1 28.5 27.3 26.3 25.3 24.6 23.8
20.5 20.4 20.3 20.2 20.1 20.2 20.4 21.0 21.6 22.2 22.7 23.0 23.3 23.4 23.4 23.2 23.0 22.7 22.2 21.8 21.5 21.2 20.9 20.7
20.9 20.4 20.2 19.8 19.6 19.8 20.6 22.4 24.4 26.2 27.8 28.8 29.7 30.2 30.2 29.6 28.7 27.7 26.1 24.9 23.9 23.0 22.3 21.5
18.2 18.0 17.9 17.8 17.7 17.8 18.1 18.7 19.4 20.1 20.6 21.0 21.3 21.5 21.5 21.3 20.9 20.6 20.0 19.6 19.3 18.9 18.7 18.4
15.7 15.2 14.9 14.6 14.3 14.6 15.3 17.3 19.4 21.3 23.1 24.2 25.1 25.7 25.7 25.0 24.1 22.9 21.2 20.0 19.0 17.9 17.2 16.4
13.8 13.6 13.4 13.3 13.2 13.3 13.7 14.5 15.4 16.3 17.1 17.5 17.9 18.2 18.2 17.9 17.5 17.0 16.2 15.7 15.2 14.8 14.4 14.1
11.1 10.6 10.2 9.9 9.7 9.9 10.7 12.7 14.9 16.9 18.6 19.8 20.7 21.3 21.3 20.6 19.7 18.5 16.7 15.5 14.4 13.4 12.6 11.8
10.2 9.9 9.7 9.5 9.4 9.5 10.0 11.1 12.3 13.4 14.4 15.1 15.6 15.9 15.9 15.5 15.0 14.3 13.3 12.7 12.1 11.4 11.0 10.6
8.6 8.2 7.8 7.5 7.3 7.5 8.3 10.1 12.2 14.0 15.6 16.7 17.6 18.1 18.1 17.4 16.6 15.5 13.9 12.7 11.7 10.7 10.0 9.2
8.6 8.2 7.8 7.5 7.3 7.5 8.3 9.6 10.9 12.2 13.3 14.1 14.6 15.0 15.0 14.6 14.0 13.2 12.2 11.3 10.7 10.0 9.5 9.0
Equipment: One computer and a personal printer are used, for which an allowance of 10.76 W/m2 is to be accommodated by the cooling system, for a total of 130 W for the room. Operation is from 8:00 AM to 5:00 PM. Infiltration: For purposes of this example, assume the building is maintained under positive pressure during peak cooling conditions and therefore has no infiltration. Assume that infiltration during peak heating conditions is equivalent to one air change per hour. Weather data: Per Chapter 14, for Atlanta, Georgia, latitude = 33.64, longitude = 84.43, elevation = 313 m above sea level, 99.6% heating design dry-bulb temperature = –5.6°C. For cooling load calculations, use 5% dry-bulb/coincident wet-bulb monthly design day profile calculated per Chapter 14. See Table 27 for temperature profiles used in these examples. Indoor design conditions: 22.2°C for heating; 23.9°C with 50% rh for cooling.
appropriate RTS to the radiant portion, and sum the convective and radiant cooling load components to determine total cooling load at the designated time. Using Equation (1), the lighting heat gain profile, based on the occupancy schedule indicated is
Cooling Loads Using RTS Method
The convective portion is simply the lighting heat gain for the hour being calculated times the convective fraction for non-in-ceiling fluorescent luminaire (pendant), from Table 3: Qc,15 = (110)(43%) = 47.3 W
Traditionally, simplified cooling load calculation methods have estimated the total cooling load at a particular design condition by independently calculating and then summing the load from each component (walls, windows, people, lights, etc). Although the actual heat transfer processes for each component do affect each other, this simplification is appropriate for design load calculations and useful to the designer in understanding the relative contribution of each component to the total cooling load. Cooling loads are calculated with the RTS method on a component basis similar to previous methods. The following example parts illustrate cooling load calculations for individual components of this single room for a particular hour and month. Equations used are summarized in Table 26. Part 1. Internal cooling load using radiant time series. Calculate the cooling load from lighting at 3:00 PM for the previously described office. Solution: First calculate the 24 h heat gain profile for lighting, then split those heat gains into radiant and convective portions, apply the
q1 q2 q3 q4 q5 q6 q7 q8 q9 q10 q11 q12
= (110 W)(0%) = 0 = (110 W)(0%) = 0 = (110 W)(0%) = 0 = (110 W)(0%) = 0 = (110 W)(0%) = 0 = (110 W)(0%) = 0 = (110 W)(100%) = 110 = (110 W)(100%) = 110 = (110 W)(100%) = 110 = (110 W)(100%) = 110 = (110 W)(100%) = 110 = (110 W)(100%) = 110
q13 = (110 W)(100%) = 110 q14 = (110 W)(100%) = 110 q15 = (110 W)(100%) = 110 q16 = (110 W)(100%) = 110 q17 = (110 W)(100%) = 110 q18 = (110 W)(100%) = 110 q19 = (110 W)(0%) = 0 q20 = (110 W)(0%) = 0 q21 = (110 W)(0%) = 0 q22 = (110 W)(0%) = 0 q23 = (110 W)(0%) = 0 q24 = (110 W)(0%) = 0
The radiant portion of the cooling load is calculated using lighting heat gains for the current hour and past 23 h, the radiant fraction from Table 3 (57%), and radiant time series from Table 19, in accordance with Equation (33). From Table 19, select the RTS for medium-weight construction, assuming 50% glass and carpeted floors as representative of the described construction. Thus, the radiant cooling load for lighting is Qr,15 = r0(0.48)q15 + r1(0.48)q14 + r2(0.48)q13 + r3(0.48)q12 + … + r23(0.48)q16 = (0.49)(0.57)(110) + (0.17)(0.57)(110) + (0.09)(0.57)(110) + (0.05)(0.57)(110) + (0.03)(0.57)(110) + (0.02)(0.57)(110) + (0.02)(0.57)(110) + (0.01)(0.57)(110) + (0.01)(0.57)(110) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.01)(0.57)(0)
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18.48
2017 ASHRAE Handbook—Fundamentals (SI) Table 28
Cooling Load Component: Lighting, W
Licensed for single user. © 2017 ASHRAE, Inc.
Heat Gain Hour
Usage Profile, %
Heat Gain
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
0 0 0 0 0 0 100 100 100 100 100 100 100 100 100 100 100 100 0 0 0 0 0 0
Total
Nonsolar RTS Zone Type 8, %
Radiant Cooling Load
Total Sensible Cooling Load
— — — — — — 63 63 63 63 63 63 63 63 63 63 63 63 — — — — — —
49 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0
8 8 7 6 6 5 35 45 50 53 54 55 55 55 56 56 57 58 28 18 13 10 9 8
8 8 7 6 6 5 82 92 97 100 101 102 102 102 103 104 104 105 28 18 13 10 9 8
— — — — — — — — — — — — — — — — — — — — — — — —
8 8 7 6 6 5 82 92 97 100 101 102 102 102 103 104 104 105 28 18 13 10 9 8
752
100
752
1319
—
1319
Convective
Radiant
43%
57%
— — — — — — 110 110 110 110 110 110 110 110 110 110 110 110 — — — — — —
— — — — — — 47 47 47 47 47 47 47 47 47 47 47 47 — — — — — —
1319
567
+ (0.01)(0.57)(0) + (0.01)(0.57)(0) + (0.00)(0.57)(0) + (0.00)(0.57)(440) + (0.00)(0.57)(440) + (0.00)(0.57)(440) = 55.77 W The total lighting cooling load at the designated hour is thus Qlight = Qc,15 + Qr,15 = 47.3 + 55.7 = 103 W See Table 28 for the office’s lighting usage, heat gain, and cooling load profiles. Part 2. Wall cooling load using sol-air temperature, conduction time series and radiant time series. Calculate the cooling load contribution from the spandrel wall section facing 60° west of south at 3:00 PM local standard time in July for the previously described office. Solution: Determine the wall cooling load by calculating (1) sol-air temperatures at the exterior surface, (2) heat input based on sol-air temperature, (3) delayed heat gain through the mass of the wall to the interior surface using conduction time series, and (4) delayed space cooling load from heat gain using radiant time series. First, calculate the sol-air temperature at 3:00 PM local standard time (LST) (4:00 PM daylight saving time) on July 21 for a vertical, dark-colored wall surface, facing 60° west of south, located in Atlanta, Georgia (latitude = 33.64, longitude = 84.43), solar clear sky optical depth for beam irradiance b (“taub”) = 0.515 and d (“taud”) for diffuse irradiance = 2.066 from monthly Atlanta weather data for July (Table 1 in Chapter 14). From Table 27, the calculated outdoor design temperature for that month and time is 33.3°C. The ground reflectivity is assumed g = 0.2. Sol-air temperature is calculated using Equation (30). For the darkcolored wall, /ho = 0.053, and for vertical surfaces, R/ho = 0. The solar irradiance Et on the wall must be determined using the equations in Chapter 14: Solar Angles: = southwest orientation = +60° = surface tilt from horizontal (where horizontal = 0°) = 90° for vertical wall surface 3:00 PM LST = hour 15
% Lighting to Return 0%
Room Sensible Cooling Load
Calculate solar altitude, solar azimuth, surface solar azimuth, and incident angle as follows: From Table 2 in Chapter 14, solar position data and constants for July 21 are ET = –6.4 min = 20.4° Eo = 1324 W/m2 Local standard meridian (LSM) for Eastern Time Zone = 75°. Apparent solar time AST AST = LST + ET/60 + (LSM – LON)/15 = 15 + (–6.4/60) + [(75 – 84.43)/15] = 14.2647 Hour angle H, degrees H = 15(AST – 12) = 15(14.2647 – 12) = 33.97° Solar altitude sin = cos L cos cos H + sin L sin = cos (33.64) cos (20.4) cos (33.97) + sin (33.64) sin (20.4) = 0.841 = sin–1(0.841) = 57.2° Solar azimuth cos = (sin sin L – sin )/(cos cos L) = [(sin (57.2)sin (33.64) – sin (20.4)]/[cos (57.2) cos (33.64)] = 0.258 = cos–1(0.253) = 75.05° Surface-solar azimuth = – = 75.05 – 60 = 15.05° Incident angle cos = cos cos g sin + sin cos = cos (57.2) cos (15.05) sin (90) + sin (57.2) cos (90)
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Nonresidential Cooling and Heating Load Calculations = 0.523 = cos–1(0.523) = 58.45° Beam normal irradiance Eb Eb = Eo exp(–b mab) m = relative air mass = 1/[sin +0.50572(6.07995 + )–1.6364], expressed in degrees = 1.18905 ab = beam air mass exponent = 1.454 – 0.406b – 0.268d + 0.021bd = 0.713566 Eb = 1324 exp[–0.556(1.89050.7055705)] = 805 W/m2 Surface beam irradiance Et,b Et,b = Eb cos = (805) cos (58.5) = 421 W/m2
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Ratio Y of sky diffuse radiation on vertical surface to sky diffuse radiation on horizontal surface Y = 0.55 + 0.437 cos + 0.313 cos 2 = 0.55 + 0.437 cos (58.45) + 0.313 cos2 (58.45) = 0.8644 Diffuse irradiance Ed – Horizontal surfaces Ed = Eo exp(–d mad) ad = diffuse air mass exponent = 0.507 + 0.205b – 0.080 d – 0.190 bd = 0.245137 Ed = Eo exp(–d mad) = 1324 exp(–2.202(1.89050.2369528)] = 133 W/m2 Diffuse irradiance Ed – Vertical surfaces Et,d = EdY = (133)(0.864) = 115 W/m2 Ground reflected irradiance Et,r Et,r = (Eb sin + Ed) g (l – cos = [805 sin (57.2) + 133](0.2)[1 – cos (90)]/2 = 81 W/m2 Total surface irradiance Et Et = ED + Ed + Er = 421 + 115 + 81 = 617 W/m2 Sol-air temperature [from Equation (29)]: Te = to + Et /ho – R/ho = 33.3 + (0.053)(617) – 0 = 66.0°C This procedure is used to calculate the sol-air temperatures for each hour on each surface. Because of the tedious solar angle and intensity calculations, using a simple computer spreadsheet or other computer software can reduce the effort involved. A spreadsheet was used to calculate a 24 h sol-air temperature profile for the data of this example. See Table 29A for the solar angle and intensity calculations and Table 29B for the sol-air temperatures for this wall surface and orientation. Conductive heat gain is calculated using Equations (30) and (31). First, calculate the 24 h heat input profile using Equation (30) and the sol-air temperatures for a southwest-facing wall with dark exterior color: qi,1 qi,2 qi,3 qi,4 qi,5 qi,6 qi,7 qi,8 qi,9 qi,10 qi,11 qi,12 qi,13
= = = = = = = = = = = = =
(0.44)(5.57)(22.3 – 23.9) (0.44)(5.57)(21.9 – 23.9) (0.44)(5.57)(21.6 – 23.9) (0.44)(5.57)(21.2 – 23.9) (0.44)(5.57)(21.0 – 23.9) (0.44)(5.57)(21.2 – 23.9) (0.44)(5.57)(22.0 – 23.9) (0.44)(5.57)(23.9 – 23.9) (0.44)(5.57)(26.1 – 23.9) (0.44)(5.57)(27.9 – 23.9) (0.44)(5.57)(29.7 – 23.9) (0.44)(5.57)(30.8 – 23.9) (0.44)(5.57)(31.7 – 23.9)
= = = = = = = = = = = = =
–4 W –5 –6 –7 –7 –7 –5 0 5 10 14 17 19
18.49 qi,14 qi,15 qi,16 qi,17 qi,18 qi,19 qi,20 qi,21 qi,22 qi,23 qi,24
= = = = = = = = = = =
(0.44)(5.57)(32.2 – 23.9) (0.44)(5.57)(32.2 – 23.9) (0.44)(5.57)(31.6 – 23.9) (0.44)(5.57)(30.7 – 23.9) (0.44)(5.57)(29.6 – 23.9) (0.44)(5.57)(27.8 – 23.9) (0.44)(5.57)(26.6 – 23.9) (0.44)(5.57)(25.6 – 23.9) (0.44)(5.57)(24.6 – 23.9) (0.44)(5.57)(23.8 – 23.9) (0.44)(5.57)(23.0 – 23.9)
= = = = = = = = = = =
20 20 19 17 14 10 7 4 2 0 –2
Next, calculate wall heat gain using conduction time series. The preceding heat input profile is used with conduction time series to calculate the wall heat gain. From Table 16, the most similar wall construction is wall number 1. This is a spandrel glass wall that has similar mass and thermal capacity. Using Equation (31), the conduction time factors for wall 1 can be used in conjunction with the 24 h heat input profile to determine the wall heat gain at 3:00 PM LST: q15
= c0qi,15 + c1qi,14 + c2qi,13 + c3qi,12 + … + c23qi,16 = (0.18)(20) + (0.58)(20) + (0.20)(19) + (0.04)(17) + (0.00)(14) + (0.00)(10) + (0.00)(5) + (0.00)(0) + (0.00)(–5) + (0.00)(–7) + (0.00)(–7) + (0.00)(–7) + (0.00)(–6) + (0.00)(–5) +(0.00)(–4) + (0.00)(–2) + (0.00)(0) + (0.00)(2) + (0.00)(4) + (0.00)(7) + (0.00)(10) + (0.00)(14) + (0.00)(17) + (0.00)(19) = 83.5 W
Because of the tedious calculations involved, a spreadsheet is used to calculate the remainder of a 24 h heat gain profile indicated in Table 29B for the data of this example. Finally, calculate wall cooling load using radiant time series. Total cooling load for the wall is calculated by summing the convective and radiant portions. The convective portion is simply the wall heat gain for the hour being calculated times the convective fraction for walls from Table 14 (54%): Qc = (83.5)(0.54) = 45 W The radiant portion of the cooling load is calculated using conductive heat gains for the current and past 23 h, the radiant fraction for walls from Table 14 (46%), and radiant time series from Table 19, in accordance with Equation (33). From Table 19, select the RTS for medium-weight construction, assuming 50% glass and carpeted floors as representative for the described construction. Use the wall heat gains from Table 29B for 24 h design conditions in July. Thus, the radiant cooling load for the wall at 3:00 PM is Qr,15 = r0(0.46)qi,15 + r1(0.46) qi,14 + r2(0.46) qi,13 + r3(0.46) qi,12 + … + r23(0.46) qi,16 = (0.49)(0.46)(84) + (0.17)(0.46)(63) + (0.09)(0.46)(44) + (0.05)(0.46)(35) + (0.03)(0.46)(28) + (0.02)(0.46)(20) + (0.02)(0.46)(12) + (0.01)(0.46)(3) + (0.01)(0.46)(–2) + (0.01)(0.46)(–4) + (0.01)(0.46)(–4) + (0.01)(0.46)(–3) + (0.01)(0.46)(–2) + (0.01)(0.46)(–1) + (0.01)(0.46)(0) + (0.01)(0.46)(2) + (0.01)(0.46)(4) + (0.01)(0.46)(8) + (0.01)(0.46)(17) + (0.01)(0.46)(43) + (0.00)(0.46)(75) + (0.00)(0.46)(96) + (0.00) (0.46)(103) + (0.00)(0.46)(99) = 27 W The total wall cooling load at the designated hour is thus Qwall = Qc + Qr15 = 45 + 27 = 72 W Again, a simple computer spreadsheet or other software is necessary to reduce the effort involved. A spreadsheet was used with the heat gain profile to split the heat gain into convective and radiant portions, apply RTS to the radiant portion, and total the convective and radiant loads to determine a 24 h cooling load profile for this example, with results in Table 29B. Part 3. Window cooling load using radiant time series. Calculate the cooling load contribution, with and without indoor shading (venetian blinds) for the window area facing 60° west of south at 3:00 PM in July for the conference room example.
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18.50
2017 ASHRAE Handbook—Fundamentals (SI) Table 29A Conduction: Wall Component of Solar Irradiance (Month 7) Direct Beam Solar
Diffuse Solar Heat Gain
Licensed for single user. © 2017 ASHRAE, Inc.
Beam Surface Diffuse Total Solar Sky Subtotal Surface Solar Normal Incident Surface Horizontal Ground Local Apparent Hour Solar Angle Altitude Azimuth Air Mass Y Diffuse, Diffuse, Irradiance, Diffuse, Angle Direct, Eb , Ed , Standard Solar Hour Time H W/m2 W/m2 m W/m2 Ratio W/m2 W/m2 W/m2 W/m2 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
0.26 1.26 2.26 3.26 4.26 5.26 6.26 7.26 8.26 9.26 10.26 11.26 12.26 13.26 14.2647 15.26 16.26 17.26 18.26 19.26 20.26 21.26 22.26 23.26
–176 –161 –146 –131 –116 –101 –86 –71 –56 –41 –26 –11 4 19 33.97 49 64 79 94 109 124 139 154 169
–36 –33 –27 –19 –9 3 14 27 39 51 63 74 76 69 57.2 45 32 20 8 –3 –14 –23 –30 –35
–175 –159 –144 –132 –122 –113 –105 –98 –90 –81 –67 –39 16 57 75.05 86 94 102 109 117 127 138 151 167
— — — — — 16.91455 3.98235 2.22845 1.58641 1.27776 1.11740 1.04214 1.02872 1.07337 1.18905 1.41566 1.86186 2.89735 6.84406 — — — — —
0.0 0.0 0.0 0.0 0.0 27.5 333.0 531.9 647.4 717.2 758.5 779.3 783.1 770.5 739.6 684.6 593.7 440.8 173.7 0.0 0.0 0.0 0.0 0.0
117.4 130.9 144.5 158.1 171.3 172.5 159.5 145.9 132.3 118.8 105.6 92.6 80.2 68.7 58.45 50.4 45.8 45.5 49.7 57.5 67.5 79.0 91.3 104.2
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 132.9 280.4 387.0 436.2 414.2 308.9 112.2 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 21.2 73.0 107.2 131.0 147.7 158.5 164.3 165.4 161.8 153.4 139.6 119.4 90.7 48.3 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 2.2 15.5 34.5 53.8 70.8 83.7 91.2 92.6 87.9 77.5 62.3 43.8 24.2 7.3 0.0 0.0 0.0 0.0 0.0
0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4553 0.5306 0.6332 0.7505 0.8644 0.9555 1.0073 1.0100 0.9631 0.8755 0.7630 0.6452 0.5403 0.4618
0.0 0.0 0.0 0.0 0.0 9.6 32.8 48.2 59.0 66.4 72.2 87.2 104.7 121.5 132.6 133.4 120.3 91.6 46.6 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 11.8 48.4 82.7 112.8 137.3 155.9 178.4 197.4 209.4 210.1 195.7 164.1 115.7 53.8 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 11.8 48.4 82.7 112.8 137.3 155.9 178.4 330.3 489.8 597.1 631.9 578.3 424.7 166.0 0.0 0.0 0.0 0.0 0.0
Table 29B Conduction: Wall Component of Sol-Air Temperatures, Heat Input, Heat Gain, Cooling Load (Month 7) Total Surface Outdoor Local Standard Irradiance, Temp., °C W/m2 Hour 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
0.0 0.0 0.0 0.0 0.0 11.8 48.4 82.7 112.8 137.3 155.9 178.4 330.3 489.8 597.1 631.9 578.3 424.7 166.0 0.0 0.0 0.0 0.0 0.0
23.2 22.8 22.4 22.1 21.9 22.1 22.9 24.8 26.9 28.8 30.6 31.7 32.6 33.1 33.1 32.4 31.6 30.4 28.7 27.5 26.5 25.5 24.7 23.9
Heat Gain, W
Sol-Air Temps., °C
Indoor Temp., °C
Heat Input, W
CTS Type 1, %
Total
23.2 22.8 22.4 22.1 21.9 22.7 25.4 29.2 32.9 36.1 38.8 41.1 50.0 59.0 64.7 65.8 62.1 52.9 37.5 27.5 26.5 25.5 24.7 23.9
23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9 23.9
–2 –3 –4 –4 –5 –3 4 13 22 30 36 42 64 86 99 102 93 71 33 9 6 4 2 0
18 57 20 4 1 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 –1 –2 –3 –4 –4 –2 4 12 21 28 35 44 62 81 95 99 91 70 39 17 8 5 2
Solution: First, calculate the 24 h heat gain profile for the window, then split those heat gains into radiant and convective portions, apply the appropriate RTS to the radiant portion, then sum the convective and radiant cooling load components to determine total window cooling load for the time. The window heat gain components are calculated using Equations (12) to (14). From Part 2, at hour 15 LST (3:00 PM):
Convective Radiant 54% 46% 0 –1 –1 –2 –2 –2 –1 2 6 11 15 19 24 33 44 51 53 49 38 21 9 4 3 1
Nonsolar RTS Zone Type 8, %
0 –1 –1 –2 –2 –2 –1 2 5 9 13 16 20 28 37 44 46 42 32 18 8 4 2 1
Radiant Cooling Load, W
Total Cooling Load, W
4 4 3 2 2 2 2 3 5 8 10 13 15 20 27 32 36 36 32 24 16 11 8 6
5 3 1 1 0 –1 1 5 12 19 26 32 39 54 71 84 89 85 69 45 25 15 10 7
49 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0
Et,b = 421 W/m2 Et,d = 115 W/m2 Er = 81 W/m2
= 58.45°
From Chapter 15, Table 10, for glass type 5d,
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Nonresidential Cooling and Heating Load Calculations
18.51
Table 30 Window Component of Heat Gain (No Blinds or Overhang) (Month 7) Beam Solar Heat Gain
Conduction Heat Gain
Diffuse Solar Heat Gain
Licensed for single user. © 2017 ASHRAE, Inc.
Total Diffuse Beam ConSolar Solar Diffuse duction Window Sky Adjusted Heat Local Beam Surface Surface Subtotal Heat Heat Outdoor Heat Hor. Ground Diffuse, Diffuse, Hemis. Gain, Ed , Y Std. Normal, Incident Beam, Beam Beam Gain, Diffuse, Gain, Temp., Gain, W/m2 IAC Hour W/m2 Angle W/m2 Ratio W/m2 W/m2 SHGC W/m2 SHGC W W °C W W 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
0.0 0.0 0.0 0.0 0.0 27.5 333.0 531.9 647.4 717.2 758.5 779.3 783.1 770.5 739.6 684.6 593.7 440.8 173.7 0.0 0.0 0.0 0.0 0.0
117.4 130.9 144.5 158.1 171.3 172.5 159.5 145.9 132.3 118.8 105.6 92.6 80.2 68.7 58.4 50.4 45.8 45.5 49.7 57.5 67.5 79.0 91.3 104.2
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 132.9 280.4 387.0 436.2 414.2 308.9 112.2 0.0 0.0 0.0 0.0 0.0
0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 0.166 0.321 0.398 0.438 0.448 0.449 0.441 0.403 0.330 0.185 0.000 0.000
1.000 1.000 1.000 1.000 1.000 0.000 0.000 0.000 0.000 0.000 0.000 0.000 1.000 1.000 1.000 1.000 1.000 1.000 1.000 0.000 0.000 0.000 1.000 1.000
0 0 0 0 0 0 0 0 0 0 0 0 82 334 572 710 690 515 184 0 0 0 0 0
0.0 0.0 0.0 0.0 0.0 21.2 73.0 107.2 131.0 147.7 158.5 164.3 165.4 161.8 153.4 139.6 119.4 90.7 48.3 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 2.2 15.5 34.5 53.8 70.8 83.7 91.2 92.6 87.9 77.5 62.3 43.8 24.2 7.3 0.0 0.0 0.0 0.0 0.0
SHGC() = SHGC(58.45) = 0.3978 (interpolated) SHGCD = 0.41
From Chapter 15, Table 14B, for light-colored blinds (assumed louver reflectance = 0.8 and louvers positioned at 45° angle) on doubleglazed, heat-absorbing windows (Type 5d from Table 13B of Chapter 15), IAC(0) = 0.74, IAC(60) = 0.65, IAC(diff) = 0.79, and radiant fraction = 0.54. Without blinds, IAC = 1.0. Therefore, window heat gain components for hour 15, without blinds, are qb15 = AEt,b SHGC()(IAC) = (3.72)(421)(0.3978)(1.00) = 623 W qd15 = A(Et,d + Er)SHGCD(IAC) = (3.72)(115 + 81)(0.41)(1.00) = 299 W qc15 = UA(tout – tin) = (3.18)(3.72)(33.3 – 23.9) = 111 W This procedure is repeated to determine these values for a 24 h heat gain profile, shown in Table 30. Total cooling load for the window is calculated by summing the convective and radiant portions. For windows with indoor shading (blinds, drapes, etc.), the direct beam, diffuse, and conductive heat gains may be summed and treated together in calculating cooling loads. However, in this example, the window does not have indoor shading, and the direct beam solar heat gain should be treated separately from the diffuse and conductive heat gains. The direct beam heat gain, without indoor shading, is treated as 100% radiant, and solar RTS factors from Table 20 are used to convert the beam heat gains to cooling loads. The diffuse and conductive heat gains can be totaled and split into radiant and convective portions according to Table 14, and nonsolar RTS factors from Table 19 are used to convert the radiant portion to cooling load. The solar beam cooling load is calculated using heat gains for the current hour and past 23 h and radiant time series from Table 20, in accordance with Equation (38). From Table 20, select the solar RTS for medium-weight construction, assuming 50% glass and carpeted floors for this example. Using Table 30 values for direct solar heat gain, the radiant cooling load for the window direct beam solar component is
0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4500 0.4553 0.5306 0.6332 0.7505 0.8644 0.9555 1.0073 1.0100 0.9631 0.8755 0.7630 0.6452 0.5403 0.4618
0.0 0.0 0.0 0.0 0.0 9.6 32.8 48.2 59.0 66.4 72.2 87.2 104.7 121.4 132.6 133.4 120.3 91.6 46.6 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 11.8 48.4 82.7 112.8 137.3 155.8 178.3 197.3 209.4 210.1 195.7 164.0 115.7 53.8 0.0 0.0 0.0 0.0 0.0
0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410 0.410
0 0 0 0 0 18 74 126 172 209 237 272 301 319 320 298 250 176 82 0 0 0 0 0
23.2 22.8 22.4 22.1 21.9 22.1 22.9 24.8 26.9 28.8 30.6 31.7 32.6 33.1 33.1 32.4 31.6 30.4 28.7 27.5 26.5 25.5 24.7 23.9
–8 –13 –17 –21 –24 –21 –12 11 36 58 79 92 102 109 109 101 91 77 57 43 31 19 10 0
–8 –13 –17 –21 –24 –3 62 137 208 268 316 364 485 762 1001 1109 1031 769 323 43 31 19 10 0
Qb,15 = r0qb,15 + r1qb,14 + r2qb,13 + r3qb,12 + … + r23qb,14 = (0.54)(783) + (0.16)(362) + (0.08)(89) + (0.04)(0) + (0.03)(0) + (0.02)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.01)(0) + (0.00)(0) + (0.00)(254) + (0.00)(610) + (0.00)(779) + (0.00)(783) = 488 W This process is repeated for other hours; results are listed in Table 31. For diffuse and conductive heat gains, the radiant fraction according to Table 14 is 46%. The radiant portion is processed using nonsolar RTS coefficients from Table 19. The results are listed in Tables 30 and 31. For 3:00 PM, the diffuse and conductive cooling load is 380 W. The total window cooling load at the designated hour is thus Qwindow = Qb + Qdiff + cond = 488 + 380 = 868 W Again, a computer spreadsheet or other software is commonly used to reduce the effort involved in calculations. The spreadsheet shown in Table 30 is expanded in Table 31 to include splitting the heat gain into convective and radiant portions, applying RTS to the radiant portion, and totaling the convective and radiant loads to determine a 24 h cooling load profile for a window without indoor shading. If the window has an indoor shading device, it is accounted for with the indoor attenuation coefficients (IAC), the radiant fraction, and the RTS type used. If a window has no indoor shading, 100% of the direct beam energy is assumed to be radiant and solar RTS factors are used. However, if an indoor shading device is present, the direct beam is assumed to be interrupted by the shading device, and a portion immediately becomes cooling load by convection. Also, the energy is assumed to be radiated to all surfaces of the room, therefore nonsolar RTS values are used to convert the radiant load into cooling load. IAC values depend on several factors: (1) type of shading device, (2) position of shading device relative to window, (3) reflectivity of shading device, (4) angular adjustment of shading device, as well as (5) solar position relative to the shading device. These factors are discussed
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18.52
2017 ASHRAE Handbook—Fundamentals (SI) Table 31 Window Component of Cooling Load (No Blinds or Overhang) (Month 7) Unshaded Direct Beam Solar Cooling Load (IFAC = 1)
Licensed for single user. © 2017 ASHRAE, Inc.
Beam Local Solar Stan- Heat dard Gain Hour W 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Shaded Direct Beam (AC 1.0) + Diffuse + Conduction Cooling Load
Solar Beam ConducConRTS Solar Diffuse Total tion vective Radiant Zone Radi- Cooling Heat Heat Heat Heat 100%, Type 8, ant, 0%, Load, Gain, Gain, Gain, Gain, W W % W W W W W W
0 0 0 0 0 0 0 0 0 0 0 0 82 334 572 710 690 515 184 0 0 0 0 0
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 0 0 0 0 0 0 0 0 0 0 0 82 334 572 710 690 515 184 0 0 0 0 0
54 16 8 4 3 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0 0.0
31 31 31 31 31 31 31 30 27 21 14 7 46 194 369 505 548 480 290 135 80 54 40 33
31 31 31 31 31 31 31 30 27 21 14 7 46 194 369 505 548 480 290 135 80 54 40 33
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 0 0 0 0 18 74 126 172 209 237 272 301 319 320 298 250 176 82 0 0 0 0 0
in detail in Chapter 15. For this example with venetian blinds, the IAC for beam radiation is treated separately from the diffuse solar gain. The direct beam IAC must be adjusted based on the profile angle of the sun. At 3:00 PM in July, the profile angle of the sun relative to the window surface is 58°. Calculated using Equation (38) from Chapter 15, the beam IAC = 0.653. The diffuse IAC is 0.79. Thus, the window heat gains, with light-colored blinds, at 3:00 PM are qb15 = AED SHGC()(IACb) = (3.72)(421)(0.3978)(0.653) = 406 W qd15 = A(Ed + Er)SHGCD (IACd)= (3.72)(115 + 81)(0.41)(0.79) = 236 W qc15 = UA(tout – tin) = (3.18)(3.72)(33.3 – 23.9) = 111 W Because the same radiant fraction and nonsolar RTS are applied to all parts of the window heat gain when indoor shading is present, those loads can be totaled and the cooling load calculated after splitting the radiant portion for processing with nonsolar RTS. This is shown by the spreadsheet results in Table 32. The total window cooling load with venetian blinds at 3:00 PM = 636 W. Part 4. Window cooling load using radiant time series for window with overhang shading. Calculate the cooling load contribution for the previous example with the addition of a 3 m overhang shading the window. Solution: In Chapter 15, methods are described and examples provided for calculating the area of a window shaded by attached vertical or horizontal projections. For 3:00 PM LST IN July, the solar position calculated in previous examples is Solar altitude = 57.2° Solar azimuth 75.1° Surface-solar azimuth = 15.1° From Chapter 15, Equation (32), profile angle is calculated by tan = tan /cos = tan(57.2)/cos(15.1) = 1.6087 = 58.1° From Chapter 15, Equation (34), shadow height SH is
–8 –13 –17 –21 –24 –21 –12 11 36 58 79 92 102 109 109 101 91 77 57 43 31 19 10 0
–8 –13 –17 –21 –24 –3 62 137 208 268 316 364 403 428 429 399 341 254 139 43 31 19 10 0
Convective Radiant 54%, 46%, W W –4 –7 –9 –11 –13 –2 33 74 112 145 171 196 218 231 232 216 184 137 75 23 17 10 5 0
–4 –6 –8 –10 –11 –1 28 63 96 123 145 167 185 197 197 184 157 117 64 20 14 9 5 0
Nonsolar RTS Zone 8, % 49 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0
Window Radi- Cooling Cooling ant, Load, Load, W W W 18 15 13 10 8 11 25 46 69 91 110 128 145 158 165 163 150 127 94 60 44 34 27 22
13 8 3 –1 –5 9 58 120 181 235 281 325 363 389 397 378 334 264 169 83 61 44 32 22
44 39 34 30 26 40 89 150 208 256 295 332 409 583 766 883 882 745 459 744 141 98 72 54
SH = PH tan = 3.05(1.6087) = 4.9 m Because the window is 1.95 m tall, at 3:00 PM the window is completely shaded by the 3 m deep overhang. Thus, the shaded window heat gain includes only diffuse solar and conduction gains. This is converted to cooling load by separating the radiant portion, applying RTS, and adding the resulting radiant cooling load to the convective portion to determine total cooling load. Those results are in Table 33. The total window cooling load = 322 W. Part 5. Room cooling load total. Calculate the sensible cooling loads for the previously described office at 3:00 PM in July. Solution: The steps in the previous example parts are repeated for each of the internal and external loads components, including the southeastfacing window, spandrel and brick walls, the southwest-facing brick wall, the roof, people, and equipment loads. The results are tabulated in Table 34. The total room sensible cooling load for the office is 1077 W at 3:00 PM in July. When this calculation process is repeated for a 24 h design day for each month, it is found that the peak room sensible cooling load actually occurs in July at hour 14 (2:00 PM solar time) at 1077 W as indicated in Table 35.
Although simple in concept, these steps involved in calculating cooling loads are tedious and repetitive, even using the “simplified” RTS method; practically, they should be performed using a computer spreadsheet or other program. The calculations should be repeated for multiple design conditions (i.e., times of day, other months) to determine the maximum cooling load for mechanical equipment sizing. Example spreadsheets for computing each cooling load component using conduction and radiant time series are available from ASHRAE. To illustrate the full building example discussed previously, those individual component spreadsheets have been compiled to allow calculation of cooling and heating loads on a room by room basis as well as for a “block” calculation for analysis of overall areas or buildings where detailed room-by-room data are not available.
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Nonresidential Cooling and Heating Load Calculations Table 32
Window Component of Cooling Load (with Blinds, No Overhang) (Month 7)
Unshaded Direct Beam Solar Cooling Load (IF AC = 1)
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Solar Beam Local Solar ConRTS Stan- Heat vective Radiant Zone 100%, Type 8, dard Gain, 0%, W W Hour W % 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
Table 33
1 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
Shaded Direct Beam (AC 1.0) + Diffuse + Conduction Cooling Load
Beam ConSolar Diffuse duction Total Radi- Cooling Heat Heat Heat Heat ant, Load, Gain, Gain, Gain, Gain, W W W W W W 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0
0 0 0 0 0 0 0 0 0 0 0 0 53 217 373 474 472 361 133 0 0 0 0 0
Local Direct Stan- Surface Shadow Shadow Sunlit dard Solar Profile Width, Height, Area, Hour Azimuth Angle m m2 m –235 –219 –204 –192 –182 –173 –165 –158 –150 –141 –127 –99 –44 –3 15 26 34 42 49 57 67 78 91 107
52 40 29 19 9 –3 –15 –28 –43 –58 –73 –87 80 69 58 48 38 26 12 –6 –32 –64 87 67
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0 0 0 0 0 14 58 100 136 165 188 215 238 252 253 236 197 139 65 0 0 0 0 0
–8 –13 –17 –21 –24 –21 –12 11 36 58 79 92 102 109 109 101 91 77 57 43 31 19 10 0
–8 –13 –17 –21 –24 –7 46 111 172 224 266 307 393 578 735 811 760 578 254 43 31 19 10 0
Convective Radiant 46%, 54%, W W –4 –6 –8 –10 –11 –3 21 51 79 103 123 141 181 266 338 373 350 266 117 20 14 9 5 0
–4 –7 –9 –11 –13 –4 25 60 93 121 144 166 212 312 397 438 410 312 137 23 17 10 5 0
Nonsolar RTS Zone 8, % 49 17 9 5 3 2 2 1 1 1 1 1 1 1 1 1 1 1 1 1 0 0 0 0
Window Radi- Cooling Cooling Load, ant, Load, W W W 29 25 23 20 18 20 34 55 77 98 116 132 160 220 285 333 342 303 207 118 80 58 45 36
26 19 15 10 7 17 55 106 156 201 239 273 341 486 623 706 692 569 324 138 94 67 50 36
26 19 15 10 7 17 55 106 156 201 239 273 341 486 623 706 692 569 324 138 94 67 50 36
Window Component of Cooling Load (with Blinds and Overhang) (Month 7) Shaded Direct Beam (AC 1.0) + Diffuse + Conduction Cooling Load
Overhang and Fins Shading Calculations
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
18.53
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 2.0 2.0 2.0 2.0 2.0 1.5 0.7 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.9 2.5 0.0 0.0 0.0 0.0 0.0
Beam Solar Heat Gain, W
Diffuse Heat Gain, W
Conduction Heat Gain, W
Total Heat Gain, W
0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 85 88 0 0 0 0 0
0 0 0 0 0 14 58 100 136 165 188 215 238 252 253 236 197 139 65 0 0 0 0 0
–8 –13 –17 –21 –24 –21 –12 11 36 58 79 92 102 109 109 101 91 77 57 43 31 19 10 0
–8 –13 –17 –21 –24 –7 46 111 172 224 266 307 340 361 362 337 288 302 210 43 31 19 10 0
Convect. Radiant % % 46%, 54%, W W –4 –7 –9 –11 –13 –4 25 60 93 121 144 166 184 195 195 182 156 163 113 23 17 10 5 0
–4 –6 –8 –10 –11 –3 21 51 79 103 123 141 156 166 166 155 133 139 96 20 14 9 5 0
NonsoWindow lar RTS Cooling Cooling Load, Zone Radiant, Load, W 8, % W W 49% 17% 9% 5% 3% 2% 2% 1% 1% 1% 1% 1% 1% 1% 1% 1% 1% 1% 1% 1% 0% 0% 0% 0%
16 13 11 8 6 8 20 38 57 76 93 109 123 134 139 137 127 127 107 63 44 33 26 20
12 6 1 –3 –7 5 45 98 150 197 237 274 307 329 334 319 282 290 220 86 61 43 31 20
12 6 1 –3 –7 5 45 98 150 197 237 274 307 329 334 319 282 290 220 86 61 43 31 20
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Table 34 Single-Room Example Cooling Load (July 3:00 PM) for ASHRAE Example Office Building, Atlanta, GA
Table 35 Single-Room Example Peak Cooling Load (Sept. 5:00 PM) for ASHRAE Example Office Building, Atlanta, GA
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‘
9.2
SINGLE-ROOM EXAMPLE PEAK HEATING LOAD
Although the physics of heat transfer that creates a heating load is identical to that for cooling loads, a number of traditionally used simplifying assumptions facilitate a much simpler calculation procedure. As described in the Heating Load Calculations section, design heating load calculations typically assume a single outdoor temperature, with no heat gain from solar or internal sources, under steady-state conditions. Thus, space heating load is determined by computing the heat transfer rate through building envelope elements (UAT) plus heat required because of outdoor air infiltration. Part 6. Room heating load. Calculate the room heating load for the previous described office, including infiltration airflow at one air change per hour. Solution: Because solar heat gain is not considered in calculating design heating loads, orientation of similar envelope elements may be ignored and total areas of each wall or window type combined. Thus, the total spandrel wall area = 5.57 + 5.57 = 11.14 m2, total brick wall area = 5.57 + 3.72 = 9.29 m2, and total window area = 3.72 + 3.72 = 7.44 m2. For this example, use the U-factors that were used for cooling load conditions. In some climates, higher prevalent winds in winter
should be considered in calculating U-factors (see Chapter 25 for information on calculating U-factors and surface heat transfer coefficients appropriate for local wind conditions). The 99.6% heating design dry-bulb temperature for Atlanta is –5.6°C and the indoor design temperature is 22.2°C. The room volume with a 2.74 m ceiling = 2.74 12.7 = 33.2 m3 = 33 200 L. At one air change per hour, the infiltration airflow = 32 200/3600 = 9.2 L/s. Thus, the heating load is Windows: Spandrel wall: Brick wall: Roof: Infiltration:
3.18 7.44[22.2 – (–5.6)] 0.4411.14[22.2 – (–5.6)] 0.459.29[22.2 – (–5.6)] 0.1812.08[22.2 – (–5.6)] 9.21.23[22.2 – (–5.6)]
= = = = =
Total room heating load:
9.3
658 W 136 116 60 315 1285 W
WHOLE-BUILDING EXAMPLE
Because a single-room example does not illustrate the full application of load calculations, a multistory, multiple-room example building has been developed to show a more realistic case. A hypothetical project development process is described to illustrate its effect on the application of load calculations.
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Nonresidential Cooling and Heating Load Calculations
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Table 36 Block Load Example: Envelope Area Summary, m2 Brick Areas
Spandrel/Soffit Areas
North
South
East
West
North
South
East
West
North
South
East
West
First floor Second floor
1765 1459
63.17 47.38
52.02 36.23
37.16 27.87
37.16 27.87
130.6 96.62
125.40 85.47
96.62 50.17
33.44 50.17
55.74 52.02
92.90 78.04
11.15 33.44
33.44 33.44
Building total
3224
110.55
88.25
65.03
65.03
227.22
210.87
146.79
83.61
107.76
170.94
44.59
66.89
Design Process and Shell Building Definition
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Window Areas
Floor Area
A development company has acquired a piece of property in Atlanta, GA, to construct an office building. Although no tenant or end user has yet been identified, the owner/developer has decided to proceed with the project on a speculative basis. They select an architectural design firm, who retains an engineering firm for the mechanical and electrical design. At the first meeting, the developer indicates the project is to proceed on a fast-track basis to take advantage of market conditions; he is negotiating with several potential tenants who will need to occupy the new building within a year. This requires preparing shell-andcore construction documents to obtain a building permit, order equipment, and begin construction to meet the schedule. The shell-and-core design documents will include finished design of the building exterior (the shell), as well as permanent interior elements such as stairs, restrooms, elevator, electrical rooms and mechanical spaces (the core). The primary mechanical equipment must be sized and installed as part of the shell-and-core package in order for the project to meet the schedule, even though the building occupant is not yet known. The architect selects a two-story design with an exterior skin of tinted, double-glazed vision glass; opaque, insulated spandrel glass, and brick pilasters. The roof area extends beyond the building edge to form a substantial overhang, shading the second-floor windows. Architectural drawings for the shell-and-core package (see Figures 17 to 22) include plans, elevations, and skin construction details, and are furnished to the engineer for use in “block” heating and cooling load calculations. Mechanical systems and equipment must be specified and installed based on those calculations. (Note: Fullsize, scalable electronic versions of the drawings in Figures 17 to 22, as well as detailed lighting plans, are available from ASHRAE at www.ashrae.org and in the ASHRAE Handbook Online version of this chapter, on the Additional Features tab.) The HVAC design engineer meets with the developer’s operations staff to agree on the basic HVAC systems for the project. Based on their experience operating other buildings and the lack of specific information on the tenant(s), the team decides on two variablevolume air-handling units (AHUs), one per floor, to provide operating flexibility if one floor is leased to one tenant and the other floor to someone else. Cooling will be provided by an air-cooled chiller located on grade across the parking lot. Heating will be provided by electric resistance heaters in parallel-type fan-powered variable-airvolume (VAV) terminal units. The AHUs must be sized quickly to confirm the size of the mechanical rooms on the architectural plans. The AHUs and chiller must be ordered by the mechanical subcontractor within 10 days to meet the construction schedule. Likewise, the electric heating loads must be provided to the electrical engineers to size the electrical service and for the utility company to extend services to the site. The mechanical engineer must determine the (1) peak airflow and cooling coil capacity for each AHU, (2) peak cooling capacity required for the chiller, and (3) total heating capacity for sizing the electrical service. Solution: First, calculate “block” heating and cooling loads for each floor to size the AHUs, then calculate a block load for the whole building determine chiller and electric heating capacity.
Based on the architectural drawings, the HVAC engineer assembles basic data on the building as follows: Location: Atlanta, GA. Per Chapter 14, latitude = 33.64, longitude = 84.43, elevation = 313 m above sea level, 99.6% heating design dry-bulb temperature = –5.6°C. For cooling load calculations, use 5% dry-bulb/coincident wet-bulb monthly design day profile from Chapter 14 (on CD-ROM). See Table 27 for temperature profiles used in these examples. Indoor design conditions: 22.2°C for heating; 23.9°C with 50% rh for cooling. Building orientation: Plan north is 30° west of true north. Gross area per floor: 1765 m2 first floor and 1459 m2 second floor. Total building gross area: 3224 m2. Windows: Bronze-tinted, double-glazed. Solar heat gain coefficients, U-factors are as in the single-room example. Walls: Part insulated spandrel glass and part brick-and-block clad columns. The insulation barrier in the soffit at the second floor is similar to that of the spandrel glass and is of lightweight construction; for simplicity, that surface is assumed to have similar thermal heat gain/loss to the spandrel glass. Construction and insulation values are as in single-room example. Roof: Metal deck, topped with board insulation and membrane roofing. Construction and insulation values are as in the singleroom example. Floor: 127 mm lightweight concrete slab on grade for first floor and 127 mm lightweight concrete on metal deck for second floor Total areas of building exterior skin, as measured from the architectural plans, are listed in Table 36. The engineer needs additional data to estimate the building loads. Thus far, no tenant has yet been signed, so no interior layouts for population counts, lighting layouts, or equipment loads are available. To meet the schedule, assumptions must be made on these load components. The owner requires that the system design must be flexible enough to provide for a variety of tenants over the building’s life. Based on similar office buildings, the team agrees to base the block load calculations on the following assumptions: Occupancy: 7.54 people per 100 m2 = 13.3 m2/person Lighting: 11.8 W/m2 Tenant’s office equipment: 10.76 W/m2 Normal use schedule is assumed at 100% from 7:00 AM to 7:00 PM and unoccupied/off during other hours. With interior finishes not finalized, the owner commits to using light-colored interior blinds on all windows. The tenant interior design could include carpeted flooring or acoustical tile ceilings in all areas, but the more conservative assumption, from a peak load standpoint, is chosen: carpeted flooring and no acoustical tile ceilings (no ceiling return plenum). For block loads, the engineer assumes that the building is maintained under positive pressure during peak cooling conditions and that infiltration during peak heating conditions is equivalent to one air change per hour in a 3.5 m deep perimeter zone around the building. To maintain indoor air quality, outdoor air must be introduced into the building. Air will be ducted from roof intake hoods to the AHUs where it will be mixed with return air before being cooled and dehumidified by the AHU’s cooling coil. ASHRAE Standard 62.1 is
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2017 ASHRAE Handbook—Fundamentals (SI) Table 38 Block Load Example—Second Floor Loads for ASHRAE Example Office Building, Atlanta, GA
the design basis for ventilation rates; however, no interior tenant layout is available for application of Standard 62.1 procedures. Based on past experience, the engineer decides to use 9.44 L/s of outdoor air per person for sizing the cooling coils and chiller. Block load calculations were performed using the RTS method, and results for the first and second floors and the entire building are summarized in Tables 37, 38, and 39. Based on these results, the engineer performs psychrometric coil analysis, checks capacities versus vendor catalog data, and prepares specifications and schedules for the equipment. This information is released to the contractor with the shell-and-core design documents. The air-handling units and chiller are purchased, and construction proceeds.
documents. The first step in this process is to prepare room-byroom peak heating and cooling load calculations, which will then be used for design of the air distribution systems from each of the VAV air handlers already installed. The HVAC engineer must perform a room-by-room “takeoff” of the architect’s drawings. For each room, this effort identifies the floor area, room function, exterior envelope elements and areas, number of occupants, and lighting and equipment loads. The tenant layout calls for a dropped acoustical tile ceiling throughout, which will be used as a return air plenum. Typical 600 by 1200 mm fluorescent, recessed, return-air-type lighting fixtures are selected. Based on this, the engineer assumes that 20% of the heat gain from lighting will be to the return air plenum and not enter rooms directly. Likewise, some portion of the heat gain from the roof will be extracted via the ceiling return air plenum. From experience, the engineer understands that return air plenum paths are not always predictable, and decides to credit only 30% of the roof heat gain to the return air, with the balance included in the room cooling load. For the open office areas, some areas along the building perimeter will have different load characteristics from purely interior spaces because of heat gains and losses through the building skin.
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Table 37 Block Load Example—First Floor Loads for ASHRAE Example Office Building, Atlanta, GA
Tenant Fit Design Process and Definition About halfway through construction, a tenant agrees to lease the entire building. The tenant will require a combination of open and enclosed office space with a few common areas, such as conference/ training rooms, and a small computer room that will operate on a 24 h basis. Based on the tenant’s space program, the architect prepares interior floor plans and furniture layout plans, and the electrical engineer prepares lighting design plans. Those drawings are furnished to the HVAC engineer to prepare detailed design
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Nonresidential Cooling and Heating Load Calculations Table 39 Block Load Example—Overall Building Loads for ASHRAE Example Office Building, Atlanta, GA
18.57 cooling loads for a single point in time may miss the peak and result in inadequate cooling for that room. • Often, in real design processes, not all data are known. Reasonable assumptions based on past experience must be made. • Heating and air-conditioning systems often serve spaces whose use changes over the life of a building. Assumptions used in heating and cooling load calculations should consider reasonable possible uses over the life of the building, not just the first use of the space. • The relative importance of each cooling and heating load component varies, depending on the portion of the building being considered. Characteristics of a particular window may have little effect on the entire building load, but could have a significant effect on the supply airflow to the room where the window is located and thus on the comfort of the occupants of that space.
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10.
Although those perimeter areas are not separated from other open office spaces by walls, the engineer knows from experience that they must be served by separate control zones to maintain comfort conditions.
Room-by-Room Cooling and Heating Loads The room-by-room results of RTS method calculations, including the month and time of day of each room’s peak cooling load, as well as peak heating loads for each room and all input data, are available at www.ashrae.org and in the ASHRAE Handbook Online version of this chapter (on the Additional Features tab) in spreadsheet format similar to Table 39. These results are used by the HVAC engineer to select and design room air distribution devices and to schedule airflow rates for each space. That information is incorporated into the tenant fit drawings and specifications issued to the contractor.
Conclusions The example results illustrate issues that should be understood and accounted for in calculating heating and cooling loads: • First, peak room cooling loads occur at different months and times, depending on the exterior exposure of the room. Calculation of
PREVIOUS COOLING LOAD CALCULATION METHODS
Procedures described in this chapter are the most current and scientifically derived means for estimating cooling load for a defined building space, but methods in earlier editions of the ASHRAE Handbook are valid for many applications. These earlier procedures are simplifications of the heat balance principles, and their use requires experience to deal with atypical or unusual circumstances. In fact, any cooling or heating load estimate is no better than the assumptions used to define conditions and parameters such as physical makeup of the various envelope surfaces, conditions of occupancy and use, and ambient weather conditions. Experience of the practitioner can never be ignored. The primary difference between the HB and RTS methods and the older methods is the newer methods’ direct approach, compared to the simplifications necessitated by the limited computer capability available previously. The transfer function method (TFM), for example, required many calculation steps. It was originally designed for energy analysis with emphasis on daily, monthly, and annual energy use, and thus was more oriented to average hourly cooling loads than peak design loads. The total equivalent temperature differential method with time averaging (TETD/TA) has been a highly reliable (if subjective) method of load estimating since its initial presentation in the 1967 Handbook of Fundamentals. Originally intended as a manual method of calculation, it proved suitable only as a computer application because of the need to calculate an extended profile of hourly heat gain values, from which radiant components had to be averaged over a time representative of the general mass of the building involved. Because perception of thermal storage characteristics of a given building is almost entirely subjective, with little specific information for the user to judge variations, the TETD/TA method’s primary usefulness has always been to the experienced engineer. The cooling load temperature differential method with solar cooling load factors (CLTD/CLF) attempted to simplify the twostep TFM and TETD/TA methods into a single-step technique that proceeded directly from raw data to cooling load without intermediate conversion of radiant heat gain to cooling load. A series of factors were taken from cooling load calculation results (produced by more sophisticated methods) as “cooling load temperature differences” and “cooling load factors” for use in traditional conduction (q = UAt) equations. The results are approximate cooling load values rather than simple heat gain values. The simplifications and assumptions used in the original work to derive those factors limit this method’s applicability to those building types and conditions for which the CLTD/CLF factors were derived; the method should not be used beyond the range of applicability.
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2017 ASHRAE Handbook—Fundamentals (SI)
Although the TFM, TETD/TA, and CLTD/CLF procedures are not republished in this chapter, those methods are not invalidated or discredited. Experienced engineers have successfully used them in millions of buildings around the world. The accuracy of cooling load calculations in practice depends primarily on the availability of accurate information and the design engineer’s judgment in the assumptions made in interpreting the available data. Those factors have much greater influence on a project’s success than does the choice of a particular cooling load calculation method. The primary benefit of HB and RTS calculations is their somewhat reduced dependency on purely subjective input (e.g., determining a proper time-averaging period for TETD/TA; ascertaining appropriate safety factors to add to the rounded-off TFM results; determining whether CLTD/CLF factors are applicable to a specific unique application). However, using the most up-to-date techniques in real-world design still requires judgment on the part of the design engineer and care in choosing appropriate assumptions, just as in applying older calculation methods.
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REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. Abushakra, B., J.S. Haberl, and D.E. Claridge. 2004. Overview of literature on diversity factors and schedules for energy and cooling load calculations (1093-RP). ASHRAE Transactions 110(1):164-176. Armstrong, P.R., C.E. Hancock, III, and J.E. Seem. 1992a. Commercial building temperature recovery—Part I: Design procedure based on a step response model. ASHRAE Transactions 98(1):381-396. Armstrong, P.R., C.E. Hancock, III, and J.E. Seem. 1992b. Commercial building temperature recovery—Part II: Experiments to verify the step response model. ASHRAE Transactions 98(1):397-410. ASHRAE. 2010. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2010. ASHRAE. 2010. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2010. ASHRAE. 2016. Energy standard for building except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2016. ASHRAE. 2012. Updating the climatic design conditions in the ASHRAE Handbook—Fundamentals (RP-1613). ASHRAE Research Project, Final Report. ASHRAE. 2013. Underfloor air distribution (UFAD) design guide, 2nd ed. ASTM. 2008. Practice for estimate of the heat gain or loss and the surface temperatures of insulated flat, cylindrical, and spherical systems by use of computer programs. Standard C680-08. American Society for Testing and Materials, West Conshohocken, PA. Bach, C., and O. Sarfraz. 2017. Update to measurements of office heat gain data. ASHRAE Research Project RP-1742, Progress Report. Bauman, F., T. Webster, P. Linden, and F. Buhl. 2007. Energy performance of UFAD systems. CEC-500-2007-050, Final Report to CEC PIER Buildings Program. Center for the Built Environment, University of California, Berkeley. www.energy.ca.gov/2007publications/CEC-500-2007 -050/CEC-500-2007-050.PDF. Bauman, F., S. Schiavon, T. Webster, and K.H. Lee. 2010. Cooling load design tool for UFAD systems. ASHRAE Journal (September):62-71. escholarship.org/uc/item/9d8430v3. Bliss, R.J.V. 1961. Atmospheric radiation near the surface of the ground. Solar Energy 5(3):103. CFR. Annual. Energy efficiency program for certain commercial and industrial equipment. Code of Federal Regulations 10 CFR 431. U.S. Government Publishing Office, Washington, D.C. www.ecfr.gov. Chantrasrisalai, C., D.E. Fisher, I. Iu, and D. Eldridge. 2003. Experimental validation of design cooling load procedures: The heat balance method. ASHRAE Transactions 109(2):160-173. Claridge, D.E., B. Abushakra, J.S. Haberl, and A. Sreshthaputra. 2004. Electricity diversity profiles for energy simulation of office buildings (RP1093). ASHRAE Transactions 110(1):365-377.
Eldridge, D., D.E. Fisher, I. Iu, and C. Chantrasrisalai. 2003. Experimental validation of design cooling load procedures: Facility design (RP-1117). ASHRAE Transactions 109(2):151-159. Feng, J., S. Schiavon, and F. Bauman. 2012. Comparison of zone cooling load for radiant and air conditioning systems. Proceedings of the International Conference on Building Energy and Environment. Boulder, CO. escholarship.org/uc/item/9g24f38j. Fisher, D.R. 1998. New recommended heat gains for commercial cooking equipment. ASHRAE Transactions 104(2):953-960. Fisher, D.E., and C. Chantrasrisalai. 2006. Lighting heat gain distribution in buildings (RP-1282). ASHRAE Research Project, Final Report. Fisher, D.E., and C.O. Pedersen. 1997. Convective heat transfer in building energy and thermal load calculations. ASHRAE Transactions 103(2): 137-148. Gordon, E.B., D.J. Horton, and F.A. Parvin. 1994. Development and application of a standard test method for the performance of exhaust hoods with commercial cooking appliances. ASHRAE Transactions 100(2). Hittle, D.C. 1999. The effect of beam solar radiation on peak cooling loads. ASHRAE Transactions 105(2):510-513. Hittle, D.C., and C.O. Pedersen. 1981. Calculating building heating loads using the frequency of multi-layered slabs. ASHRAE Transactions 87(2): 545-568. Hosni, M.H., and B.T. Beck. 2008. Update to measurements of office equipment heat gain data (RP-1482). ASHRAE Research Project, Progress Report. Hosni, M.H., B.W. Jones, J.M. Sipes, and Y. Xu. 1998. Total heat gain and the split between radiant and convective heat gain from office and laboratory equipment in buildings. ASHRAE Transactions 104(1A):356-365. Hosni, M.H., B.W. Jones, and H. Xu. 1999. Experimental results for heat gain and radiant/convective split from equipment in buildings. ASHRAE Transactions 105(2):527-539. Incropera, F.P., and D.P. DeWitt. 1990. Fundamentals of heat and mass transfer, 3rd ed. Wiley, New York. Iu, I., and D.E. Fisher. 2004. Application of conduction transfer functions and periodic response factors in cooling load calculation procedures. ASHRAE Transactions 110(2):829-841. Iu, I., C. Chantrasrisalai, D.S. Eldridge, and D.E. Fisher. 2003. Experimental validation of design cooling load procedures: The radiant time series method (RP-1117). ASHRAE Transactions 109(2):139-150. Jones, B.W., M.H. Hosni, and J.M. Sipes. 1998. Measurement of radiant heat gain from office equipment using a scanning radiometer. ASHRAE Transactions 104(1B):1775-1783. Karambakkam, B.K., B. Nigusse, and J.D. Spitler. 2005. A one-dimensional approximation for transient multi-dimensional conduction heat transfer in building envelopes. Proceedings of the 7th Symposium on Building Physics in the Nordic Countries, The Icelandic Building Research Institute, Reykjavik, vol. 1, pp. 340-347. Kerrisk, J.F., N.M. Schnurr, J.E. Moore, and B.D. Hunn. 1981. The custom weighting-factor method for thermal load calculations in the DOE-2 computer program. ASHRAE Transactions 87(2):569-584. Komor, P. 1997. Space cooling demands from office plug loads. ASHRAE Journal 39(12):41-44. Kong, M., and J. Zhang. 2016. Life-cycle cost and benefit analysis of utilizing hoods for light-duty cooking appliances in commercial kitchens (RP1631, part 2). Science and Technology for the Built Environment 22(6): 866-882. Kong, M., J. Zhang, B. Guo, and K. Han. 2016. Measurements of grease emission and heat generation rates of electric countertop appliances (RP1631, part 1). Science and Technology for the Built Environment 22(6): 845-865. Kusuda, T. 1967. NBSLD, the computer program for heating and cooling loads for buildings. BSS 69 and NBSIR 74-574. National Bureau of Standards. Latta, J.K., and G.G. Boileau. 1969. Heat losses from house basements. Canadian Building 19(10):39. LBL. 2015. WINDOW 7.4.6: A PC program for analyzing window thermal performance for fenestration products. LBL-48255. Windows and Daylighting Group. Lawrence Berkeley Laboratory, Berkeley. Liesen, R.J., and C.O. Pedersen. 1997. An evaluation of inside surface heat balance models for cooling load calculations. ASHRAE Transactions 103(2):485-502. Marn, W.L. 1962. Commercial gas kitchen ventilation studies. Research Bulletin 90(March). Gas Association Laboratories, Cleveland, OH.
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Nonresidential Cooling and Heating Load Calculations McClellan, T.M., and C.O. Pedersen. 1997. Investigation of outdoor heat balance models for use in a heat balance cooling load calculation procedure. ASHRAE Transactions 103(2):469-484. McQuiston, F.C., and J.D. Spitler. 1992. Cooling and heating load calculation manual, 2nd ed. ASHRAE. Miller, A. 1971. Meteorology, 2nd ed. Charles E. Merrill, Columbus. Nigusse, B.A. 2007. Improvements to the radiant time series method cooling load calculation procedure. Ph.D. dissertation, Oklahoma State University. Parker, D.S., J.E.R. McIlvaine, S.F. Barkaszi, D.J. Beal, and M.T. Anello. 2000. Laboratory testing of the reflectance properties of roofing material. FSEC-CR670-00. Florida Solar Energy Center, Cocoa. Pedersen, C.O., D.E. Fisher, and R.J. Liesen. 1997. Development of a heat balance procedure for calculating cooling loads. ASHRAE Transactions 103(2):459-468. Pedersen, C.O., D.E. Fisher, J.D. Spitler, and R.J. Liesen. 1998. Cooling and heating load calculation principles. ASHRAE. PG&E. 2010-2016. Dishwashing machine performance reports: Application of ASTM F2474, standard test method for heat gain to space performance of commercial kitchen ventilation/appliance systems. PG&E Food Service Technology Center, San Ramon, CA. www.fishnick.com /publications/appliancereports/dishmachines/. Rees, S.J., J.D. Spitler, M.G. Davies, and P. Haves. 2000. Qualitative comparison of North American and U.K. cooling load calculation methods. International Journal of Heating, Ventilating, Air-Conditioning and Refrigerating Research 6(1):75-99. Rock, B.A. 2005. A user-friendly model and coefficients for slab-on-grade load and energy calculation. ASHRAE Transactions 111(2):122-136. Rock, B.A., and D.J. Wolfe. 1997. A sensitivity study of floor and ceiling plenum energy model parameters. ASHRAE Transactions 103(1):16-30. Schiavon, S., F. Bauman, K.H. Lee, and T. Webster. 2010a. Simplified calculation method for design cooling loads in underfloor air distribution (UFAD) systems. Energy and Buildings 43(1-2):517-528. escholarship .org/uc/item/5w53c7kr. Schiavon, S., K.H. Lee, F. Bauman, and T. Webster. 2010b. Influence of raised floor on zone design cooling load in commercial buildings. Energy and Buildings 42(5):1182-1191. escholarship.org/uc/item/2bv611dt. Schiavon, S., F. Bauman, K.H. Lee, and T. Webster. 2010c. Development of a simplified cooling load design tool for underfloor air distribution systems. Final Report to CEC PIER Program, July. escholarship.org/uc/item /6278m12z. Smith, V.A., R.T. Swierczyna, and C.N. Claar. 1995. Application and enhancement of the standard test method for the performance of commercial kitchen ventilation systems. ASHRAE Transactions 101(2). Sowell, E.F. 1988a. Cross-check and modification of the DOE-2 program for calculation of zone weighting factors. ASHRAE Transactions 94(2). Sowell, E.F. 1988b. Load calculations for 200,640 zones. ASHRAE Transactions 94(2):716-736. Spitler, J.D., and D.E. Fisher. 1999a. Development of periodic response factors for use with the radiant time series method. ASHRAE Transactions 105(2):491-509. Spitler, J.D., and D.E. Fisher. 1999b. On the relationship between the radiant time series and transfer function methods for design cooling load calculations. International Journal of Heating, Ventilating, Air-Conditioning and Refrigerating Research (now Science and Technology for the Built Environment) 5(2):125-138. Spitler, J.D., D.E. Fisher, and C.O. Pedersen. 1997. The radiant time series cooling load calculation procedure. ASHRAE Transactions 103(2). Spitler, J.D., S.J. Rees, and P. Haves. 1998. Quantitive comparison of North American and U.K. cooling load calculation procedures—Part 1: Methodology, Part II: Results. ASHRAE Transactions 104(2):36-46, 47-61. Sun, T.-Y. 1968. Shadow area equations for window overhangs and side-fins and their application in computer calculation. ASHRAE Transactions 74(1):I-1.1 to I-1.9. Swierczyna, R., P. Sobiski, and D. Fisher. 2008. Revised heat gain and capture and containment exhaust rates from typical commercial cooking appliances (RP-1362). ASHRAE Research Project, Final Report. Swierczyna, R., P.A. Sobiski, and D.R. Fisher. 2009. Revised heat gain rates from typical commercial cooking appliances from RP-1362. ASHRAE Transactions 115(2):138-160. Talbert, S.G., L.J. Canigan, and J.A. Eibling. 1973. An experimental study of ventilation requirements of commercial electric kitchens. ASHRAE Transactions 79(1):34.
18.59 Walton, G. 1983. Thermal analysis research program reference manual. National Bureau of Standards. Webster, T., F. Bauman, F. Buhl, and A. Daly. 2008. Modeling of underfloor air distribution (UFAD) systems. SimBuild 2008, University of California, Berkeley. Wilkins, C.K., and M.R. Cook. 1999. Cooling loads in laboratories. ASHRAE Transactions 105(1):744-749. Wilkins, C.K., and M.H. Hosni. 2000. Heat gain from office equipment. ASHRAE Journal 42(6):33-44. Wilkins, C.K., and M. Hosni. 2011. Plug load design factors. ASHRAE Journal 53(5):30-34. Wilkins, C.K., and N. McGaffin. 1994. Measuring computer equipment loads in office buildings. ASHRAE Journal 36(8):21-24. Wilkins, C.K., R. Kosonen, and T. Laine. 1991. An analysis of office equipment load factors. ASHRAE Journal 33(9):38-44. Zhou, X., S.J. Lochhead, Z. Zhong, and C.V. Huynh. 2016. Low energy LED lighting heat distribution in buildings. ASHRAE Research Project RP1681, Final Report.
BIBLIOGRAPHY Alereza, T., and J.P. Breen, III. 1984. Estimates of recommended heat gain due to commercial appliances and equipment. ASHRAE Transactions 90(2A):25-58. ASHRAE. 1975. Procedure for determining heating and cooling loads for computerized energy calculations, algorithms for building heat transfer subroutines. ASHRAE. 1979. Cooling and heating load calculation manual. BLAST Support Office. 1991. BLAST user reference. University of Illinois, Urbana–Champaign. Buffington, D.E. 1975. Heat gain by conduction through exterior walls and roofs—Transmission matrix method. ASHRAE Transactions 81(2):89. Burch, D.M., B.A. Peavy, and F.J. Powell. 1974. Experimental validation of the NBS load and indoor temperature prediction model. ASHRAE Transactions 80(2):291. Burch, D.M., J.E. Seem, G.N. Walton, and B.A. Licitra. 1992. Dynamic evaluation of thermal bridges in a typical office building. ASHRAE Transactions 98:291-304. Butler, R. 1984. The computation of heat flows through multi-layer slabs. Building and Environment 19(3):197-206. Ceylan, H.T., and G.E. Myers. 1985. Application of response-coefficient method to heat-conduction transients. ASHRAE Transactions 91:30-39. Chiles, D.C., and E.F. Sowell. 1984. A counter-intuitive effect of mass on zone cooling load response. ASHRAE Transactions 91(2A):201-208. Chorpening, B.T. 1997. The sensitivity of cooling load calculations to window solar transmission models. ASHRAE Transactions 103(1). Clarke, J.A. 1985. Energy simulation in building design. Adam Hilger Ltd., Boston. Davies, M.G. 1996. A time-domain estimation of wall conduction transfer function coefficients. ASHRAE Transactions 102(1):328-208. Falconer, D.R., E.F. Sowell, J.D. Spitler, and B.B. Todorovich. 1993. Electronic tables for the ASHRAE load calculation manual. ASHRAE Transactions 99(1):193-200. Harris, S.M., and F.C. McQuiston. 1988. A study to categorize walls and roofs on the basis of thermal response. ASHRAE Transactions 94(2):688-714. Hittle, D.C. 1981. Calculating building heating and cooling loads using the frequency response of multilayered slabs, Ph.D. dissertation, Department of Mechanical and Industrial Engineering, University of Illinois, Urbana–Champaign. Hittle, D.C., and R. Bishop. 1983. An improved root-finding procedure for use in calculating transient heat flow through multilayered slabs. International Journal of Heat and Mass Transfer 26:1685-1693. Kimura and Stephenson. 1968. Theoretical study of cooling loads caused by lights. ASHRAE Transactions 74(2):189-197. Kusuda, T. 1969. Thermal response factors for multilayer structures of various heat conduction systems. ASHRAE Transactions 75(1):246. Mast, W.D. 1972. Comparison between measured and calculated hour heating and cooling loads for an instrumented building. ASHRAE Symposium Bulletin 72(2). McBridge, M.F., C.D. Jones, W.D. Mast, and C.F. Sepsey. 1975. Field validation test of the hourly load program developed from the ASHRAE algorithms. ASHRAE Transactions 1(1):291.
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18.60 Mitalas, G.P. 1968. Calculations of transient heat flow through walls and roofs. ASHRAE Transactions 74(2):182-188. Mitalas, G.P. 1969. An experimental check on the weighting factor method of calculating room cooling load. ASHRAE Transactions 75(2):22. Mitalas, G.P. 1972. Transfer function method of calculating cooling loads, heat extraction rate, and space temperature. ASHRAE Journal 14(12):52. Mitalas, G.P. 1973. Calculating cooling load caused by lights. ASHRAE Transactions 75(6):7. Mitalas, G.P. 1978. Comments on the Z-transfer function method for calculating heat transfer in buildings. ASHRAE Transactions 84(1):667-674. Mitalas, G.P., and J.G. Arsenault. 1970. Fortran IV program to calculate Ztransfer functions for the calculation of transient heat transfer through walls and roofs. Use of Computers for Environmental Engineering Related to Buildings, pp. 633-668. National Bureau of Standards, Gaithersburg, MD. Mitalas, G.P., and K. Kimura. 1971. A calorimeter to determine cooling load caused by lights. ASHRAE Transactions 77(2):65. Mitalas, G.P., and D.G. Stephenson. 1967. Room thermal response factors. ASHRAE Transactions 73(2):III.2.1. Nevins, R.G., H.E. Straub, and H.D. Ball. 1971. Thermal analysis of heat removal troffers. ASHRAE Transactions 77(2):58-72. NFPA. 2012. Health care facilities code. Standard 99-2012. National Fire Protection Association, Quincy, MA. Ouyang, K., and F. Haghighat. 1991. A procedure for calculating thermal response factors of multi-layer walls—State space method. Building and Environment 26(2):173-177. Peavy, B.A. 1978. A note on response factors and conduction transfer functions. ASHRAE Transactions 84(1):688-690. Peavy, B.A., F.J. Powell, and D.M. Burch. 1975. Dynamic thermal performance of an experimental masonry building. NBS Building Science Series 45 (July). Romine, T.B., Jr. 1992. Cooling load calculation: Art or science? ASHRAE Journal, 34(1):14. Rudoy, W. 1979. Don’t turn the tables. ASHRAE Journal 21(7):62. Rudoy, W., and F. Duran. 1975. Development of an improved cooling load calculation method. ASHRAE Transactions 81(2):19-69. Seem, J.E., S.A. Klein, W.A. Beckman, and J.W. Mitchell. 1989. Transfer functions for efficient calculation of multidimensional transient heat transfer. Journal of Heat Transfer 111:5-12. Sowell, E.F., and D.C. Chiles. 1984a. Characterization of zone dynamic response for CLF/CLTD tables. ASHRAE Transactions 91(2A):162-178.
2017 ASHRAE Handbook—Fundamentals (SI) Sowell, E.F., and D.C. Chiles. 1984b. Zone descriptions and response characterization for CLF/CLTD calculations. ASHRAE Transactions 91(2A): 179-200. Spitler, J.D. 1996. Annotated guide to load calculation models and algorithms. ASHRAE. Spitler, J.D., F.C. McQuiston, and K.L. Lindsey. 1993. The CLTD/SCL/CLF cooling load calculation method. ASHRAE Transactions 99(1): 183-192. Spitler, J.D., and F.C. McQuiston. 1993. Development of a revised cooling and heating calculation manual. ASHRAE Transactions 99(1):175-182. Stephenson, D.G. 1962. Method of determining non-steady-state heat flow through walls and roofs at buildings. Journal of the Institution of Heating and Ventilating Engineers 30:5. Stephenson, D.G., and G.P. Mitalas. 1967. Cooling load calculation by thermal response factor method. ASHRAE Transactions 73(2):III.1.1. Stephenson, D.G., and G.P. Mitalas. 1971. Calculation of heat transfer functions for multi-layer slabs. ASHRAE Transactions 77(2):117-126. Sun, T.-Y. 1968. Computer evaluation of the shadow area on a window cast by the adjacent building. ASHRAE Journal (September). Todorovic, B. 1982. Cooling load from solar radiation through partially shaded windows, taking heat storage effect into account. ASHRAE Transactions 88(2):924-937. Todorovic, B. 1984. Distribution of solar energy following its transmittal through window panes. ASHRAE Transactions 90(1B):806-815. Todorovic, B. 1987. The effect of the changing shade line on the cooling load calculations. In ASHRAE videotape, Practical applications for cooling load calculations. Todorovic, B. 1989. Heat storage in building structure and its effect on cooling load; Heat and mass transfer in building materials and structure. Hemisphere Publishing, New York. Todorovic, B., and D. Curcija. 1984. Calculative procedure for estimating cooling loads influenced by window shading, using negative cooling load method. ASHRAE Transactions 2:662. Todorovic, B., L. Marjanovic, and D. Kovacevic. 1993. Comparison of different calculation procedures for cooling load from solar radiation through a window. ASHRAE Transactions 99(2):559-564. Wilkins, C.K. 1998. Electronic equipment heat gains in buildings. ASHRAE Transactions 104(1B):1784-1789. York, D.A., and C.C. Cappiello. 1981. DOE-2 engineers manual (Version 2.1A). Lawrence Berkeley Laboratory and Los Alamos National Laboratory.
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11. BUILDING EXAMPLE DRAWINGS
Fig. 17 First Floor Shell and Core Plan
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Fig. 18 Second Floor Shell and Core Plan
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Fig. 19 East/West Elevations, Elevation Details, and Perimeter Section
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Fig. 20 First Floor Tenant Plan
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Fig. 21 Second Floor Tenant Plan
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Fig. 22 3D View
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ENERGY ESTIMATING AND MODELING METHODS General Considerations ........................................................... 19.1 Degree-Day and Bin Methods ................................................. 19.6 Thermal Loads Modeling......................................................... 19.8 HVAC Component Modeling ................................................. 19.15
Low-Energy System Modeling................................................ Data-Driven Modeling ........................................................... Model Calibration .................................................................. Validation and Testing ...........................................................
NERGY requirements of HVAC systems directly affect a building’s operating cost and indirectly affect the environment. This chapter discusses methods for estimating energy use for two purposes: modeling for building and HVAC system design and associated design optimization (forward modeling), and modeling energy use of existing buildings for establishing baselines, calculating retrofit savings, and implementing model predictive control (data-driven modeling) (Armstrong et al. 2006a; Gayeski et al. 2012; Krarti 2010).
energy use of such a building can then be predicted or simulated by the forward-simulation model. The primary benefits of this method are that it is based on sound engineering principles and has gained widespread acceptance by the design and professional community. Major simulation codes, such as DOE-2, EnergyPlus, ESP-r, and TRNSYS are based on forward-simulation models. Although procedures for estimating energy requirements vary considerably in their degree of complexity, they all have three common elements: calculation of (1) space load, (2) secondary equipment load and energy requirements, and (3) primary equipment energy requirements. Here, secondary refers to equipment that distributes the heating, cooling, or ventilating medium to conditioned spaces, whereas primary refers to central plant equipment that converts fuel or electric energy to heating or cooling effect. The space load is the amount of energy that must be added to or extracted from a space to maintain thermal comfort. The simplest procedures assume that the energy required to maintain comfort is only a function of the outdoor dry-bulb temperature. More detailed methods consider humidity, solar effects, internal gains, heat and moisture storage in walls and interiors, and effects of wind on both building envelope heat transfer and infiltration. The section on Thermal Loads Modeling addresses some of these factors. ASHRAE Standard 183 and Chapters 17 and 18 discuss load calculation in detail. Although energy calculations are similar to the heating and cooling design load calculations used to size equipment, they are not the same. Energy calculations are based on average use and typical weather conditions rather than on maximum use and worst-case weather. Currently, most procedures are based on hourly profiles for climatic conditions and operational characteristics for a number of typical days of the year or on 8760 hours of operation per year. The space load is converted to a load on the secondary equipment. This can be a simple estimate of duct or piping losses or gains, or a complex hour-by-hour simulation of an air system, such as variableair-volume with outdoor-air cooling. This step must include calculation of all forms of energy required by the secondary system (e.g., electrical energy to operate fans and/or pumps, energy in heated or chilled water). The secondary equipment load is converted to the fuel and energy required by the primary equipment and the peak demand on the utility system. It considers equipment efficiencies and part-load characteristics. It is often necessary to keep track of the different forms of energy, such as electrical, natural gas, and/or oil. In some cases, where calculations are required to ensure compliance with codes or standards, these energies must be converted to source energy or resource consumed, as opposed to energy delivered to the building boundary. Previously, the steps were performed independently: each step was completed for the entire year and hourly results were passed to the next step. Current software usually performs all steps at each time interval, allowing effects such as insufficient plant capacity to be reflected in room conditions. Often, energy calculations lead to an economic analysis to establish the cost effectiveness of efficiency measures (as in ASHRAE
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E
1.
GENERAL CONSIDERATIONS 1.1
MODELS AND APPROACHES
A mathematical model is a description of the behavior of a system. It is made up of three components (Beck and Arnold 1977): • Input variables (statisticians call these regressor variables, whereas physicists call them forcing variables), which act on the system. There are two types: controllable by the experimenter (e.g., internal gains, thermostat settings), and uncontrollable (e.g., climate). • System structure and parameters/properties, which provide the necessary physical description of the system (e.g., thermal mass or mechanical properties of the elements). • Output (response, or dependent) variables, which describe the reaction of the system to the input variables. Energy use is often a response variable. The science of mathematical modeling as applied to physical systems involves determining the third component of a system when the other two components are given or specified. There are two broad but distinct approaches to modeling: forward (classical) and data driven (inverse). The choice of approach is dictated by the objective or purpose of the investigation (Rabl 1988).
Forward (Classical) Approach The objective is to predict the output variables of a specified model with known structure and known parameters when subject to specified input variables. To ensure accuracy, models have tended to become increasingly detailed. This approach presumes knowledge not only of the various natural phenomena affecting system behavior but also of the magnitude of various interactions (e.g., effective thermal mass, heat and mass transfer coefficients). The main advantage of this approach is that the system need not be physically built to predict its behavior. The forward-modeling approach is ideal in the preliminary design when design details are limited. Forward modeling of building energy use begins with a physical description of the building system or component of interest. For example, building geometry, geographical location, physical characteristics (e.g., wall material and thickness), type of equipment and operating schedules, type of HVAC system, building operating schedules, plant equipment, etc., are specified. The peak and average The preparation of this chapter is assigned to TC 4.7, Energy Calculations.
19.1 Copyright © 2017, ASHRAE
19.24 19.27 19.34 19.37
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Standard 90.1). Thus, thorough energy analysis provides intermediate data, such as time of energy use and maximum demand, so utility charges can be accurately estimated. Although not part of the energy calculations, capital equipment costs should also be estimated to assess the life-cycle costs of alternative efficiency measures.
1.2
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Data-Driven (Inverse) Approach In this approach, input and output variables are known and measured, and the objective is to determine a mathematical description of the system and to estimate system parameters. In contrast to the forward approach, the data-driven approach is relevant only when the system has already been built and actual performance data are available for model development, calibration (see the section on Model Calibration), and/or identification. Two types of performance data can be used: nonintrusive and intrusive. Intrusive data are gathered under conditions of predetermined or planned experiments on the system to elicit system response under a wider range of system performance than would occur under normal system operation to allow more accurate model identification. When constraints on system operation do not allow such tests to be performed, the model must be identified from nonintrusive data obtained under normal operation. The data-driven model has to meet requirements very different from the forward model. The data-driven model can only contain a relatively small number of parameters because of the limited and often repetitive information contained in the performance data. (For example, building operation from one day to the next is fairly repetitive.) It is thus a much simpler model that contains fewer terms representative of aggregated or macroscopic parameters (e.g., overall building heat loss coefficient and time constants). Because model parameters are deduced from actual building performance, it is much more likely to accurately capture as-built system performance, thus allowing more accurate prediction of future system behavior under certain specific circumstances. Performance data collection and model formulation need to be appropriately tailored for the specific circumstance, which often requires a higher level of user skill and expertise. In general, data-driven models are less flexible than forward models in evaluating energy implications of different design and operational alternatives, and so are not substitutes in this regard. To better understand the uses of data-driven models, consider some of the questions that a building professional may ask about an existing building with known energy consumption (Rabl 1988): • How does energy consumption compare with design predictions (and are any discrepancies caused by anomalous weather, unintended building operation, improper operation, as-built deficiency, etc.)? • How would consumption change if thermostat settings, ventilation rates, or indoor lighting levels were changed? • How much energy could be saved by retrofits to the building shell, changes to air handler operation from constant volume (CV) to variable air volume (VAV), or changes in the various control settings? • If retrofits are implemented, can one verify that the savings are due to the retrofit and not to other causes (e.g., weather)? • How can one detect faults in HVAC equipment and optimize control and operation? All these questions are better addressed by the data-driven approach. The forward approach could also be used, for example, by going back to the blueprints of the building and of the HVAC system, and repeating the analysis performed at the design stage using actual building schedules and operating modes, but this is tedious and labor intensive, and materials and equipment often perform differently in reality than as specified. Tuning the forward-simulation model is often problematic, although it is an option (see the section on Model Calibration).
OVERALL MODELING STRATEGIES
In developing a simulation model for building energy prediction, two basic issues must be considered: (1) modeling components or subsystems and (2) overall modeling strategy. Modeling components, discussed in the sections on Thermal Loads Modeling and HVAC Component Modeling, results in sets of equations describing the individual components. The overall modeling strategy refers to the sequence and procedures used to solve these equations. The accuracy of results and the computer resources required to achieve these results depend on the modeling strategy. Early building energy programs, including some still in common use, execute load models for every space for every hour of the simulation period. (Most models of this type use 1 h as the time step, which excludes any information on phenomena occurring in a shorter time span.) Next, the program runs models for every secondary system, one at a time, for every hour of the simulation. Finally, the plant simulation model is executed again for the entire period. This procedure is shown in Figure 1. Solid lines represent data passed from one model to the next; dashed lines represent information, usually provided by the user, about one model passed to the preceding model. For example, the system information may consist of a piecewise-linear function of zone temperature that gives system capacity. Because of this loads-systems-plants sequence, certain phenomena cannot be modeled precisely. For example, if the heat balance method for computing loads is used, and some component in the system simulation model cannot meet the load, the program can only report the current load. In actuality, the space temperature should readjust until the load matches equipment capacity, but this cannot be modeled, because loads have been precalculated and fixed. If the weighting-factor method is used for loads, this problem is partially overcome, because loads are continually readjusted during the system simulation. However, the weighting-factor technique is based on linear mathematics, and wide departures of room temperatures from those used during execution of the load program can introduce errors. A similar problem arises in plant simulation. For example, in an actual building, as load on the central plant varies, the supply chilled-water temperature also varies. This variation in turn affects the capacity of secondary system equipment. In an actual building, when the central plant becomes overloaded, space temperatures should rise to reduce load. However, with this separate load-systemplant approach, this condition cannot occur; thus, only the overload condition can be reported. These are some of the penalties associated with decoupling the load, system, and plant models. More recent building energy programs use an alternative strategy, in which all calculations are performed at each time step. Here, the load, system, and plant equations are solved simultaneously at each time interval. With this strategy, unmet loads and imbalances cannot occur; conditions at the plant are immediately reflected to the secondary system and then to the load model, forcing them to readjust to the instantaneous conditions throughout the building. The results of this modeling strategy are superior, although the magnitude and importance of the improvement are case specific.
Fig. 1 Overall Modeling Strategy
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The principal disadvantage of this approach, and the reason that it was not widely used in the past, is that it demands more computing resources. However, most current desktop computers can now run programs using the alternative approach in a reasonable amount of time. Programs that, to one degree or another, implement simultaneous solution of the loads, system, and plant models have been developed by ESRU (2016), Klein et al. (1994), Park et al. (1985), Taylor et al. (1990, 1991), and U.S. Department of Energy (19962016). Some of these programs can simulate the loads, systems, and plants using subhourly time steps. An economic model, as shown in Figure 1, calculates energy costs (and sometimes capital costs) based on the estimated required input energy. Thus, the simulation model calculates energy use and cost for any given input weather and internal loads. By applying this model (i.e., determining output for given inputs) at each hour (or other suitable interval), the hour-by-hour energy consumption and cost can be determined. Maintaining running sums of these quantities yields monthly or annual energy usage and costs. Because detailed models are computationally intensive, several simplified methods have been developed, including the degree-day, bin, and correlation methods. See the section on Degree-Day and Bin Methods for more information.
1.3
SIMULATING SECONDARY AND PRIMARY SYSTEMS
Traditionally, most energy analysis programs include a set of preprogrammed models that represent various systems (e.g., variable air volume, terminal reheat, multizone, variable refrigerant flow). In this scheme, the equations for each system are arranged so they can be solved sequentially. If this is not possible, the smallest number of equations that must be solved simultaneously is solved using an appropriate technique. Furthermore, individual equations may vary hourly in the simulation, depending on controls and operating conditions. For example, a dry coil uses different equations than a wet coil. The primary disadvantage of this scheme is that it is relatively inflexible: to modify a system, the program source code may have to be modified and recompiled. An alternative strategy views the system as a series of components (e.g., fan, coil, pump, duct, pipe, damper, thermostat) that may be organized in a component library (Park et al. 1985; TRNSYS 2012). Users specify connections between the components, and the program then resolves the specification of components and connections into a set of simultaneous equations. A refinement of component-based modeling is known as equation-based modeling (Buhl et al. 1993; Sowell and Moshier 1995). These models do not follow predetermined rules for a solution, and the user can specify which variables are inputs and which are outputs. Current research in this area centers on use of the Modelica object-oriented modeling language (Wetter et al. 2011).
1.4
HISTORY OF SIMULATION METHOD DEVELOPMENT
Significant historical events, such as World Wars I and II, the development of analog and digital computers and programming languages (1950s), and repeated oil crises (1967, 1973, and 1979) spurred the development of analysis methods and computer simulation programs now used to simulate building energy use. Before the 1950s, most of the fundamentals (i.e., gas laws, heat transfer properties, and thermodynamics) of today’s HVAC systems had been studied and published, which contributed significantly to the rapid development of today’s building energy simulation technology. Also, during this same period, researchers and engineers developed the essential methods for calculating and analyzing the dynamic heat gain through multilayer walls, overall dynamic building cooling and
19.3 heating loads, solar heating performance, and internal reflected illuminance (Oh 2013; Oh and Haberl 2016a, 2016b, 2016c). In the 1920s, one of the most important methods for calculating the dynamic heat gain through walls and roofs for whole-building energy simulation [i.e., the response factor method (RFM)] was developed and published by André Nessi and Léon Nisolle at École Centrale Paris (French University) in France (Nessi and Nisolle 1925). The RFM concept is widely used in most of today’s cooling and heating loads calculations, such as with the weighting factor method or the transfer function method (Mitalas and Stephenson 1967; Stephenson and Mitalas 1967, 1971) used in DOE-2.1E (Winkelmann et al. 1993), DOE-2.2/eQUEST (LBNL 1998), TRACE (Trane 2013), and HAP (Carrier 2013); the CLTD/CLF method used in TRACE; the heat balance method (Pedersen et al. 1997) used in BLAST (Hittle 1977) and EnergyPlus (Crawley et al. 2001); and the radiant time series method (Spitler et al. 1997) used in TRACE. In the 1940s, the concept of the resistance-capacitance (RC) network analysis method (i.e., the thermal network method) used in simulation programs for buildings was first introduced by Victor Paschkis, a research engineer at Columbia University (Paschkis 1942; Paschkis and Baker 1942). The thermal network concept is widely used today in whole-building simulation programs such as DOE-2.1E, DOE-2.2/eQUEST, and EnergyPlus and in detailed solar simulation programs such as TRNSYS and SUNREL (Deru et al. 2002). In the 1950s, the most important method (i.e., the utilizability method) for calculating the performance of solar heating systems for design programs was developed (Whillier 1953). The utilizability method is used today in simplified solar design programs such as RETScreen (NRC 2004), F-Chart (Klein and Beckman 2001a), and PV F-Chart (Klein and Beckman 2001b). Also during the 1950s, analog computers were first used to study the behavior of dynamic heat gain/loss and the impact of dynamic heat gain/loss on HVAC systems (Buchberg 1955, 1958; Nottage and Parmelee 1954; Willcox et al. 1954). During the 1960s, digital computers and the analysis methods suitable for the digital computers were developed and became widely used. Digital computers were quickly substituted for analog computers during this period because the digital computer was more convenient to program, more flexible, and more agile at describing the governing equations and driving functions than the methods used in analog computers. Finally, the scientific applications of the digital computer were considerably improved by the FORTRAN programming language, a high-level, scientific programming language that was first commercially released in 1957 by IBM. In the 1960s, FORTRAN allowed computer codes written on one computer to be run on another computer by a different analyst, which accelerated the wide-spread availability of building energy simulation programs. The combination of computer advances with the analysis methods developed by numerous researchers facilitated the development of today’s wholebuilding energy, solar energy, and daylighting simulation programs.
1.5
USING ENERGY MODELS
Typical Applications Energy models are useful for a range of applications, and the typical application of forward-modeling tools falls into three categories: • Comparison. A common application is to use energy models to compare the estimated performance of design alternatives for new buildings or to compare retrofit alternatives for existing buildings. In such design studies, the goal is to estimate the relative performance of two or more options. Examples include comparing alternatives for building form and orientation, comparing different glazing selections, or comparing chiller alternatives. Energy models are especially well suited to such comparison studies,
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19.4 because the relative performance is likely to be estimated with reasonable accuracy even if some inputs, such as actual operating hours or plug loads, are not accurately known at the time of the study. • Compliance. Using models for energy code compliance (e.g., ASHRAE Standard 90.1) and for building rating system credit (e.g., U.S. Green Building Council’s LEED® rating system) are very common applications. In most cases, these studies involve comparing two energy models: one representing a baseline building that meets the minimum code requirements, and the other representing the proposed building design, with the result an estimate of relative performance. Some utility and other incentive programs also use the same method for estimating energy savings for new and retrofit construction projects. Note that the models will not necessarily be good predictors of actual energy use unless the inputs have been carefully specified to represent the expected operation of the actual building. • Prediction. An application of growing importance is using energy models to predict energy consumption. Some industry efforts, such as Architecture 2030 (AIA 2014), encourage building designers to set targets for energy consumption. In those cases, the goal is to design a building that meets an absolute energy target, often specified in terms of kWh/m2 ·yr. In addition, some owners seek zero-energy buildings that offset consumption with renewable energy. During the design of zero-energy buildings, accurate prediction of energy consumption is important for proper sizing of renewable energy systems. Energy models used for prediction require very careful attention to providing realistic inputs to represent the expected building operation.
Choosing Measures for Evaluation Forward-modeling programs are useful for evaluating many types of design elements, whereas they may be unnecessary or inappropriate for evaluating other elements. They are most useful for evaluating elements that affect heating and cooling loads. Many programs also provide the ability to compare the energy performance of different HVAC designs. The following are examples of design elements that are typically good candidates for evaluation with forward-modeling programs: • Building form and orientation. Varying building dimensions. • Opaque envelope. Varying levels of insulation and thermal mass. • Fenestration. Alternative glazing and framing types, and both fixed and operable shading devices. • Daylighting controls. Requires a program with an adequately accurate daylight illuminance calculation. • Cooling equipment efficiency. Full- and partial-load performance. • Heating equipment efficiency. Full- and partial-load performance, condensing versus noncondensing equipment. • Economizer cooling. Air- and water-side economizers. • Variable-flow controls. Air and hydronic distribution system control. • Basic HVAC controls. Space temperature set points, HVAC supply temperature reset controls, start/stop scheduling, load management scheduling. • Combined heat and power systems. Some programs can apply waste heat from cogeneration systems to building thermal demands. Some design elements do not necessarily require an energy model for evaluation, and a reasonably accurate savings estimate may be possible using other methods. Although it may still be useful to represent them in an energy model along with other measures, it may not be worth the effort for an analysis looking exclusively at these measures:
2017 ASHRAE Handbook—Fundamentals (SI) • Plug loads. The majority of the impact is from direct savings that may be calculated simply with an estimate of connected power and operating hours. An energy model may be appropriate if the effect on heating and cooling loads is expected to be significant. • Interior lighting power. Similar issues to plug loads. • Exterior lighting power and control. Has no effect on heating and cooling loads. • Control of constant-speed/constant-volume equipment. The savings from an adjustment in flow may be simple to calculate without using an energy model. • Photovoltaic and other on-site generation systems. Although some programs offer the capability to model photovoltaic systems or other on-site electricity production, separate calculations may be equally accurate and require less effort. Limitations to the capabilities of most forward-modeling programs are important to understand. Some programs may not be appropriate candidates for performing an evaluation, depending on the feature to be analyzed: • Daylight-redirecting devices. Design elements such as light shelves or light-redirecting glazing are sometimes used to increase the daylighted area. It is important to understand the capabilities of the tool used for analysis. • Dynamic glazing. Dynamic glazing may be used to control glare and solar heat gain. It is important to understand the capabilities of the modeling tool for this design element. • Natural ventilation. Not all programs can model natural ventilation or the integration of mechanical and natural ventilation. See the section on Natural Ventilation for more information. • Specific HVAC system types. The selected program might not have the ability to explicitly model a desired HVAC system configuration. • Specific HVAC system control strategies. Programs vary in capability to represent varying control strategies.
When to Use Energy Models Energy modeling has appropriate applications throughout the stages of design, construction, and operation of a building. At the earliest conceptual design phase, comparative studies help designers compare options for building form, system types, and other decisions that will be hard to change at later stages. During subsequent phases, energy modeling studies can be used to refine design decisions such as component specifications. At the conclusion of design, the models can be used to document energy code or ratingsystem compliance. Although use of energy models during construction is less common, their use during commissioning to compare expected performance to observed performance is growing. Similarly, building performance can be verified by comparing actual energy consumption to modeled performance, and any observed differences can help the building operator identify potential areas for improvement. The American Institute of Architects has published a guide to integrating energy modeling in the design process (AIA 2012). Proposed ASHRAE Standard 209P specifies mandatory and optional modeling tasks for new construction projects.
Energy Modelers Creation of energy models has typically been done by specialists, and specialists continue today to perform a significant amount of energy modeling work. However, increasing user-friendliness of modeling software and the integration of building information models and energy modeling tools have made energy modeling more accessible to nonspecialists. Anyone performing energy modeling should understand the way their program of choice works, understand assumptions being used by the program, and know how to interpret the results obtained from the program.
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19.5
ASHRAE offers a Building Energy Modeling Professional (BEMP) qualification exam for modeler certification (www.ashrae .org/education--certification/certification/bemp-building-energy -modeling-professional-certification).
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1.6
UNCERTAINTY IN MODELING
Uncertainty is inherent in modeling. When modeling a complex system, there are often uncertainties in identifying system inputs. In addition, there are uncertainties and variability in known model input parameters, and uncertainties in the model itself. It is important for modelers to characterize the uncertainties to understand how to interpret the uncertainty in the model prediction (Macdonald and Strachan 2001; de Wit and Augenbroe 2002). In reporting a model prediction, providing a value without providing its confidence interval implies 0% uncertainty, which is not possible because of many inherent uncertainties in the modeling process. Therefore, it is advisable to provide the predicted value as an interval rather than a precise value. Uncertainties can be broken into two broad categories: aleatoric and epistemic (Kiureghaian and Ditlevsen 2009). Aleatoric uncertainties are those related to inherent randomness and variability of the system and the inputs to the system that cannot be reduced by better knowledge or system observation. They are typically characterized by probabilities. In other words, aleatoric uncertainties are “unknown unknowns.” Examples of these in energy modeling are weather-related model inputs or occupant-related loads. Epistemic uncertainties are those related to uncertainty of the system model itself or parameters to the model that are fixed/static but not known precisely and that could be reduced with better knowledge of the system or the inputs. Epistemic uncertainty is also referred to as state-of-knowledge uncertainty, subjective uncertainty, or reducible uncertainty. In other words, epistemic uncertainties are “known unknowns.” Examples include building geometry, thermal properties of enclosure materials, or efficiencies of equipment (Hopfe and Hensen 2011; Sun et al. 2014). Knowing the type of uncertainty associated with a parameter is important in understanding how to propagate that uncertainty through the model. Uncertainties can be propagated through a model using a variety of methods, including Monte Carlo, importance sampling, local expansion, most-probable-point methods, and numeric integration (Lee and Chen 2008). When using “black box” simulation engines, Monte Carlo methods and importance sampling are the most applicable. Once the prediction uncertainty is quantified, there are various uses for the information. The output can be analyzed to obtain most probable values and/or variances. Output probability distributions can be analyzed for understanding the risk of underperformance. Uncertainty statistics can also be used for verifying that a model meets standards or guidelines for baseline predictions in measurement and verification, along with determining the uncertainty in cost savings calculations. ASHRAE Guideline 14 provides instruction on maximum levels of uncertainty in baseline models, along with simplified methods for assessing the quantifiable uncertainty in savings computations.
1.7
CHOOSING AN ANALYSIS METHOD
The most important step in selecting an energy analysis method is matching method capabilities with project requirements. The method must be capable of evaluating all design options with sufficient accuracy to make correct choices. The following factors apply generally (Sonderegger 1985): • Accuracy. The method should be sufficiently accurate to allow correct choices. Because of the many parameters involved in energy estimation, absolutely accurate energy prediction is not
• •
•
• •
possible (Waltz 1992). ANSI/ASHRAE Standard 140 was developed to identify and diagnose differences in predictions that may be caused by algorithmic differences, modeling limitations, coding errors, or input errors. More information on model validation and testing can be found in the Validation and Testing section of this chapter and in ANSI/ASHRAE Standard 140. Sensitivity. The method should be sensitive to the design options being considered. The difference in energy use between two choices should be adequately reflected. Versatility. The method should allow analysis of all options under consideration. When different methods must be used to consider different options, an accurate estimate of the differential energy use cannot be made. Speed and cost. The total time (gathering data, preparing input, calculations, and analysis of output) to make an analysis should be appropriate to the potential benefits gained. With greater speed, more options can be considered in a given time. The cost of analysis is largely determined by the total time of analysis. Reproducibility. The method should not allow so many vaguely defined choices that different analysts would get completely different results (Corson 1992). Ease of use. This affects both the economics of analysis (speed) and the reproducibility of results.
Selecting Energy Analysis Computer Programs Selecting a building energy analysis program depends on its application, number of times it will be used, experience of the user, and hardware available to run it. The first criterion is the ability of the program to deal with the application. For example, if the effect of a shading device is to be analyzed on a building that is also shaded by other buildings part of the time, the ability to analyze detached shading is an absolute requirement, regardless of any other factors. The cost of the computer facilities and the software itself are typically a small part of running a building energy analysis; the major costs are of learning to use the program and of using it. Major issues that influence the cost of learning a program include (1) complexity of input procedures, (2) quality of the user’s manual, and (3) of a good support system to answer questions. As the user becomes more experienced, the cost of learning becomes less important, but the need to obtain and enter a complex set of input data continues to consume the time of even an experienced user until data are readily available in electronic form compatible with simulation programs. Complexity of input is largely influenced by the availability of default values for the input variables. Default values can be used as a simple set of input data when detail is not needed or when building design is very conventional, but additional detail can be supplied when needed. Secondary defaults, which can be supplied by the user, are also useful in the same way. Some programs allow the user to specify a level of detail. Then the program requests only the information appropriate to that level of detail, using default values for all others. However, whenever default values are used, even internally in the software, the user should take care to understand the default values and whether they apply to the situation being analyzed. Quality of output is another factor to consider. Reports should be easy to read and uncluttered. Titles and headings should be unambiguous. Units should be stated explicitly. The user’s manual should explain the meanings of data presented. Graphic output can be very helpful. In most cases, simple summaries of overall results are the most useful, but very detailed output is needed for certain studies and also for debugging program input during the early stages of analysis. Before a final decision is made, manuals for the most suitable programs should be obtained and reviewed, and, if possible, demonstration versions of the programs should be obtained and run, and
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19.6
2017 ASHRAE Handbook—Fundamentals (SI)
support from the software supplier should be tested. The availability of training should be considered when choosing a more complex program. Availability of weather data and a weather data processing subroutine or program are major features of a program. Some programs include subroutine or supplementary programs that allow the user to create a weather file for any site for which weather data are available. Programs that do not have this capability must have weather files for various sites created by the program supplier. In that case, the available weather data and the terms on which the supplier will create new weather data files must be checked. More information on weather data can be found in Chapter 14. Auxiliary capabilities, such as economic analysis and design calculations, are a final concern in selecting a program. An economic analysis may include only the ability to calculate annual energy bills from utility rates, or it might extend to calculations or even to life-cycle cost optimization. An integrated program may save time because some input data have been entered already for other purposes. Results of computer calculations should be accepted with caution, because software vendors do not accept responsibility for the correctness of calculation methods and have no control over program use. Manual calculation should be done to develop a good understanding of underlying physical processes and building behavior. In addition, the user should (1) review the computer program documentation to determine what calculation procedures are used, (2) compare results with manual calculations and measured data, and (3) conduct sample tests to confirm that the program delivers acceptable results. The most accurate methods for calculating building energy consumption are the costliest because of their intense computational requirements and the expertise needed by the designer or analyst. Simulation programs that assemble component models into system models and then exercise those models with weather and occupancy data are preferred by experts for determining energy use in buildings. Often, energy consumption at a system or whole-building level must be estimated quickly to study trends, compare systems, or study building effects such as envelope characteristics. For these purposes, simpler methods, such as degree-day and bin, may be used. The International Building Performance Simulation Association (IBPSA-USA) maintains a website (www.buildingenergysoftware tools.com) (IBPSA 2015) with information about hundreds of available software tools. Crawley et al. (2005) compare the capabilities of many building simulation tools.
2.
DEGREE-DAY AND BIN METHODS 2.1
DEGREE-DAY METHOD
As described in Huang et al. (2001), the degree-day method was developed in the 1930s to estimate the heating energy consumption of a building. A degree-day is a measure of how often and by how many degrees the average daily temperature (the average of the daily maximum and minimum) for a location is above (for cooling) or below (for heating) a base temperature. For example, a day where the average daily temperature is 12 degrees lower than the base temperature would represent 12 degree-days; an equivalent would be 2 days, each of which was 6 degrees below the base temperature. A year’s summation of the number of heating or cooling degree-days for a location is a convenient single number indicating that location’s climate severity. The base temperature is that at which the building's internal heat gains counterbalance the heat losses to the outdoors, so that the building requires neither heating nor cooling. Below that “balance point” temperature, the building requires heating, whereas
above that temperature the building requires cooling, in proportion to the difference from the base temperature. Base 18°C heating degree-days and, to a lesser extent, base 18°C cooling degree-days have become widely accepted as the most convenient, simple indicators of climate severity. In the United States, heating degree-days vary from fewer than 500 in Miami, FL, 1000 to 3000 in the south, and 3000 to 7000 in the north, to extremes of more than 8000 in Bismarck, ND and 10 000 in Anchorage, AK. Correspondingly, cooling degree-days vary from 0 in Anchorage, AK, to fewer than 100 in Seattle, WA, 500 to 1200 in the north, 1200 to 3000 in the south, and more than 3000 in Phoenix, AZ and Miami, FL. In the degree-day method, the building heat load (i.e., the amount of heating energy input or cooling energy extraction) is estimated as the number of degree-days times 24 (to convert to degree-hours), times the overall building heat loss coefficient W/K. The overall heat loss coefficient is the sum of the U-factor area for all external surfaces, such as walls, windows, doors, roof, and losses or gains from infiltration. The cooling load equation is similar but uses cooling degree-days instead of heating degree-days. Example 1. Estimate the heating energy requirements for an 170 m2 residence in Denver (3300 K-days) using the degree-day method. The volume of the residence is 400 m3, the overall heat loss coefficient is 200 W/K, 0.7 air changes per hour (ach). The residence is heated by a furnace with an AFUE of 0.78, with an average duct loss factor of 0.10. Solution: The heating load equation is
h Q h = HDD 65 24 -------- day U i Ai + Infiltration air changes/h Volume i Volumetric heat capacity for air
To derive the heating or cooling energy consumption, the heating or cooling loads must be divided by the efficiencies of the HVAC system. 3300 24 200 + 0.7 400 1 3600 1210 Qh = ----------------------------------------------------------------------------------------------------------- = 33 200 kWh 0.78 0.9
The degree-day method, as demonstrated in Example 1, is limited in that it does not consider the effects of solar heat gain or building thermal mass, nor can it account for variations in infiltration rates, thermostat settings (such as night setback), or occupant actions such as window venting on cool summer nights or during the spring and fall seasons. Efforts to improve the degree-day method, such as by using variable base temperatures, calculating degree-hours instead of days, etc., are described in the following section. The gains in accuracy from these modifications have been offset by increased difficulty in computation, nearly always requiring a computer program. Because the capability of personal computers has grown to handle programs using methods that capture the transient heat flows that dominate building thermal processes, degree-day methods have fallen increasingly out of favor, because they remain fundamentally steadystate calculations. Even with these limitations, the degree-day method can still be of use in providing a quick answer that can be used as a starting point or check for more detailed calculations. The degree-day method may be adequate for cases in which gathering the additional data on climate and building conditions may be unwarranted or impractical, such as for estimating heating energy use in lightconstruction residential buildings with low solar gain in cold
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Energy Estimating and Modeling Methods climates. The degree-day method can also be useful in analyzing results from more detailed, often hard-to-interpret calculations.
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Variable-Base Degree-Day Method Further described in Huang et al. (2001), the variable-base degree-day (VBDD) method accounts for balance point temperatures varying between buildings and even within a building, depending on the time of day. The 18°C base temperature is based on the assumption that heat gains from the sun and internal processes contribute on average 3 K of “free heat,” allowing a set point of 21°C with heating required only when the outdoor temperature drops below 18°C. Better-insulated, more airtight buildings can have lower balance point temperatures.The average balance point temperature for modern residential buildings is now 13 to 14°C. Commercial buildings’ low surface-to-volume ratios, high windowto-wall ratios, and high internal gains allow balance point temperatures of 10°C or lower. Balance point temperatures differ markedly between daytime and nighttime hours. Figure 2 shows the variation of the balance point temperature for a typical house (Nisson and Dutt 1985). In daytime, the house receives heat gain from the sun and from occupant activity (e.g., equipment, lights). The balance point temperature may be around 8 K below the indoor temperature set point. However, at night, with no solar heat gain and reduced human activity, it may be only 2 K lower than set point. The variation can be even greater for commercial buildings because of higher internal heat gains during the day and very low heat gains at night. The VBDD method accounts for these different building conditions by calculating (1) the balance point temperature of a building and (2) the heating and cooling degree-hours at that temperature. To divide the daytime and nighttime, degree-hours must be used instead of degree-days. Instead of using the average between the daily maximum and minimum temperatures as for degree-days, degree-hours are calculated from the temperature for each hour. The number of degree-hours divided by 24 can be slightly to significantly larger than the number of degree-days at the same base temperature. An example of why this can occur is a spring day where the average daily temperature may be below the base temperature, but several afternoon hours are above it. Such a day would have no cooling degree-days but a number of cooling degree-hours; these differences can be particularly significant for cooling. The calculations are repeated for daytime and nighttime conditions for each month of the year using degree-hour totals for each period with different base temperatures.
Fig. 2 Variation of Balance Point Temperature and Internal Gains for Typical House (Nisson and Dutt 1985)
19.7 For cooling, the VBDD method can be adjusted to identify the number of cooling degree hours that are actually met by the HVAC system. The cooling balance point changes dramatically depending on whether the building is being vented. If the windows are open, heat gains are flushed from the building and do not affect its cooling load. However, if the windows are closed, solar and internal gains make the balance temperature significantly lower than the thermostat set point for cooling. The VBDD method accounts for these different conditions by calculating the cooling degree-hours using the balance temperature with the windows closed, but not counting degree-hours when the temperatures are below the thermostat set point, when the windows are assumed to be open. Figure 3 shows the one-to-one relationship between cooling degree-hours and the temperature difference between the balance point and outdoor air temperature. Those degree-hours occurring in the ventilation section when the outdoor temperatures are below the thermostat set point are not added to the running total of cooling degree-hours. VBDD method calculations are, in most cases, too tedious for either hand calculations or implementation in a spreadsheet. In the early 1980s, several PC programs were written using the VBDD method, such as CIRA (Computerized Instrumented Residential Audit) (Sonderegger et al. 1982). Variable-base degree-day programs such as CIRA represent the apex of the VBDD method’s application; however, it remains a steady-state method that does not recognize a building's thermal history.
Sources of Degree-Day Data The most general and directly usable source for degree-day data is the ASHRAE Weather Data Viewer DVD (ASHRAE 2013). This application performs useful aggregations for more than 6000 locations. In addition to design conditions, the Weather DataViewer derives annual and monthly degree-days relative to any base temperature, and also provides bin data. Results are in spreadsheet form. Precalculated annual and monthly degree-days for several common bases are included in the climatic data that accompany Chapter 14, and additional online and printed collections of climatic data are listed in the section on Other Sources of Climatic Information in that chapter. Annual and monthly degree-days relative to an arbitrary base can be estimated using the Weather Data Viewer or the procedures documented in the section on Estimation of Degree-Days in Chapter 14. Other degree-day estimation methods include the Erbs et al. (1983) model, which needs as input only the average to for each month of the year.
Fig. 3 Uncounted Ventilation Degree-Hours versus Counted Cooling Degree-Hours
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19.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Sample Annual Bin Data Bin 39/ 41
Site Chicago, IL Dallas/Ft. Worth, TX Denver, CO Los Angeles, CA Miami, FL Nashville, TN Seattle, WA
Licensed for single user. © 2017 ASHRAE, Inc.
2.2
36/ 38
33/ 35
74 4 170 322 81 4 10 9 14 4 82
30/ 32
27/ 29
24/ 26
21/ 23
176 431 512 960 511 922 1100 1077 217 406 390 570 16 56 194 1016 648 2147 2581 1852 366 717 756 1291 10 88 139 330
18/ 20
15/ 17
12/ 14
9/ 11
6/ 8
3/ 5
660 750 726 1874 734 831 497
591 803 712 2280 390 693 898
780 870 902 2208 202 801 1653
510 581 809 843 100 670 1392
770 728 783 227 76 858 1844
686 418 750 23 14 639 1127
0/ 2
–3/ –1
–6/ –4
–9/ –12/ –15/ –18/ –21/ –7 –10 –13 –16 –19
1671 380 304 125 464 37 3 1467 446 216 106 2 793 141 715 40
89 26
66
49
11
4
85
52
44
8
29 1
BIN AND MODIFIED BIN METHODS
For many applications, the degree-day or variable-base degreeday methods should not be used, because the overall building heat loss coefficient Ktot , the efficiency h of the HVAC system, or the balance point temperature tbal may not be sufficiently constant. Heat pump efficiency, for example, varies strongly with outdoor temperature; efficiency of HVAC equipment may be affected indirectly by the outdoor temperature to when efficiency varies with load (common for boilers and chillers). Furthermore, in most commercial buildings, occupancy has a pronounced pattern that affects heat gain, indoor temperature, and ventilation rate. In such cases, steady-state calculation can yield good results for annual energy consumption if different temperature intervals and time periods are evaluated separately. This approach is known as the bin method, because consumption is calculated for several values of to and multiplied by the number of hours Nbin in the temperature interval (bin) centered on that temperature: K tot Qbin = Nbin ---------- [tbal – to]+ h
(1)
The superscript plus sign indicates that only positive values are counted; no heating is needed when to is above tbal. Equation (1) is evaluated for each bin, and the total consumption is the sum of the Qbin over all bins. The underlying assumption of the bin method is that, for a given temperature at the same general time of day (morning, afternoon, evening, etc.), the heating and cooling loads of a building should be roughly the same. Therefore, one can derive a building’s annual heating and cooling loads by calculating its loads for a set of “snapshots” defined by temperature “bins,” multiplying the calculated loads by the number of hours represented by each bin, and then totaling the sums to derive the building’s annual heating and cooling loads. Table 1 provides sample annual bin data for several U.S. sites. In the original formulation of the bin method, there was no accounting for effects such as occupancy, solar gain, or wind on the calculated loads. The modified bin method (Knebel 1983) extended the basic bin method to account for weekday/weekend and partial-day occupancy effects, to calculate net building loads (conduction, infiltration, internal loads, and solar loads) at four temperatures, rather than interpolate from design values, and to better describe secondary and primary equipment performance. Other versions of the bin method accounted for solar and wind effects by using more detailed binned data that gave the average wind speeds and solar gains by month, and the number of hours within bins disaggregated by month as well as time of day. Vadon et. al (1991) also presented an improvement for correlating the solar gains to the temperature bins. Example 2. Estimate the energy requirements for a residence with a design heat loss Qdes of 11 700 W at 30°C design temperature difference. The indoor design temperature is 21°C. Average internal heat gains are
Fig. 4 Heat Pump Capacity and Building Load estimated to be 1250 W. Assume a 10.5 kW heat pump with the characteristics given in Columns E and H of Table 2 and in Figure 4. Solution: The design heat loss is based on no internal heat generation. The heat pump system energy input is the net heat requirement of the space (i.e., envelope loss minus internal heat generation). The net heat loss per degree and the heating/cooling balance temperature may be computed: Ktot = Qdes/t = 11 700/30 = 390 W/K The balance temperature tbal is defined as the temperature at which the external heat loss equals the internal heat gain Qint, and therefore Qint = Ktot (to – tbal) and tbal = to – (Qint /Ktot) thus tbal = 21 – (1250/390) = 17.8°C Table 2 is then computed, resulting in 7618 kWh.
3. THERMAL LOADS MODELING 3.1
SPACE SENSIBLE LOAD CALCULATION METHODS
Calculating instantaneous space sensible load is a key step in any building energy simulation. The heat balance, comprehensive room transfer function (CRTF), and weighting-factor methods are used for these calculations. A fourth method, the thermalnetwork method, is similar in rigor to the heat balance and CRTF methods but not as widely used.
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Energy Estimating and Modeling Methods
19.9
Table 2 Calculation of Annual Heating Energy Consumption for Example 2 Climate A
B
House C
Temp. Temp. Weather Bin, Diff., tbal Data – tbin Bin, h °C 16 13 10 7 4 1 –2 –5 –8 –11
1.8 4.8 7.8 10.8 13.8 16.8 19.8 22.8 25.8 28.8
693 801 670 858 639 793 141 89 29 0
D
Heat Pump E
F
G
Supplemental I
J
K
Heat Adjusted Heat Pump Heat Integrated Cycling Rated Heat Pump Seasonal Heat Loss Pump Electric Operating Supplied Pump Electric Heating Capacity Rate, Capacity, Adjustment Capacity, Input, Time Heating, Consumption, kW kW Fractionc kWhd kWhe kWb Factora kW 0.70 1.87 3.04 4.21 5.38 6.55 7.72 8.89 10.06 11.23
12.80 12.01 11.22 9.80 8.49 7.98 7.47 6.95 6.48 5.69
0.764 0.789 0.818 0.857 0.908 0.955 1.000 1.000 1.000 1.000
9.78 9.48 9.18 8.40 7.71 7.62 7.47 6.95 6.48 —
aCycling
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H
Capacity Adjustment Factor = 1 Cd(1 x), where Cd = degradation coefficient (default = 0.25 unless part-load factor is known) and x = building heat loss per unit capacity at temperature bin. Cycling capacity = 1 at balance point and below. Cycling capacity adjustment factor should be 1.0 at all temperature bins if manufacturer includes cycling effects in heat pump capacity (Column E) and associated electrical input (Column H).
3.74 3.63 3.52 3.40 3.18 3.10 3.02 2.93 2.85 —
0.072 0.197 0.331 0.501 0.698 0.860 1.000 1.000 1.000 — Totals:
488 1496 2036 3611 3439 5196 1053 618 188 — 18 125
187 573 781 1462 1418 2114 426 261 83 — 7305
bColumn
G = Column E Column F time fraction equals smaller of 1 or Column D/Column G dColumn J = Column I Column G Column C eColumn K = Column I Column H Column C cOperating
L
M
N
Total SuppleElectric mental Space Heating Energy Load, Required, ConsumpkWhf kWhg tion, kWhh 485 1497 2037 3612 3438 5195 1089 791 292 — 18 436 fColumn
gColumn
— — — — — — 36 173 104 — 313
187 573 781 1462 1418 2114 462 434 187 — 7618
L = Column C Column D M = Column L – Column J
hColumn N = Column K + Column M
The instantaneous space sensible load is the rate of heat flow into the space air mass. This quantity, sometimes called the cooling load, differs from heat gain, which usually contains a radiative component that passes through the air and is absorbed by other bounding surfaces. Instantaneous space sensible load is entirely convective; even loads from internal equipment, lights, and occupants enter the air by convection from the surface of such objects or by convection from room surfaces that have absorbed the radiant component of energy emitted from these sources. However, some adjustment must be made when radiant cooling and heating systems are evaluated because some of the space load is offset directly by radiant transfer without convective transfer to the air mass. For equilibrium, the instantaneous space sensible load must match the heat removal rate of the conditioning equipment. Any imbalance in these rates changes the energy stored in the air mass. Customarily, however, the thermal mass (heat capacity) of the air itself is ignored in analysis, so the air is always assumed to be in thermal equilibrium. Under these assumptions, the instantaneous space sensible load and rate of heat removal are equal in magnitude and opposite in sign. The weighting-factor, CRTF, and heat balance methods use conduction transfer functions (or their equivalents) to calculate transmission heat gain or loss. The main difference is in the methods used to calculate the subsequent internal heat transfers to the room. Experience has shown that all the methods produce similar results, provided model coefficients are determined for the specific building under analysis and room temperature variations are moderate.
rigorous. However, the heat balance method requires more calculations at each point in the simulation process, using more computer time. This method has been validated in several studies (Chantrasrisalai et al. 2003a, 2003b; Eldridge et al. 2003; Iu et al. 2003). The heat balance method allows the net instantaneous sensible heating and/or cooling load to be calculated on the space air mass. Generally, a heat balance equation is written for each enclosing surface, plus one equation for room air. Although not necessary, linearization is commonly used to simplify the radiative transfer formulation. This set of equations can then be solved for the unknown surface and air temperatures. Once these temperatures are known, they can be used to calculate the convective heat flow to or from the space air mass. The heat balance method is developed in Chapter 18 for use in design cooling load calculations. The procedure described in Chapter 18 is aimed at obtaining the design cooling load for a fixed zone air temperature. For building energy analysis purposes, it is preferable to know the actual heat extraction rate. This may be determined by recasting Equation (27) of Chapter 18 so that the system heat transfer is determined simultaneously with the zone air temperature. The system heat transfer is the rate at which heat is transferred to the space by the system. Although this can be done by simultaneously modeling the zone and the system (Taylor et al. 1990, 1991), it is often convenient to make a simple, linearized representation of the system known as a control profile. This usually takes the form
Heat Balance Method
where
The heat balance method for calculating net space sensible loads is described in Chapter 18 and in the ASHRAE Toolkit for Building Load Calculations (Barnaby et al. 2005, 2009; Iu and Fisher 2004; Pedersen et al. 2001, 2003). Its development relies on the first law of thermodynamics (conservation of energy) at each surface. Because the heat balance method involves fewer assumptions than the weighting-factor method, it is more flexible and physically
q sys = a + bt a j
j
(2)
q sys j = system heat transfer at time step j, W a, b = coefficients that apply over a certain range of zone air temperatures t a j = zone air temperature at time step j, °C
System heat transfer q sysj may be considered positive when heating is provided to the space and negative when cooling is provided.
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19.10
2017 ASHRAE Handbook—Fundamentals (SI)
It is equal in magnitude but opposite in sign to the zone cooling load, as defined in Chapter 18, when zone air temperature is fixed. Substituting Equation (2) into Equation (27) of Chapter 18 and solving for zone air temperature, a + A i h ci t si + cV infil t o + cV vent t v + q c ,int N
i, j
i=1
j
j
j
j
– b + A i h ci + cV infil + cV vent
j
t a = --------------------------------------------------------------------------------------------------------------------------N
j
j
(3)
j
i=1
where N = number of zone surfaces Ai = area of ith surface, m2 hci = convection coefficient for ith surface, W/(m2 ·K) t si = surface temperature for ith surface at time step j, °C i, j
= density, kg/m3 c = specific heat of air, J/(kg ·K) V = volumetric flow rate of air, m3/s t o = outdoor air temperature at time step j, °C j
t v = ventilation air temperature at time step j, °C
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j
q c int = sum of convective portions of all internal heat gains at time j step j, W
The zone air heat balance equation [Equation (3)] must be solved simultaneously with the interior and exterior surface heat balance equations [Equations (26) and (25) in Chapter 18]. Also, the correct temperature range must be found to use the proper set of a and b coefficients; this may be done iteratively. Once the zone air temperature is found, the actual system heat transfer rate may be found directly from Equation (2). Beyond treatment of system heat transfer, other considerations that may be important in building energy analysis programs include treatment of radiant cooling and heating systems, treatment of interzone heat transfer, modeling convection heat transfer, and modeling radiation heat transfer. The heat balance method in Chapter 18 assumes the use of a single design day. In a building energy analysis program, it is most commonly used with a year’s worth of design weather data. In this case, the first day of the year is usually simulated repeatedly until a steady-periodic response is obtained. Then, each day is simulated sequentially, and, where needed, historical data for surface temperatures and heat fluxes from the previous day are used. When radiant cooling and heating systems are evaluated, the radiant source should be identified as a room surface. The calculation procedure considers the radiant source in the heat balance analysis. Therefore, the heat balance method is preferred over the weighting-factor method for evaluating radiant systems. Strand and Pedersen (1997) describe implementation of heat source conduction transfer functions that may be used for modeling radiant panels within a heat balance-based building simulation program. In principle, this method extends directly to multiple spaces, with heat transfer between zones. In this case, some surface temperatures appear in the surface heat balance equations for two different zones. In practice, however, the size of the coefficient array required for solving the simultaneous equations becomes prohibitively large, and the solution time excessive. For this reason, some programs solve only one space at a time and assume that adjacent space temperatures are the same as the prior time step, or use iterative schemes to achieve a simultaneous solution. Other approaches may remove this limitation (Walton 1980). Relatively simple exterior and interior convection models may be used for design cooling load calculation procedures. However, more sophisticated exterior convection models (Cooper and Tree 1973; Fracastoro et al. 1982; Melo and Hammond 1991; U.S.
Department of Energy 1996-2016; Walton 1983; Yazdanian and Klems 1994) that incorporate the effects of wind speed, wind direction, surface orientation, etc., may be preferable. More detailed interior convection correlations for use in buildings are also available (Alamdari and Hammond 1982, 1983; Altmayer et al. 1983; Bauman et al. 1983; Bohn et al. 1984; Chandra and Kerestecioglu 1984; U.S. Department of Energy 1996-2016 Goldstein and Novoselac 2010; Khalifa and Marshall 1990; Peeters et al. 2011; Spitler et al. 1991; Walton 1983). Also, more detailed models of exterior [e.g., Cole (1976); Walton (1983)] and interior [e.g., Carroll (1980); Davies (1988); Kamal and Novak (1991); Steinman et al. (1989); Walton (1980)] long-wave radiation transfer have been implemented in detailed building simulation programs.
Weighting-Factor Method The weighting-factor method of calculating instantaneous space sensible load is a compromise between simpler methods (e.g., steady-state calculation) that ignore the ability of building mass to store energy, and more complex methods (e.g., complete energy balance calculations). With this method, space heat gains at constant space temperature are determined from a physical description of the building, ambient weather conditions, and internal load profiles. Along with the characteristics and availability of heating and cooling systems for the building, space heat gains are used to calculate air temperatures and heat extraction rates. This discussion is in terms of heat gains, cooling loads, and heat extraction rates. Heat losses, heating loads, and heat addition rates are merely different terms for the same quantities, depending on the direction of the heat flow. The weighting factors represent Z-transfer functions (Kerrisk et al. 1981; York and Cappiello 1982). The Z-transform is a method for solving differential equations with discrete data. Two groups of weighting factors are used: heat gain and air temperature. Heat gain weighting factors represent transfer functions that relate space cooling load to instantaneous heat gains. A set of weighting factors is calculated for each group of heat sources that differ significantly in the (1) relative amounts of energy appearing as convection to the air versus radiation, and (2) distribution of radiant energy intensities on different surfaces. Air temperature weighting factors represent a transfer function that relates room air temperature to the net energy load of the room. Weighting factors for a particular heat source are determined by introducing a unit pulse of energy from that source into the room’s network. The network is a set of equations that represents a heat balance for the room. At each time step, including the initial introduction, the energy flow to the room air represents the amount of the pulse that becomes a cooling load. Thus, a long sequence of cooling loads can be generated, from which weighting factors are calculated. Similarly, a unit pulse change in room air temperature can be used to produce a sequence of cooling loads. A two-step process is used to determine the air temperature and heat extraction rate of a room or building zone for a given set of conditions. First, the room air temperature is assumed to be fixed at some reference value, usually the average air temperature expected for the room over the simulation period. Instantaneous heat gains are calculated based on this constant air temperature. Various types of heat gains are considered. Some, such as solar energy entering through windows or energy from lighting, people, or equipment, are independent of the reference temperature. Others, such as conduction through walls, depend directly on the reference temperature. A space sensible cooling load for the room, defined as the rate at which energy must be removed from the room to maintain the reference value of the air temperature, is calculated for each type of instantaneous heat gain. The cooling load generally differs from the instantaneous heat gain because some energy from heat gain is
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Energy Estimating and Modeling Methods
19.11
absorbed by walls or furniture and stored for later release to the air. At time , the calculation uses present and past values of the instantaneous heat gain (q, q–1), past values of the cooling load (Q–1, Q–2, ...), and the heat gain weighting factors (v0, v1, v2, ..., w1, w2, ...) for the type of heat gain under consideration. Thus, for each type of heat gain q, cooling load Q is calculated as
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Q = v0q + v1q–1 + … – w1Q–1 – w2Q–2 – …
(4)
The heat gain weighting factors are a set of parameters that determine how much of the energy entering a room is stored and how rapidly stored energy is released later. Mathematically, the weighting factors are coefficients in a Z-transfer function relating the heat gain to the cooling load. These weighting factors differ for different heat gain sources because the relative amounts of convective and radiative energy leaving various sources differ and because the distribution of radiative energy can differ. Heat gain weighting factors also differ for different rooms because room construction affects the amount of incoming energy stored by walls or furniture and the rate at which it is released. Sowell (1988) showed the effects of 14 zone design parameters on zone dynamic response. After the first step, cooling loads from various heat gains are added to give a total cooling load for the room. In the second step, the total cooling load is used (with information on the room’s HVAC system and a set of air temperature weighting factors) to calculate the actual heat extraction rate and air temperature. The actual heat extraction rate differs from the cooling load (1) because, in practice, air temperature can vary from the reference value used to calculate the cooling load, or (2) because of HVAC system characteristics. Deviation of air temperature t from the reference value at hour is calculated as t = 1/g0 + [(Q – ER) + P1(Q–1 – ER–1) + P2(Q–2 – ER–2) + … – g1t–1 – g2t–2 – …]
(5)
where ER is the energy removal rate of the HVAC system at hour , and g0, g1, g2, …, P1, P2, … are air temperature weighting factors, which incorporate information about the room, particularly thermal coupling between the air and the storage capacity of the building mass. Example values of weighting factors for typical building rooms are presented in the following table. One of the three groups of weighting factors, for light, medium, and heavy construction rooms, can be used to approximate the behavior of any room. Some automated simulation techniques allow weighting factors to be calculated specifically for the building under consideration. This option improves the accuracy of the calculated results, particularly for a building with an unconventional design. McQuiston and Spitler (1992) provided electronic tables of weighting factors for a large number of parametrically defined zones. Normalized Coefficients of Space Air Transfer Functions Room Envelope Construction Light
g 0*
g 1*
g 2*
W/(m2 · K)
P0
P1
Dimensionless
+9.54
9.82
+0.28
1.0
0.82
Medium
+10.28
10.73
+0.45
1.0
0.87
Heavy
+10.50
11.07
+0.57
1.0
0.93
Two assumptions are made in the weighting-factor method. First, the processes modeled are linear. This assumption is necessary because heat gains from various sources are calculated independently and summed to obtain the overall result (i.e., the superposition principle is used). Therefore, nonlinear processes such as radiation or natural convection must be approximated linearly. This assumption is not a significant limitation because these processes can be linearly approximated with sufficient accuracy for most
calculations. The second assumption is that system properties influencing the weighting factors are constant (i.e., they are not functions of time). Often, a single set of weighting factors is used during the entire simulation period; however, multiple sets of weighting factors can be used to represent daily or seasonal variations. This assumption can limit the use of weighting factors in situations where important room properties vary more frequently during the calculation (e.g., the distribution of solar radiation incident on the interior walls of a room, which can vary over the day, and indoor surface heat transfer coefficients). When the weighting-factor method is used, a combined radiative/ convective heat transfer coefficient is used as the indoor surface heat transfer coefficient. This value is assumed constant even though, in a real room, (1) radiant heat transferred from a surface depends on the temperature of other room surfaces (not on room air temperature) and (2) the combined heat transfer coefficient is not constant. Under these circumstances, an average value of the property must be used to determine the weighting factors. Cumali et al. (1979) investigated extensions to the weighting-factor method to eliminate this limitation. Weighting factors are derived by introducing a unit pulse in a heat balance analysis. These analyses are performed relatively quickly as a preprocess step. Once the weighting factors are established, the simulation time can be significantly reduced, compared with performing a heat balance at every time step. There is an implicit trade-off between the limitations of the weighting-factor method and its improved computation speed compared with that of the heat balance method.
Comprehensive Room Transfer Function The comprehensive room transfer function method (CRTF) is important as a reasonably accurate (comparable to weighting factor) model that is fast enough to be embedded in optimizations. One important application is model-predictive control. CRTF parameters may be estimated using forward (Seem et al. 1989) or inverse (Armstrong et al. 2006b; Gayeski et al. 2012) modeling. The interior terminating point of each wall is a common star node. Instead of separate surface flux and temperature for each wall, only the net star-node flux and temperature are evaluated. Exogenous radiant fluxes (solar and radiant shares of internal gains) are imposed on the star node as well. Radiant and convective exchange between walls also occurs through the star node. Convective heating and cooling by the system, on the other hand, as well as infiltration, ventilation, and the convective share of internal gains, enter the room model through its air node, which is coupled to the star node by a single, relatively small resistance. The resulting model has the topology of n conduction transfer functions terminating on a massless star node that is connected by a single resistance to an air capacitance node. However the c and d coefficients of all walls are combined, thus reducing the computational effort at each simulation time step. The CRTF thus has (up to the star node) the mathematical form of a multivariate autoregressive-moving-average with exogenous inputs (ARMAX) model. Methods of evaluating the common c and d coefficients and the star and air node resistances are described by Armstrong et al. (1992) and Seem et al. (1989).
Thermal-Network Methods Although implementations of the thermal-network method vary, they all have in common the discretization of the building into a network of nodes connected by heat transfer paths. In many respects, thermal-network models may be considered a variant of the heat balance method. Thermal-network models can include an arbitrary number of nodes as required for the problem under consideration. For example, heat balance models generally use simple methods for distributing radiation from lights; thermal-network models may model the lamp, ballast, and luminaire housing separately.
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2017 ASHRAE Handbook—Fundamentals (SI)
Thermal-network models depend on a heat balance at each node to determine node temperature and energy flow between all connected nodes. Energy flows may include conduction, convection, and short- or long-wave radiation. Methods have been developed that reduce the number of node interconnections (e.g., by replacing the general delta radiant exchange network with a star network) (Carroll 1980, 1981). For any mode of energy flow, a range of finite-difference or finite-volume techniques may be used to model the energy flow between nodes. Taking transient conduction heat transfer as an example, the simplest thermal-network model would be a one- or two-capacitor resistance/capacitance network (Hammarsten 1987; Sonderegger 1977; Sowell 1990). Others have used more refined network discretization (Clarke 2001; Lewis and Alexander 1990; Walton 1993). Advanced thermal-network models generally use a set of algebraic and differential equations. In most implementations, the solution procedure is separated from the models so that, in theory, different solvers might be used to perform the simulation. In contrast, in most heat balance and weighting factor programs the solution technique takes advantage of a constrained model structure. Various solution techniques have been used in conjunction with thermal-network models. Examples include graph theory combined with Newton-Raphson and predictor/corrector ordinary differential equation integration (Buhl et al. 1990) and the use of Euler explicit integration combined with sparse matrix techniques (Walton 1993). Of the four sensible-load zone models discussed, thermalnetwork models are the most flexible and have the greatest potential for high accuracy. However, they also require the most computation time, and, in current implementations, require more user effort to take advantage of the flexibility.
3.2
ENVELOPE COMPONENT MODELING
Above-Grade Opaque Surfaces Heat transfer through above-grade, opaque envelope components (e.g., walls, roofs, ceilings) is often approximated using transient one-dimensional (1D) calculations. This approximation is common to most building energy simulation tools. Two- and threedimensional effects, such as thermal bridging, are often accounted for by adjusting the properties of the materials in a 1D construction. Chapter 25 discusses the calculation of heat transfer through opaque building surfaces, including the overall effective thermal transmittance (or U-factor) of a flat building assembly. Thermal properties of building materials and other solids can be found in Chapter 26 (Table 1) and Chapter 33 (Table 3), respectively. Transient 1D heat transfer is calculated for multilayered constructions using one of three methods: • Response factors (Kusuda 1969). These factors relate the heat flux through the surface to temperature changes. The response factors are generated by calculating the response to a simple transient event (e.g., a step change) and observing the resulting changes in heat flux over time. Once obtained, response factors can be reused throughout a simulation to estimate the heat flux response to other transient temperature changes. • Conduction transfer functions (CTFs). Similar to response factors, CTFs relate the current heat flux through a surface to the heat fluxes of previous time steps. CTFs are also generated by imposing a simple transient event and observing the resulting heat flux over time. • Direct numerical methods (e.g., finite difference, finite element). These methods are required for nonlinear analysis such as for phase change materials and variable thermal properties.
Iu and Fisher (2004) provide an overview and comparison of response factors and CTFs.
Below-Grade Opaque Surfaces Thermal modeling of building foundations (Claridge et al. 1993), including guidelines for placement of insulation, is described in Chapter 27 of this volume and Chapter 44 of the 2015 ASHRAE Handbook—HVAC Applications. Chapter 18 of this volume provides information for calculating transmission heat losses through slab foundations and through basement walls and floors. These calculations are appropriate for design loads but are not intended for estimating annual energy usage. This section provides information about calculation methods suitable for energy estimates over time periods of arbitrary length. The magnitude of foundation heat transfer relative to other loads in the building depends on several factors, including the insulation design of the foundation and the shape and size of the foundation relative to the overall size of the building. Foundation heat losses (or gains) are more significant in buildings with lower foundation areato-perimeter ratios (Bahnfleth and Pedersen 1990; Kruis 2015). Low area-to-perimeter ratios often coincide with small buildings (e.g., detached residential construction) and building footprints with narrow cross sections. Where higher ground-coupled floor area-toperimeter ratios may occur (e.g., in warehouses, shopping malls, other low-rise commercial buildings), thermal mass effects related to the core-area (nonperimeter) portion of the ground-coupled floor may also be important. Ground-coupled foundation heat transfer involves threedimensional (3D) thermal conduction, moisture transport, and the long-time-constant heat storage properties of the ground. During the early 1990s, only simplified models could be routinely used for calculating ground heat transfer. These models were based on 1D steady-state conduction or 1D dynamic thermal diffusion modeling. Because of continuing increases in computing power, the state of the art in ground heat transfer modeling has improved. Several calculation methods have been developed and applied to building energy simulation software (Bahnfleth and Pedersen 1990; Beausoleil-Morrison 1996; Beausoleil-Morrison and Mitalas 1997; Clements 2004; ISO 1998). These methods are further described and compared to more detailed 3D numerical methods (Ben-Nakhi 2007; Crowley 2007; Deru 2003; Thornton 2007) as part of the IEA BESTEST ground-coupled heat transfer modeling test cases (Neymark et al. 2008). Other examples of ground heat transfer modeling techniques applied to floors and basements are described in Andolsun et al. (2010), Krarti (1994a, 1994b), and Krarti and Chuangchid (1999), and a simplification of that method in Chapter 19 of the 2009 ASHRAE Handbook—Fundamentals, and in Krarti et al. (1988a, 1988b) and Winkelmann (2002). Methods of estimating foundation heat transfer can be highly constrained simplifications to ease calculation requirements, or very detailed and computationally intensive. Although the state of the art in ground heat transfer modeling is improving, many wholebuilding simulation tools still rely on a loosely defined set of ground temperatures that are applied as an exterior boundary condition for a simplified 1D approximation. These temperatures are often based on lagged, monthly average outdoor dry-bulb temperatures. In reality, the temperature in the ground varies both spatially and temporally, and is heavily influenced by the presence of a conditioned building. Finding a balance between computation speed and numerical accuracy is the focus of the dissertation work performed by Kruis (2015). Kruis found that two-dimensional (2D) calculations capture much of the accuracy of the more detailed 3D methods with computation times reasonable for whole-building energy simulation tools.
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Energy Estimating and Modeling Methods Fenestration Fenestration systems (windows, skylights, and doors) affect building energy use through four basic mechanisms: (1) thermal heat transfer, (2) solar heat gain, (3) air leakage, and (4) daylighting. Details of each of these effects are covered in Chapter 15. Fenestration performance is characterized primarily by its overall coefficient of heat transfer (U-factor), solar heat gain coefficient (SHGC), and visible transmittance (VT). These metrics are typically used to rate fenestration products for a specific set of conditions that are not necessarily representative of the range of conditions encountered in a typical whole-building energy simulation. The energy impact of fenestration systems depends on the thermal and spectral properties of each component, including
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• • • • •
Glazing substrates (panes) Glazing coatings Fill gas Opaque frame and divider Spacers between panes
Fenestration system models span a wide range of complexity and input requirements. Some of the simpler models apply a constant Ufactor and SHGC to the system, whereas more advanced models may require finer details of the fenestration system that are not often available to energy modelers, such as • Glazing reflectance (front and back) and transmittance at wavelengths spanning the solar, visible, and infrared spectrums, as well as at different angles of solar incidence • Temperature-dependent thermal properties of fill gases • Two-dimensional heat transfer through frames, dividers, and spacers Many of these inputs can be found or generated in the suite of fenestration-related software tools developed by Lawrence Berkeley National Lab (LBNL 2016): • WINDOW (for analyzing composite fenestration system thermal and optical properties) • THERM (for analyzing two-dimensional heat transfer through building products) • Optics (for analyzing optical properties of glazing systems) • International Glazing Database (IGDB) (optical data for glazing products) In most energy modeling applications, the level of detail required to compose a fenestration system using these software tools is not readily available. Instead, energy modelers are at most provided with the product’s rated U-factor, SHGC, and VT, or specific values defined by compliance rules and incentive programs (e.g., ASHRAE Standard 90.1, ENERGY STAR®). With this use case in mind, Arasteh et al. (2009) developed an approach in EnergyPlus to represent fenestration systems with angle-dependent optical properties using only the rated U-factor, SHGC, and VT. This approach offers a higher level of accuracy than the assumption of constant values, but does not explicitly simulate the opaque frame and dividers components or the fill gas in the fenestration system.
Infiltration Infiltration is defined as the flow of outdoor air via (1) leakage through unintentional openings (e.g., cracks, porosities) in the building envelope and (2) natural ventilation through exterior windows, doors, vents, etc. It is typically treated as an additional load for the purposes of whole-building energy modeling. The flow is driven by pressure differences and buoyancy forces, and may be counteracted by pressurizing the building or by sealing openings with air barriers, weatherstripping, and similar measures. Outdoor air that has entered a conditioned space will at some point need to be heated or cooled to
19.13 maintain the space at the desired temperature. There are a number of procedures available for calculating infiltration and associated loads. Empirical models, based on the results of blower door tests, are most readily available for residential buildings because of more readily available measured data, such as effective leakage area and flow exponent. Models that treat the building as a single zone are most directly applicable to smaller buildings (e.g., houses). Models that treat buildings as a set of zones (not necessarily corresponding to thermal zones) are also available, but these models require parameters that may be difficult to determine without building-specific data. More information on these models and on the general issue can be found in Chapter 16. Discussion of natural ventilation modeling is included in this chapter’s section on Low-Energy System Modeling. In the absence of the information necessary to use a more detailed approach, infiltration rates from appropriate standards or guideline documents (e.g., ASHRAE Standard 90.1; Gowri et al. 2009) can be used to determine the inputs for an energy model. For commercial buildings, which are typically designed to be pressurized, this approach may be more successful than for residential buildings, which may not be pressurized.
3.3 INPUTS TO THERMAL LOADS MODELS Choosing Climate Data Environmental data provide boundary conditions for many energy estimating and modeling methods. Many energy estimating and modeling methods require climate data from specific sources or in specific formats. The most common application is full-year hourly simulation using a typical meteorological year (TMY) to represent typical building operation. Other applications may require historical climate data (for model calibration), real-time climate data (for model predictive controls), or projected climate data (for predicting the impact of climate change on building energy consumption). Climate data are available from a growing number of sources. Chapter 14 discusses the use of climatic design information in greater detail and includes data for more than 8000 locations worldwide. Hensen and Lamberts (2011) list several other sources of climate data and discuss the requirements of climate data for various simulation applications.
Internal Heat Gains Accurate representation of internal heat gains in building simulation is important for several reasons. These heat gains often compose a significant portion of HVAC system loads. In addition, heat sources such as lighting and office equipment may be significant direct contributors to building energy consumption. Typical sources of internal heat gain represented in a simulation include occupants, lighting, and plug loads. Some buildings also include significant heat from cooking, refrigeration, laboratory, or manufacturing equipment. The following simulation inputs are generally necessary to characterize a source of internal heat gain: • Peak heat rate. For example, installed lighting power. Note that the actual peak heat rate for sources such as office equipment is often lower than the equipment’s nameplate rating. • Time variation of heat rate. The heat gain from most sources varies by hour and may vary by day of the week and by season. • Latent heat fraction. Many sources of internal heat gain, such as lighting and most office equipment, produce only sensible heat. However, other sources, such as occupants and some cooking equipment, produce both sensible and latent heat. For those cases, the portion of the total heat gain entering the space as latent heat must be identified. • Radiant/convective split. Sensible heat gain has two components: radiant and convective. The relative proportion between the
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19.14
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two pathways affects the time delay between instantaneous heat gain and space load. Simulation programs often include default fractions for the split. Attention to these values is especially important when evaluating spaces with significant thermal mass. • Fraction of heat gain to space. Not all of the heat produced by some sources of internal heat gain ends up directly in the space. In those cases, it may be necessary to specify the portion that goes to the space and the portion that goes elsewhere. One example is recessed lighting in a suspended ceiling: some of the energy consumed by the lights enters the space and the rest goes to the space above the ceiling. Another example is cooking equipment under a range hood: a portion of the heat enters the kitchen, primarily via radiation, and the rest of the heat goes to the exhaust air. The same concept applies to laboratory equipment in a fume hood or to industrial equipment where a portion of the heat output is captured by an exhaust system. Another example is a space using a displacement ventilation air delivery strategy: a portion of the heat gain from occupants and equipment in the occupied zone rises into the stratified zone or return air and does not end up as a load in the occupied zone. Appropriate values for internal heat gain inputs may be different between energy calculations and HVAC system peak-load calculations. Although the basic form of input may be identical (e.g., both may require specifying office equipment power density), the objectives of the analyses are different. The typical objective of energy estimation is to determine likely energy consumption rather than worst-case energy consumption. Therefore, the appropriate values for internal heat gain inputs will typically be lower for energy estimation than for load calculations. ASHRAE research project RP-1093 developed different sets of diversity factors appropriate for energy calculations and for load calculations (Abushakra et al. 2000). In many cases, information about actual internal heat gains is not available to the energy modeler, either because it is a new construction project and detailed equipment specifications are not available, or the project is an existing building and a detailed survey of heat sources is not practical. Therefore, it may be necessary to make assumptions. Chapter 18 provides valuable data on peak heat rates, radiant/convective splits, and fraction of heat to space for a range of equipment types. Other sources of internal gain assumptions include COMNET (2016), Building America simulation protocols (Wilson et al. 2014), and DOE reference buildings (Deru et al. 2011).
Occupant Behavior Technologies alone do not necessarily guarantee low energy use in buildings. Occupant behavior has a significant effect on building performance. Occupants’ expectations of satisfaction with their indoor environment drive various actions, such as adjusting thermostat settings, opening windows, turning on lights, closing window blinds, consuming domestic hot water, and moving around, to satisfy their physical and nonphysical needs. These actions affect the built environment and energy use. Clearly understanding and accurately modeling occupant behavior in buildings is crucial to reducing the gap between design and actual building energy performance (Gunay et al. 2013; Hoes et al. 2009; Turner and Frankel 2008; Yan et al. 2015), especially for low-energy buildings relying more on passive design features, occupancy-controlled technologies, and occupant engagement. Occupant behavior is represented in predefined deterministic schedules or fixed settings to describe occupant movement and activities, the use of lighting, plug loads, and HVAC system operation. Assumptions of occupant input can be found in ASHRAE publications (e.g., Handbook, standards, guidelines), DOE reference buildings, simulation guides (COMNET 2016), building design documents, or measured data from existing buildings. Abushakra
2017 ASHRAE Handbook—Fundamentals (SI) and Claridge (2008) developed a method for modeling occupancy in commercial buildings based on monitored lighting and receptacles energy use. The method provides typical occupancy load shapes that can be used in both forward building energy simulations and data-driven models. When these static schedules and settings are entered into building energy modeling (BEM) programs, the results are deterministic and homogeneous, ignoring the stochastic nature, dynamics, and diversity of occupant behavior. For example, occupants can open windows for various reasons: (1) feeling hot (thermal comfort driven), (2) feeling stuffy (indoor air quality driven), (3) arriving in a space (event driven), and (4) personal habit. Alternative approaches used to integrate occupant behavior into models include embedded behavior modules adopted by BEM programs or agent-based modeling approaches that specify how agents (in this case, occupants) interact with one another and with their environment. Integration of occupant behavior with BEM programs ranges from built-in user functions to more flexible cosimulation. The user functions approach allows users, to a certain degree of flexibility, to write custom code to implement new or overwrite existing default building operation and supervisory controls. The cosimulation approach allows separate behavior tools to run simultaneously and exchange occupant information in a collaborative manner with BEM programs (Hong et al. 2016a; Wetter et al. 2011). Agent-based modeling specifies occupant attributes, behavioral rules, memory, resources, decision-making sophistication, and procedures for modifying current behavioral rules. Andrews et al. (2011), Langevin (2014), and Robinson (2011) developed agentbased modeling tools for cosimulation. Despite advancements, there are significant and fundamental scientific problems remaining to be addressed in modeling occupant behavior, including (1) choosing the best modeling approach based on available data and for a particular application, (2) standardizing the representation of occupant behavior for interoperability (Hong et al. 2015), (3) selecting methods to evaluate occupant behavior models, (4) interpreting the simulation results from the use of stochastic models, and (5) considering social and behavioral influencing factors. IEA EBC (2013-2017) Annex 66 addresses some of these problems and provides a literature database on occupant behavior research.
Thermal Zoning Strategies Configuration of thermal zones for building energy simulation commonly follows the core and perimeter zone method (Hirsch 2003; LBNL 1984.). The core and perimeter approach divides a building into an approximate perimeter of thermal zones within 4.5 to 6 m from the exterior walls and then further divides the building into separate perimeter zones for each orientation. A commonly used thermal zoning guideline is provided in ASHRAE Standard 90.1, Appendix G, which describes the standard’s performance rating method. The standard allows combining two or more actual thermal zones into a single zone if they have the same space usage, their exterior glazed surfaces have the same orientation (within 45°), and they are served by the same type of HVAC system. In certain cases, spaces that have similar thermal conditions may be aggregated into fewer zones within a simulation model with only a minor difference in simulation outcome, as stated in Lokmanhekim (1971), who proposed a method to define thermal zones by grouping spaces based on comparison of load profile plots. Several recent studies have considered different approaches for thermal zoning: • Raftery (2011) developed a method that defines the various type of thermal zones in a simulation based on four major criteria: (1) function of the space, (2) position of the zone relative to the exterior, (3) available measured data, and (4) method used to
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Energy Estimating and Modeling Methods condition the zone. Raftery’s method yields a more detailed thermal zoning plan than the traditional core and perimeter zoning method. However, the method is not automated, relying instead on a user’s subjective inputs to a simulation program. • Georgescu et al. (2012) developed a method that analyzes a detailed building energy model using an optimization method to develop zoning approximations based on observations of zone temperature. However, in this method, a detailed model of all zones must first be created to establish a baseline simulation for comparison; this can take a significant amount of time and can introduce uncertainty as the number of parameters in the model increases. Like Raftery (2011), the Georgescu et al. (2015) method also lacks an automated procedure. • Yi (2015) developed a user interface that suggests optimized thermal-zone layouts based on space arrangement. However, Yi’s method is based only on the spatial functions and does not consider varying internal load, space use, or system type. This method relies on the user’s judgment to perform the final selection of zones. • Dogan et al. (2015) present an algorithm to automatically convert arbitrary building massing models into multiple-zone building energy models. The applications include model generation during schematic design and generation of models for urban massing studies. Although these efforts provide promising steps toward improved thermal zoning for building energy simulation, a truly automated, one-size-fits-all approach remains to be developed.
4. HVAC COMPONENT MODELING 4.1
MODELING STRATEGIES
The performance of HVAC components varies based on component design, operating conditions, and control strategies. Two different strategies are commonly used to represent the performance of HVAC components: empirical (regression-based) models and firstprinciples models.
Empirical (Regression-Based) Models Although some secondary components (e.g., heat exchangers, valves) are readily described by fundamental engineering principles, the complex nature of some equipment (e.g., chillers, boilers, fans coupled with air distribution systems) has discouraged the use of first-principles models for energy calculations in favor of regression-based component models based on manufacturers’ empirical performance data. Energy consumption of complex equipment generally depends on equipment design, load conditions, environmental conditions, and equipment control strategies. For example, chiller performance depends on the basic equipment design features (e.g., heat exchange surfaces, compressor design), temperatures and flow through the condenser and evaporator, and methods for controlling the chiller at different loads and operating conditions (e.g., inlet guide vane control or variable-speed compressor control on centrifugal chillers to maintain leaving chilled-water temperature set point). In general, these variables vary constantly and require calculations on an hourly or subhourly basis. Energy consumption characteristics of primary equipment have traditionally been modeled using equations developed by regression analysis of manufacturers’ published design data. Historically, published data were available only for full-load design conditions, and additional correction functions were used to correct the full-load data to part-load conditions. However, industry efforts to standardize reporting of performance data over a range of operating conditions (proposed ASHRAE Standard 205P) are making it possible to
19.15 capture the performance of specific equipment via automated performance maps or custom regression fits. Use of such standard-format data is expected to become the preferred approach for transmitting equipment performance characteristics to energy modeling software. The functional form of the regression equations and correction functions takes many forms, including exponentials, Fourier series, and, most commonly, second- or third-order polynomials. Selection of an appropriate functional form depends on the behavior of the equipment. In some cases, energy consumption is calculated using direct interpolation from data tables. A typical approach to modeling primary equipment in energy simulation programs is to assume the following functional form for equipment power consumption: P = PIR Load PIR = PIR nom f 1 t a, t b, f 2 PLR
(6)
C avail = C nom f 3 t a t b Load PLR = --------------C avail
(7)
where P PIR PIRnom Load Cavail Cnom f1
= = = = = = =
f2 = f3 = ta, tb = PLR =
equipment power, kW power input ratio power input ratio under nominal full-load conditions power delivered to load, kW available equipment capacity, kW nominal equipment capacity, kW function relating full-load power at off-design conditions (ta, tb, …) to full-load power at design conditions fraction full-load power function, relating part-load power to full-load power function relating available capacity at off-design conditions (ta, tb, …) to nominal capacity various operating temperatures that affect power part-load ratio
The part-load ratio is the ratio of the load to the available equipment capacity at given off-design operating conditions. Like the power, the available, or full-load, capacity is a function of operating conditions. The particular forms of off-design functions f1 and f3 depend on the specific type of equipment. For example, for fossil-fuel boilers, full-load capacity and power (or fuel use) can be affected by thermal losses to ambient temperature. However, these off-design functions are typically considered to be unity in most building simulation programs. For condensing boilers, efficiency is significantly affected by boiler entering water temperature, which is therefore an important input for the f1 function (see also the section on Boilers). For chillers, both capacity and power are affected by condenser and evaporator temperatures, which are often characterized in terms of their secondary fluids. For direct-expansion air-cooled chillers, operating temperatures are typically the wet-bulb temperature of air entering the evaporator and the dry-bulb temperature of air entering the condenser. For liquid chillers, the temperatures are usually the leaving chilled-water temperature and the entering condenser water temperature (see also the section on Chillers). For fans and pumps, energy consumption is often represented as a polynomial function of airflow or fluid flow, as described in the section on Fans, Pumps, and Distribution Systems. As an example, consider the cooling performance of a directexpansion (DX) packaged single-zone rooftop unit. The nominal rated performance of these units is typically given for an outdoor air temperature of 35°C and evaporator entering coil conditions of 26.7°C db and 19.4°C wb. However, performance changes as
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19.16
2017 ASHRAE Handbook—Fundamentals (SI) Table 3
Corr. f1 f3
0
Correlation Coefficients for Off-Design Relationships
1
–1.063931 0.8740302
2
0.0306584 0.0011416
0.0001269 0.0001711
3
4
0.0154213 –0.002957
0.0000497 0.0000102
5 0.0002096 0.0000592
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conditions. However, first-principles physical models generally have several advantages over pure regression models:
Fig. 5 Possible Part-Load Power Curves outdoor temperature and entering coil conditions vary. To account for these effects, the DOE-2.1E simulation program expresses the off-design functions f1 and f3 with biquadratic functions of the outdoor dry-bulb temperature and the coil entering wet-bulb temperature. Many other programs use similar functions. f 1 t wb ent t oa = a 0 + a 1 t wb, ent 2
2
(8)
2
(9)
+ a 2 t wb, ent + a 3 t oa + a 4 t oa + a 5 t wb, ent t oa f 3 t wb ent t oa = c 0 + c 1 t wb, ent 2
+ c 2 t wb, ent + c 3 t oa + c 4 t oa + c 5 t wb, ent t oa
The constants in Equations (8) and (9) are given in Table 3. The fraction full-load power function f2 represents the change in equipment efficiency at part-load conditions and depends heavily on the control strategies used to match load and capacity. Figure 5 shows several possible shapes of these functional relationships. Curve 1 represents equipment with constant efficiency, independent of load. Curve 2 represents equipment that is most efficient in the middle of its operating range. Curve 3 represents equipment that is most efficient at full load. Note that these types of curves apply to both boilers and chillers.
First-Principles Models Engineering first principles such as thermodynamics as well as heat, mass, and momentum transfer can be used to develop models of primary equipment as well as secondary components. Gordon and Ng (1994, 1995), Gordon et al. (1995), Jiang and Reddy (2003), Lebrun et al. (1999), and others have sought to develop such models in which unknown model parameters are extracted from measured or published manufacturers’ data. The energy analyst often must choose the appropriate model for the job. For example, a complex boiler model is not appropriate if the boiler operates at virtually constant efficiency. Similarly, a regression-based model might be appropriate when the user has a full dataset of reliable in-situ measurements of the plant. A regressionbased model may also be a good choice when manufacturer performance data is provided for a full range of anticipated operating
• Physical models allow confident extrapolation outside the range of available data. • Regression is still required to obtain values for unknown physical parameters. However, the values of these parameters usually have physical significance, which can be used to estimate default parameter values, diagnose errors in data analysis through checks for realistic parameter values, and even evaluate potential performance improvements. • The number of unknown parameters is generally much smaller than the number of unknown coefficients in the typical regression model. For example, the standard ARI compressor model requires as many as 30 coefficients, 10 each for regressions of capacity, power, and refrigerant flow. By comparison, a physical compressor model may have as few as four or five unknown parameters. Thus, physical models require fewer measured data. • Data on part-load operation of chillers and boilers have been notoriously difficult to obtain but are becoming more readily available from manufacturers. Part-load corrections often represent the greatest uncertainty in the regression models, while causing the greatest effect on annual energy predictions. By comparison, physical models of full-load operation often allow direct extension to part-load operation with little additional required data. Physical models of primary HVAC equipment can be found in many HVAC textbooks, but some of the models described here are specifically based on the work of Bourdouxhe et al. (1994a, 1994b, 1994c) in developing the ASHRAE HVAC 1 Toolkit (Lebrun et al. 1999). Each elementary component’s behavior is characterized by a limited number of physical parameters, such as heat exchanger heat transfer area or centrifugal compressor impeller blade angle. Values of these parameters are identified, or tuned, based on regression fits of overall performance compared to measured or published data. Although physical models are based on physical characteristics, values obtained through a regression analysis of manufacturers’ data are not necessarily representative of the actual measured values. Strictly speaking, the parameter values are regression coefficients with estimated values, identified to minimize the error in overall system performance. In other words, errors in the fundamental models of equipment are offset by over- or underestimation of the parameter values.
4.2
TERMINAL COMPONENTS
Terminal Units and Controls Although many terminal units do not directly consume energy, they often have a significant effect on the energy consumed by primary and secondary equipment. For example, the minimum airflow set point for a variable-air-volume (VAV) terminal unit with a reheat coil affects supply fan energy, cooling energy consumption of the chiller or direct-expansion compressor, heating energy consumption of the electric reheat coil or central plant boiler, and pump energy for delivery of chilled and hot water. Therefore, the representation of terminal units in energy-estimating programs deserves special attention. Several types of terminal units are commonly represented in energy-estimating programs: single-duct VAV boxes with and
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Energy Estimating and Modeling Methods
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without reheat coils, single-duct VAV boxes with parallel or series induction fans, dual-duct VAV boxes, chilled-beam induction units, fan-coils, radiant panels, baseboard radiators, and others. Chapters 4, 5, and 20 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment provide a description of many types of terminal units and their operation. When selecting an energy-estimating program, it is important to note that not all programs explicitly represent some of these types of terminal units or they may not directly represent the desired control scheme. The following considerations regarding accuracy and energy consumption are important when representing terminal units in energy-estimating programs: • VAV box minimum airflow set point. As noted previously, this setting has a significant impact on energy consumption in a VAV system, and using default inputs may lead to inaccurate model results. In some actual systems, the minimum airflow set point is varied over time in a demand-controlled ventilation scheme, and in those cases it is important to ensure that the selected energyestimating program can represent that strategy. • VAV box damper control in heating mode. In actual systems, VAV box airflow in heating mode may remain at a fixed minimum position or may modulate to provide increased airflow as heating load increases, typically using a specific heating maximum airflow set point that may be different from the maximum cooling airflow. This later strategy is often called reverseacting damper control (dual-maximum), and it can lead to reduced reheat energy. Many energy-estimating programs can represent both control strategies, and this is an important input to verify. • Fan power for VAV boxes with induction fans. In some cases, these small induction fans are relatively inefficient, and this is an important input to verify. • Control of parallel induction fans in VAV boxes. Typically, in a parallel induction VAV box, the fan runs only in heating mode. To accurately represent this fan operation, it is important to verify in the energy-estimating program how the control for this fan is specified. • Chilled-water temperature limits for chilled-beam systems. In a typical chilled beam design, there is a minimum setpoint for entering water temperature to avoid condensation on the coils. This set point is higher than that for a typical chilled-water distribution system, and it may affect the efficiency of the chilled-water system and the pumping energy. • Temperature input requirements for radiant heating systems and radiators. Different types of radiant heating terminal units have different hot-water temperature input requirements. The corresponding hot-water supply and return temperature requirements on the boilers will affect boiler operating efficiency, especially for condensing boilers.
Underfloor Air Distribution Underfloor air distribution (UFAD) is sometimes considered as an energy-efficient HVAC strategy, and a method to estimate its energy performance is a valuable design tool. UFAD systems offer potential for energy savings in several ways. These systems typically operate at higher cooling supply air temperatures compared to overhead air delivery systems, and offer potential for additional economizer energy savings. Primary cooling equipment efficiency may improve when providing higher-temperature supply air. In addition, if the UFAD system is operated to maintain partially mixed air distribution, as described in Chapter 57 of the 2015 ASHRAE Handbook—HVAC Applications, then the zone air distribution effectiveness may be better than that of an overhead air
19.17 delivery system and allow a reduction in outdoor air ventilation rate. A limitation of many energy-estimating programs is the assumption that air within thermal zones is fully mixed, and these programs cannot directly represent the performance of a UFAD system that is operated to create some temperature stratification. Approximate methods have been used by energy modelers, such as raising the thermostat set point in the UFAD zones to represent a higher average air temperature or representing the actual zone with two stacked zones in the model. See Energy Design Resources (2012) for more information. A method for explicit modeling of UFAD systems has been developed for the EnergyPlus program, with separate models for interior and perimeter zones (Webster et al. 2013).
Thermal Displacement Ventilation Thermal displacement ventilation (TDV) is another air delivery strategy that offers energy efficiency benefits. A displacement system is typically designed to achieve fully stratified air distribution, as described in Chapter 57 of the 2015 ASHRAE Handbook— HVAC Applications. The potential efficiency benefits are similar to UFAD systems. Modelers seeking to estimate the energy performance of a TDV system face the same basic limitations as with UFAD systems, and similar approximations have been used in practice. However, a method for modeling TDV represented by three nodes (floor temperature, occupied-zone temperature, and upper-zone temperature) is implemented in EnergyPlus (U.S. Department of Energy 19962016).
Radiant Heating and Cooling Systems Heating and cooling a space using heated and cooled panels is fundamentally different from using conditioned air. In a radiant system, a significant fraction of the sensible heating and cooling occurs by radiant heat transfer between the radiant panel and the surfaces in the space. This process is described in Chapter 6 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. As noted in the Heat Balance Method section, energy-estimating programs that use the heat balance method for space load calculations are generally preferred for representing the performance of radiant systems. Use of the heat balance method allows accounting for radiant heat transfer among surfaces in a space, and a heated or cooled panel may be represented as one of those surfaces.
4.3
SECONDARY SYSTEM COMPONENTS
Secondary HVAC systems generally include all elements of the overall building energy system between a central heating and cooling plant and the building terminal units and zones. The precise definition depends heavily on the building design. A secondary system typically includes air-handling equipment; air distribution systems with associated ductwork; dampers; fans; and heating, cooling, and humidity-conditioning equipment. They also include liquid distribution systems between the central plant and the zone and airhandling equipment, including piping, valves, and pumps. Although the exact design of secondary systems varies dramatically among buildings, they are composed of a relatively small set of generic HVAC components, including distribution components (e.g., pumps/fans, pipes/ducts, valves/dampers, headers/plenums, fittings) and heat and mass transfer components (e.g., heating coils, cooling and dehumidifying coils, liquid heat exchangers, air heat exchangers, evaporative coolers, steam injectors and other humidifiers). Most secondary systems can be described by simply connecting these components to form the complete system. Energy estimation through computer simulation sometimes mimics the modular construction of secondary systems by using modular
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19.18 simulation elements [e.g., the ASHRAE HVAC2 Toolkit (Brandemuehl 1993; Brandemuehl and Gabel 1994), the simulation program TRNSYS (Klein et al. 1994; TRNSYS 2012), and Annex 10 activities of the International Energy Agency (IEA ECBCS 1987)]. To the extent that the secondary system consumes energy and transfers energy between the building and central plant, an energy analysis can be performed by characterizing the energy consumption of the individual components and the energy transferred among system components. In many energy-estimating programs, secondary systems are represented by a mix of component models and simplified system models. For example, it is common for air and hydronic distribution systems to be represented by energy models that do not explicitly include components such as pipes/ducts, coils, and valves/dampers. Those methods are described in the following section. In this chapter, secondary components are divided into two categories: (1) distribution components and (2) heat and mass transfer components.
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Fans, Pumps, and Distribution Systems The distribution system of an HVAC system affects energy consumption in two ways. First, fans and pumps consume electrical energy directly, based on the flow and pressures under which a device operates. Ducts and dampers, or pipes and valves, and the system control strategies affect flow and pressure at each fan or pump. Second, thermal energy is often transferred to (or from) the fluid by (1) heat transfer through pipes and ducts and (2) electrical input to fans and pumps. Analysis of system components should, therefore, account for both direct electrical energy consumption and thermal energy transfer. Fan and pump performance are discussed in Chapters 21 and 44 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. In addition, Chapter 21 of this volume covers pressure loss calculations for airflow in ducts and duct fittings, and the effects of system air leakage. Chapter 22 presents a similar discussion for fluid flow in pipes. Although these chapters do not specifically focus on energy estimation, energy use is governed by the same performance characteristics and engineering relationships. Strictly speaking, performance calculations for a building’s fan and air distribution system require a detailed pressure balance on the entire network. For example, in an air distribution system, airflow through the fan depends on its physical characteristics, operating speed, and pressure differential across the fan. Pressure drop through the air distribution system depends on duct construction (i.e., friction and fitting losses), coil and filter characteristics, damper positions, air leakage, and airflow through each component. Interaction between the fan and air distribution system results in a set of coupled, nonlinear algebraic equations. Models and subroutines for performing these calculations are available in the ASHRAE HVAC2 Toolkit (Brandemuehl 1993) and CONTAM (NIST 2000-2016). A simplified, physics-based model of the system curve, which represents the system pressure drop as seen by the fan, is described by Sherman and Wray (2010). Detailed analysis of a distribution system requires flow and pressure balancing among the components, but nearly all commercially available energy analysis methods approximate the effect of the interactions with part-load performance curves [empirical (regressionbased) component models]. This approach eliminates the need to calculate pressure drop through the distribution system at off-design conditions. Two simplified methods are described in the following sections. Polynomial Curve Fit. Part-load curves are often expressed in terms of a power input ratio as a function of the part-load ratio, defined as the ratio of part-load flow to design flow:
2017 ASHRAE Handbook—Fundamentals (SI) Q W PIR = ------------ = f plr ----------- W full Q full
(10)
where PIR W Wfull Q Qfull fplr
= = = = = =
power input ratio fan motor input power at part load, W fan motor input power at full load or design, W fan airflow rate at part load, m3/s fan airflow rate at full load or design, m3/s regression function, typically polynomial or power law
The exact shape of the part-load curve depends on the effect of flow control on the pressure and fan efficiency and may be calculated using a detailed analysis or measured field data. Figure 6 shows the relationship for three typical fan control strategies, as represented in a simulation program (York and Cappiello 1982). In the simulation program, the curves are represented by polynomial regression equations. Models and subroutines for performing these calculations are also available in the ASHRAE HVAC2 Toolkit (Brandemuehl 1993). Figure 7 shows an example of a similar curve for the part-load operation of a fan system in a monitored building (Brandemuehl and Bradford 1999). In this case, the fan system represents 10 separate air handlers, each with supply and return fans, operating with variablespeed fan control to maintain a set duct static pressure. Component Model. A second method can be used for fan system energy estimation in variable-airflow systems. It accounts
Fig. 6
Part-Load Curves for Typical Fan Operating Strategies (York and Cappiello 1982)
Fig. 7
Fan Part-Load Curve Obtained from Measured Field Data under ASHRAE RP-823 (Brandemuehl and Bradford 1999)
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Energy Estimating and Modeling Methods
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separately for the substantial part-load performance variations that can occur for the fans, fan drive belts, fan motor, and variable-frequency drives (VFDs). For fans, dimensionless models are used to represent fan efficiency and speed variations with airflow. This method also uses a physics-based model (often called the “system curve”) for determining flow-dependent fan pressure rise (Sherman and Wray 2010). Note that system curves with duct static pressure controls or significant pressure drops through coils and filters do not follow the commonly assumed quadratic system curve. Fan power is calculated using airflow, pressure rise, and fan efficiency at each time step. Calculating fan power and speed allows determining fan shaft torque and the resulting loads on drive components (belts, motors, and VFDs). Note that fan laws do not apply to drive components. In the absence of physics-based models, regression models based on available data can be used to represent belt, motor, and VFD part-load efficiency variations. An implementation of this approach, developed by Wray in 2010, is included in EnergyPlus, and details are provided in the EnergyPlus Engineering Reference (U.S. Department of Energy 1996-2016). A benefit of this approach is its ability to directly represent the impacts of low-pressure-drop distribution systems, air leakage reduction, and duct static-pressure reset control, which are important efficiency strategies for many systems. This approach can also represent the effects of fan system component oversizing and changes for components such as fans, belts, motors, and variablefrequency drives. A disadvantage to the component model is that the additional capabilities require additional inputs. However, generic values for model parameters are provided in the EnergyPlus documentation. Fan and Pump Heat. Heat dissipated to the airstream by fan operation increases airstream temperature. Fan shaft power is usually assumed to be dissipated fully in the airstream. Motor losses also contribute if the motor is mounted in the airstream. For pumps, these contributions are typically assumed to be zero because the motor is not mounted in the water stream, and the ratio of transport power to thermal capacitance rate is usually much less for water than for air. The following equation provides a convenient and general model to calculate the heat transferred to the fluid by a motor: qfluid = [m+ (1 – m) fm, loss]W
(11)
where qfluid = heat transferred to fluid, W fm,loss = fraction of motor heat loss transferred to fluid stream, dimensionless (= 1 if fan mounted in airstream, = 0 if fan mounted outdoor airstream) W = fan motor input power, W m = motor efficiency
The heat rejected by VFDs and the fraction of motor heat not transferred to the airstream should be accounted for as building loads.
Heat and Mass Transfer Components Secondary HVAC systems comprise heat and mass transfer components (e.g., steam or hot-water air-heating coils, chilled-water cooling and dehumidifying coils, shell-and-tube liquid heat exchangers, air-to-air heat exchangers, evaporative coolers, steam injectors). Although these components do not consume energy directly, their thermal performance dictates interactions between building loads and energy-consuming primary components (e.g., chillers, boilers). In particular, secondary component performance determines the entering fluid conditions for primary components, which in turn determine energy efficiencies of primary equipment. In addition, heat and mass transfer components may affect air and fluid flow, which affect performance of energy-consuming secondary components such as fans and pumps. Accurate energy calcula-
tions cannot be performed without appropriate models of the system heat and mass transfer components. For example, load on a chiller is typically described as the sum of zone sensible and latent loads, plus any heat gain from ducts, plenums, fans, pumps, and piping. However, the chiller’s energy consumption is determined not only by the load but also by the return chilled-water temperature and flow rate. The return water condition is determined by cooling coil performance and part-load operating strategy of the air and water distribution system. The cooling coil might typically be controlled to maintain a constant leaving air temperature by modulating water flow through the coil. In such a scenario, the cooling coil model must be able to calculate the leaving air humidity, water temperature, and water flow rate given the cooling coil design characteristics and entering air temperature and humidity, airflow, and water temperature. Virtually all building energy simulation programs include, and require, models of heat and mass transfer components. These models are generally relatively simple. Whereas a coil designer might use a detailed tube-by-tube analysis of conduction and convection heat transfer and condensation on fin surfaces to develop an optimal combination of fin and tube geometry, an energy analyst is more interested in determining changes in leaving fluid states as operating conditions vary during the year. In addition, the energy analyst is likely to have limited design data on the equipment and, therefore, requires a model with very few parameters that depend on equipment geometry and detailed design characteristics. A typical approach to modeling heat and mass transfer components for energy calculations is based on an effectiveness-NTU heat exchanger model (Kays and London 1984). The effectiveness-NTU (number of transfer units) model is described in most heat transfer textbooks and briefly discussed in Chapter 4. It is particularly appropriate for describing leaving fluid conditions when entering fluid conditions and equipment design characteristics are known. Also, this model requires only a single parameter to describe the characteristics of the exchanger: the overall transfer coefficient UA, which can be determined from limited design performance data. Effectiveness methods are used to perform energy calculations for a variety of sensible heat exchangers in HVAC systems. For typical finned-tube air-heating coils, the cross-flow configuration with both fluid streams unmixed is most appropriate. Air-to-air heat exchangers may be cross- or counterflow. For liquid-to-liquid exchangers, tube-in-tube equipment can be modeled as parallel or counterflow, depending on flow directions; correlations for shelland-tube effectiveness and NTU, which depend on the extent of baffling and the number of tube passes, are given in heat transfer texts (Mills 1999). The energy analyst must determine the UA to describe the operations of a specific heat exchanger. There are three ways to determine this important parameter: direct calculation, measurement, and manufacturers’ data. Given detailed information about the materials, geometry, and construction of the heat exchanger, fundamental heat transfer principles can be applied to calculate the overall heat transfer coefficient. However, the method most appropriate for energy estimation is use of manufacturers’ performance data or measurement of installed performance. In reporting the design performance of a heat exchanger, a manufacturer typically gives the heat transfer rate under various operating conditions, with operating conditions described in terms of entering fluid flow rates and temperatures. The effectiveness and UA can be calculated from the given heat transfer rate and entering fluid conditions. Example 3. An energy analyst seeks to evaluate a hot-water heating system that includes a hot-water heating coil. The energy analysis program uses an effectiveness-NTU model of the coil and requires the UA of the coil as an input parameter. Although detailed information on the coil geometry and heat transfer surfaces is not available, the manufacturer
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states that the one-row hot-water heating coil delivers 240 kW of heat under the following design conditions: Design Performance Entering water temperature thi = 80°C Water mass flow rate m· h = 5.0 kg/s Entering air temperature tci = 20°C Air mass flow rate m· c = 8.0 kg/s Design heat transfer q = 240 kW Solution: First determine the heat exchanger UA from design data, then use UA to predict performance at off-design conditions. EffectivenessNTU relationships are used for both steps. The key assumption is that the UA is constant for both operating conditions. a) An examination of flow rates and fluid specific heats allows calculation of the hot-fluid capacity rate Ch and the cold-fluid capacity rate Cc at design conditions, and the capacity rate ratio Z. C = m· c = (5.0)(4.195) = 20.97 kW/K h
p h
Fig. 8 Psychrometric Schematic of Cooling Coil Processes
Cc = m· c p c = (8.0)(1.007) = 8.05 kW/K Cmax = Ch
Cmin = Cc
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C min Z = ------------ = 0.384 C max where cp is specific heat and cmax and cmin are the larger and smaller of the capacity rates, respectively, b) Effectiveness can be directly calculated from the heat transfer definition. t co – t ci q Cc 240 8.05 = ---------------------- = ---------------------- = ----------------------- = 0.497 t hi – t ci t hi – t ci 80 – 20 where tco is the leaving air temperature. c) The effectiveness-NTU relationships for a cross-flow heat exchanger with both fluids unmixed allow calculation of the effectiveness in terms of the capacity rate ratio Z and the NTU [the relationships are available from most heat transfer textbooks and, specifically, in Kays and London (1984)]. Given the effectiveness and capacity rate ratio, NTU = 0.804. d) The heat transfer UA is then determined from the definition of the NTU. UA = CminNTU = (8.05)(0.804) = 6.472 kW/K
Application to Cooling and Dehumidifying Coils Analysis of air-cooling and dehumidifying coils requires coupled, nonlinear heat and mass transfer relationships. These relationships form the basis for all HVAC components with moisture transfer, including cooling coils, cooling towers, air washers, and evaporative coolers. Although the complex heat and mass transfer theory presented in many textbooks is often required for cooling coil design, simpler models based on effectiveness concepts are usually more appropriate for energy estimation. For example, the bypass factor is a form of effectiveness in the approach of the leaving air temperature to the apparatus dew-point, or coil surface, temperature. The effectiveness-NTU method is typically developed and applied in analysis of sensible heat exchangers, but it can also be used to analyze other types of exchangers, such as cooling and dehumidifying coils, that couple heat and mass transfer. By redefining the state variables, capacity rates, and overall exchange coefficient of these enthalpy exchangers, the effectiveness concept may be used to calculate heat transfer rates and leaving fluid states. For sensible heat exchangers, the state variable is temperature, the capacity is the product of mass flow and fluid specific heat, and the overall transfer coefficient is the conventional overall heat transfer coefficient. For cooling and dehumidifying coils, the state variable becomes moist air enthalpy, the capacity has units of mass flow, and the overall heat transfer coefficient is modified to reflect enthalpy exchange. This approach is the basis for models by Brandemuehl
(1993), Braun et al. (1989), and Threlkeld (1970). The same principles also underlie the coil model described in Chapter 23 of the 2012 ASHRAE Handbook—HVAC Systems and Equipment. The effectiveness model is based on the observation that, for a given set of entering air and liquid conditions, the heat and mass transfer are bounded by thermodynamic maximum values. Figure 8 shows the limits for leaving air states on a psychrometric chart. Specifically, the leaving chilled-water temperature cannot be warmer than the entering air temperature, and the leaving air temperature and humidity cannot be lower than the conditions of saturated moist air at the temperature of the entering chilled water. Figure 8 also shows that performance of a cooling coil requires evaluating two different effectivenesses to identify the leaving air temperature and humidity. An overall effectiveness can be used to describe the approach of the leaving air enthalpy to the minimum possible value. An air-side effectiveness, related to the coil bypass factor, describes the approach of the leaving air temperature to the effective wet-coil surface temperature. Effectiveness analysis is accomplished for wet coils by establishing a common state variable for both the moist air and liquid streams. As implied by the lower limit of the entering chilled-water temperature, this common state variable is the moist air enthalpy. In other words, all liquid and coil temperatures are transformed to the enthalpy of saturated moist air at the liquid or coil temperature. Changes in liquid temperature can similarly be expressed in terms of changes in saturated moist air enthalpy through a saturation specific heat cp,sat defined by the following: h l sat cp, sat = ---------------t l
(12)
Using the definition of Equation (12), the basic effectiveness relationships discussed in Chapter 4 can be written as q = Ca(ha, ent – ha, lvg) = Cl (hl, sat, lvg – hl, sat, ent)
(13)
q = Cmin(ha, ent – hl, sat, ent)
(14)
Ca = m· a
(15)
m· c p l Cl = ---------------c p, sat
(16)
Cmin = min(Ca, Cl)
(17)
where q = heat transfer from air to water, W C = fluid capacity, kg/h
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Energy Estimating and Modeling Methods
19.21
m· a = dry air mass flow rate, kg/s m· l = liquid mass flow rate, kg/s cp,l = liquid specific heat, kJ/(kg·K) cp,sat = saturation specific heat, defined by Equation (12), kJ/(kg·K) ha = enthalpy of moist air, kJ/kg hl,sat = enthalpy of saturated moist air at the temperature of the liquid, kJ/kg
The cooling coil effectiveness of Equation (14) is defined, then, as the ratio of moist air enthalpies in Figure 8. As with sensible heat exchangers, effectiveness is also a function of the physical coil characteristics and can be obtained by modeling the coil as a counterflow heat exchanger. However, because heat transfer calculations are performed based on enthalpies, the overall transfer coefficient must be based on enthalpy potential rather than temperature potential. The enthalpy-based heat transfer coefficient UAh is related to the conventional temperature-based coefficient by the specific heat: q = UA t = UA h h
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UA t UA UA h = -------------- = -------h cp
(18)
A similar analysis can be performed to evaluate the air-side effectiveness, which identifies the leaving air temperature. Whereas the overall enthalpy-based effectiveness is based on an overall heat transfer coefficient between the chilled water and air, air-side effectiveness is based on a heat transfer coefficient between the coil surface and air. As with sensible heat exchangers, the overall heat transfer coefficients UA can be determined either from direct calculation from coil properties or from manufacturers’ performance data. A sensible heat exchanger is modeled with a single effectiveness and can be described by a single parameter UA, but a wet cooling and dehumidifying coil requires two parameters to describe the two effectivenesses shown in Figure 8. These parameters are the internal and external UAs: one describes heat transfer between the chilled water and the air-side surface through the pipe wall, and the other between the surface and the moist air. UA values can be determined from the sensible and latent capacity of a cooling coil at a single rating condition. A significant advantage of the effectiveness-NTU method is that the component can be described with as little as one measured data point or one manufacturer’s design calculation.
4.4
PRIMARY SYSTEM COMPONENTS
Primary HVAC systems consume energy and deliver heating and cooling to a building, usually through secondary systems. Primary equipment generally includes chillers, boilers, cooling towers, cogeneration equipment, and plant-level thermal-storage equipment. In particular, primary equipment generally represents the major energy-consuming equipment of a building, so accurate characterization of building energy use relies on accurate modeling of primary equipment energy consumption.
Boilers The literature on boiler models is extensive, ranging from steadystate performance models (DeCicco 1990; Lebrun 1993) to detailed dynamic simulation models (Bonne and Jansen 1977; Lebrun et al. 1985), to a combination of these two schemes (Laret 1991; Malmström et al. 1985). However, the input information required for these models is not often readily available, and these models should be considered only in more complex situations (e.g., large boilers in large buildings, district heating systems, cogeneration systems), where a complete, detailed representation of heat distribution, emission, and operation and control under varying external conditions is warranted.
In most current energy-estimating programs, the performance of boilers is represented by regression equations and a full-load efficiency value as described in the section on Empirical (RegressionBased) Models. Those equations typically represent the boiler efficiency under varying load conditions and, in some cases, the equations also account for water temperature conditions. When using these regression-based models, two important issues to understand are efficiency rating conditions and the off-design efficiency calculation method used by the energy-estimating program. Rating Conditions. The primary rating condition relevant for fuel-fired boiler efficiency is the entering water temperature, which is the return water temperature in most building heating applications. Efficiency improves as the return water temperature drops, because more heat can be extracted from exhaust gases. A condensing boiler can reach efficiencies up to 98% with entering water at 27°C. The minimum return water temperature for noncondensing boilers is typically about 60°C, because lower temperatures lead to condensation and damaging corrosion, and efficiency under that condition is usually in the range of 80% to 85%. The important point to consider when entering a value for fuel-fired boiler efficiency in an energyestimating program is the assumption made by the program about rating conditions. For example, the DOE2.2 program expects noncondensing boiler efficiency values to correspond to 71°C entering water temperature and condensing water boiler efficiency values to correspond to 27°C entering water temperature. More information on boiler design and rating conditions is provided in Chapter 32 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. Off-Design Conditions. Another important consideration in the use of energy-estimating programs is the type of model used to represent off-design conditions (i.e., how the efficiency is assumed to change at partial load or as the water temperature conditions change). Some models account only for changes in efficiency with changes in load. Other models consider both load and entering water temperature, and still others consider load and leaving water temperature. Some programs offer a choice of models. For best accuracy, models that account for load and entering water temperature are preferred. Figures 9 and 10 show examples of how boiler efficiency is represented by different types of models. Figure 9 shows a model in which boiler efficiency is a function of part-load ratio alone. Different regression equation coefficients are used for atmospheric-draft and forced-draft boilers, and the forced-draft units are represented with better part-load efficiency. An implicit assumption of this model is that entering water temperature is in a range where it does not have a significant impact on efficiency, which may be reasonable for noncondensing boilers. Figure 10 shows a model for a condensing boiler in which efficiency is a function of both part-load ratio and entering water temperature. This model shows that efficiency increases as entering water temperature decreases and as load decreases.
Chillers Models representing energy consumption of chillers include both first-principles models and empirical models. Examples of firstprinciples models include detailed modeling techniques described in the ASHRAE HVAC 1 Toolkit (Lebrun et al. 1999) and the GordonNg model (Gordon 2000). A discussion of the Gordon-Ng model is provided in the Data-Driven Modeling section, presented as an example of a physical model. From a practical point of view, the data required as inputs to these first-principles models will not always be available. Most current energy-estimating programs use empirical models to represent chiller performance. Hydeman et al. (2002) compares the prediction capabilities of several types of models. Empirical chiller models typically comprise a set of regression equations that represent variations in cooling capacity and efficiency in operation. A common implementation consists of three polynomial equations. One represents variation in cooling capacity
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2017 ASHRAE Handbook—Fundamentals (SI)
as a function of temperature conditions. A second represents the change in efficiency with changing temperature conditions. The third equation represents efficiency as a function of part-load ratio. The following are common inputs to empirical chiller models: • Full-load efficiency • Part-load ratio • Chilled-water supply temperature (temperature leaving the evaporator) • For water-cooled chillers: condenser water return temperature (temperature entering the condenser) • For air-cooled chillers: outdoor air dry-bulb temperature Different types of chillers have different performance characteristics, and their performance is typically represented by different
sets of regression equation coefficients. Important chiller design differences include type of condenser (air- or water-cooled), type of compressor (reciprocating, screw, or centrifugal), and unloading mechanism (e.g., inlet vanes or variable-speed compressor control). Chapter 43 of the 2016 ASHRAE Handbook—Systems and Equipment provides descriptions of different types of chillers. Coefficients for chiller models are often supplied as defaults by energy-estimating programs. However, when the intent of analysis is to compare the energy performance of specific chiller alternatives, the recommended practice is to obtain performance data for the specific chillers to create calibrated sets of coefficients for each chiller. Manufacturers can typically provide such data, and it is very important to obtain a set of data representing the full range of potential operating conditions. The process of calculating coefficients for a chiller model is described in Hydeman and Gillespie (2002). For another discussion of empirical chiller models and the process of determining model coefficients, see the HVAC Simulation Guidebook, vol. 1 (Energy Design Resources 2012).
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Cooling Tower Model
Fig. 9
Example Boiler Model: Efficiency as Function of Part-Load Ratio
A cooling tower is used in primary systems to reject heat from the chiller condenser. Controls typically manage tower fans and pumps to maintain a desired water temperature entering the condenser. Like cooling and dehumidifying coils in secondary systems, cooling tower performance has a strong influence on the chiller’s energy consumption. In addition, tower fans consume electrical energy directly. Fundamentally, a cooling tower is a direct contact heat and mass exchanger. Equations describing the basic processes are given in Chapter 6 and in many HVAC textbooks. Chapter 40 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment describes the specific performance of cooling towers. Performance subroutines are also available in Lebrun et al. (1999) and TRNSYS (2012). For energy calculations, cooling tower performance is typically described in terms of the outdoor wet-bulb temperature, temperature drop of water flowing through the tower (range), and difference between leaving water and air wet-bulb temperatures (approach). Simple models assume constant range and approach, but more sophisticated models use rating performance data to relate leaving water temperature to the outdoor wet-bulb temperature, water flow, and airflow. Simple cooling tower models, such as those based on a single overall transfer coefficient that can be directly inferred from a single tower rating point, are often appropriate for energy calculations.
Variable-Speed Vapor-Compression Heat Pump Model
Fig. 10 Example Boiler Model: Efficiency as Function of Part-Load Ratio and Entering Water Temperature
Although packaged equipment is most often modeled by empirical functions fit to manufacturers’ data, methods are also needed to model and evaluate advanced equipment such as variable-speed heat pumps. Zakula et al. (2011) detail a component-based heat pump model, including variable-speed fan and pump models; variable refrigerant-, air-, and water-side heat transfer coefficients; variable pressure drops; and staged compressors, one or more of which may operate over a wide speed range. Such models can produce performance maps for new equipment designs, including designs in which refrigerant and transport fluid flow rates and condenser subcooling are coordinated as optimized functions of conditions and imposed load (Zakula et al. 2012). A tricubic fit to such performance maps has been shown to represent performance accurately over a wide range of conditions and load. As a special heat pump model, variable-refrigerant-flow (VRF) systems vary the refrigerant flow to meet the dynamic zone thermal loads. Hong et al. (2016) developed a new VRF heat pump model, which has physics-based component models. This VRF model, implemented in EnergyPlus, (1) enables advanced controls, including variable evaporating and condensing temperatures in the indoor and
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Energy Estimating and Modeling Methods outdoor units, and variable fan speeds based on the temperature and zone load in the indoor units; (2) adds a detailed refrigerant pipe heat loss calculation using refrigerant flow rate, operational conditions, pipe length, and pipe insulation materials; (3) increases accuracy of simulation especially in partial load conditions; and (4) improves the usability of the model by significantly reducing the number of user input performance curves. A new heat recovery VRF model was also developed and is implemented in EnergyPlus, which enables more accurate modeling of heat recovery between zones with simultaneous cooling and heating loads. The new VRF models in EnergyPlus were validated with measured performance data from real buildings or laboratory experiments.
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Ground-Coupled Systems Ground-coupled systems use the ground as the source/sink for heating or cooling a building. These types of systems are described in detail in Chapter 34 of the 2015 ASHRAE Handbook—HVAC Applications. For the most part, the HVAC system components used in these systems are similar to those used in other HVAC systems and do not require any special techniques to estimate their performance. The exception to this is simulating energy exchange with the ground. Energy is exchanged with the ground through direct use of water from the ground or through ground heat exchangers. With a directuse system, the temperature entering the HVAC system from the ground is often considered to be constant or to vary only seasonally. This assumption can be valid when the extraction and injection wells are far enough apart to prevent mixing of the aquifer. A number of aquifer storage performance models have been developed for use with HVAC systems, but these have not yet reached mainstream simulation programs. For simulating performance of ground heat exchangers, two different techniques have been developed: the duct ground storage (DST) model introduced by Hellstrom (1989) that calculates the transient thermal process of borehole fields, and a thermal response factor (g-function) approach based on work of Eskilson (1987). Both techniques give a relation between the heat extracted (or rejected) from the ground per unit borehole length and the borehole wall temperature. Both of these methods have been integrated into whole-building simulation programs.
4.5
MODELING OF SYSTEM CONTROLS
Building control systems are hierarchical: higher-level, supervisory controls typically generate set points for lower-level, local loop controls. Supervisory controls include reset and optimal control (see Chapter 42 of the 2015 ASHRAE Handbook—HVAC Applications) and often have a large effect on energy consumption. Local loop controllers may also affect energy performance; for example, proportional-only room temperature control results in a trade-off between energy use and comfort. Faults in control systems and devices can also affect energy consumption (e.g., leaking valves and dampers can significantly increase energy use). It is particularly important to account for these departures from ideal behavior when simulating performance of real buildings using calibrated models. Modeling and simulation of some supervisory control functions are increasingly handled by whole-building simulation programs, but very advanced optimal controls such as receding horizon control are not. Simulation of local loop controls also requires more specialized, component- or equation-based modeling environments. Modern control systems, particularly direct digital controls (DDC), typically use integral action to drive the controlled variable to its set point. For energy modeling purposes, the controlled variable (e.g., supply air temperature) can be treated as being at the set point unless system capacity is insufficient. The simulation must determine whether the capacity required to meet set point exceeds
19.23 available capacity. If it does, the available capacity is used to determine the actual value of the controlled variable. Where there is only proportional action, the resulting relationship between the controlled variable and the output of the system can be used to determine both values. For example, the action of a conventional pneumatic room temperature controller can be represented by a function relating heating and cooling delivery to space temperature. Similarly, supply air temperature reset control can be modeled as a relationship between outdoor or zone temperature and coil or fan discharge temperature. An accurate secondary system model must ensure that all controls are properly represented and that the governing equations are satisfied at each simulation time step. This often creates a need for iteration or for use of values from an earlier solution point. Controls on space temperature affect the interaction between loads calculations and the secondary system simulation. A realistic model might require a dead band in space temperature in which no heating or cooling is called for; within this range, the secondary system operates at zero capacity, and the true space temperature must be modeled accordingly. If the thermostat has proportional control between zero and full capacity, the space temperature rises in proportion to the load during cooling and falls similarly during heating. Capacity to heat or cool also varies with space temperature after the control device has reached its maximum because capacity is proportional to the difference between supply and space temperatures. Failure to properly model these phenomena may result in overestimating energy use.
4.6
INTEGRATION OF SYSTEM MODELS
Energy calculations for secondary systems involve construction of the complete system from the set of HVAC components. As an example, consider a process for simulating a variable-air-volume (VAV) system, a single-path system that controls zone temperature by modulating airflow while maintaining constant supply air temperature. VAV terminal units, located at each zone, adjust the quantity of air reaching each zone depending on its load requirements. Reheat coils may be included to provide required heating for perimeter zones or to prevent overcooling of lightly loaded zones. This VAV system simulation consists of a central air-handling unit and a VAV terminal unit with reheat coil located at each zone, as shown in Figure 11.The central air-handling unit includes a fan, cooling coil, preheat coil, and outdoor air economizer. Supply air leaving the air-handling unit is controlled to a fixed set point. The VAV terminal unit at each zone varies airflow to meet the cooling load. As zone cooling load decreases, the VAV terminal unit decreases zone airflow until the unit reaches its minimum position. If the cooling load continues to decrease, the reheat coil is activated to meet the zone load. As supply air volume leaving the unit decreases, fan power consumption also reduces. A variable-speed drive is used to control the supply fan. The simulation is based on system characteristics and zone design requirements. For each zone, the inputs include sensible and latent loads, zone set-point temperature, and minimum zone supplyair mass flow. System characteristics include supply air temperature set point; entering water temperature of reheat, preheat, and cooling coils; minimum mass flow of outdoor air; and economizer temperature/enthalpy set point for minimum airflow. The algorithm for performing calculations for this VAV system is shown in Figure 12. The algorithm directs sequential calculations of system performance. Calculations proceed from the zones along the return air path to the cooling coil inlet and back through the supply air path to the cooling coil discharge. An alternative finite-state machine (FSM) approach is found in Chapter 42 of the 2015 ASHRAE Handbook—HVAC Applications. Moving back along the supply air path, the fan entering air temperature is calculated by setting fan outlet air temperature to the
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Fig. 11 Schematic of Variable-Air-Volume System with Reheat
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cooling coil performance on the mixed air humidity ratio. The latent load defines the difference between zone humidity and supply air humidity. However, the humidity ratio of supply air depends on the humidity ratio entering the coil, which in turn depends on that of the return air. This calculation must be performed either by solving simultaneous equations or, as in this case, iteration. Assuming a trial value for the humidity ratio at the cooling coil discharge (e.g., 13°C, 90% rh), the humidity ratio at all other points throughout the system can be calculated. With known cooling coil inlet air conditions and a design discharge air temperature, the inverted cooling coil subroutine iterates on the coil fluid mass flow to converge on the discharge air temperature with the discharge air humidity ratio as an output. The cooling coil discharge air humidity ratio is then compared to the previous discharge humidity ratio. Iteration continues through the loop several times until the values of the cooling coil discharge air humidity ratio stabilize within a specified tolerance. This basic algorithm for simulation of a VAV system might be used in conjunction with a heat balance type of load calculation. For a weighting factor approach, it would have to be modified to allow zone temperatures to vary and consequently zone loads to be readjusted. It should also be enhanced to allow possible limits on reheat temperature and/or cooling coil limits, zone humidity limits, outdoor air control (economizers), and/or heat-recovery devices, zone exhaust, return air fan, heat gain in the return air path because of lights, the presence of baseboard heaters, and more realistic control profiles. Most current building energy programs incorporate these and other features as user options, as well as algorithms for other types of systems.
5. LOW-ENERGY SYSTEM MODELING 5.1 Fig. 12 Algorithm for Calculating Performance of VAV with System Reheat system design supply air temperature. The known fan inlet air temperature is then used as both the cooling coil and preheat coil discharge air temperature set point. Moving along the return air path, the cooling coil entering air temperature can be determined by sequentially moving through the economizer cycle and preheat coil. Unlike temperature, the humidity ratio at any point in a system cannot be explicitly determined because of the dependence of
NATURAL AND HYBRID VENTILATION
Natural ventilation refers to the introduction of outdoor air into a building via intentionally provided openings (e.g., windows and trickle ventilators) whereby airflow is driven by naturally occurring forces such as wind and buoyancy (stack effect) to provide fresh air and ventilation cooling (free cooling). The movement of air resulting from natural ventilation can also provide direct cooling [see Aynsley et al. (1977) and Olgyay (1973)]. Hybrid or mixed-mode ventilation refers to the use of mechanical means to supplement or enhance natural ventilation when the driving forces of wind and/or buoyancy are unable to meet the required ventilation rate or thermal comfort targets. Hybrid ventilation can be loosely divided into three categories:
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Energy Estimating and Modeling Methods fully parallel natural and mechanical ventilation, fan-assisted natural ventilation, and stack- and wind-assisted mechanical ventilation (Hybrid Ventilation Centre 2002). Regardless of which system is implemented, control of the system components and interaction between ventilation modes is critical to system effectiveness. The design of natural ventilation systems requires careful consideration of the climate in which the building is located. Not only must prevailing outdoor temperatures, wind speed, and direction be accounted for to gage the forces available to drive natural ventilation, but outdoor temperature and humidity can affect the level of thermal comfort attainable within the building. However, even in climates where external air temperatures are relatively high, it may be possible to use natural ventilation to deliver acceptable thermal comfort. This can be done in combination with other passive design measures such as exposed thermal mass, solar shading, and nighttime ventilation cooling, or as part of a hybrid ventilation solution (Lomas et al. 2007). Also, environmental pollutants can affect the level of indoor air quality (IAQ) that is achievable without the benefit of filtering the outdoor air as it enters the building. Simulation of natural and hybrid ventilation systems should capture as many of these aspects as possible. At the very least, consideration should be given to modeling and predicting the likely performance of the ventilation strategy in terms of ventilation rates and indoor environmental quality, including thermal comfort of the building occupants. Chapters 13 and 16, respectively, discuss modeling and ventilation (including natural ventilation) in more detail, and CIBSE (2005) provides a guide to designing natural ventilation systems in nondomestic buildings.
Natural Ventilation Tools for modeling natural ventilation systems fall into three main categories, discussed in the following sections. Although these modeling tools can be applied individually, simulation and design are better served by using the capabilities of multiple types of tools to provide a more comprehensive analysis. Simplified Models. Simplified models can be used at the initial design stages for solving algebraic (nondifferential) equations to determine climate suitability, ventilation flow rates, and zoneaverage temperatures. This can be valuable information, especially at the concept design stage, when architects require approximate information about the feasibility of implementing natural ventilation and determining size and position of ventilation openings. The equations typically used in these models can be found in Chapter 16 and in CIBSE (2005). However, there are many other examples of simplified techniques used to predict the behavior of specific aspects of natural ventilation. For example, Linden et al. (1990) developed analytical models for predicting interface heights in stratified regimes, and Chenvidyakarn and Woods (2005) used analytical methods to predict solution multiplicity in large-volume spaces. Climate-suitability analysis methods and tools are available for the predesign phase. These relatively simple models provide a firstorder estimate of the potential effectiveness of natural ventilation to provide direct ventilation cooling, nighttime cooling, and thermal comfort, based on climate data representative of the building location. Axley and Emmerich (2002) present a climate suitability design method based on a single-zone, steady-state energy balance, and thermal comfort criteria. Lomas et al. (2007) present an analysis of ambient temperature and moisture content with respect to thermal comfort criteria plotted on a psychrometric chart. One common form of analytical model involves first solving equations to determine the driving pressures caused by wind and buoyancy forces. These driving pressures are calculated with reference to the neutral pressure level (CIBSE 2005) and are used at proposed opening locations as input to mathematical representations of the airflow through the openings (e.g., the orifice flow equation).
19.25 Given the design airflow rate, these equations can then be used to calculate effective opening areas. Axley (2001) presents an opening sizing method that was subsequently formulated as a software tool (Emmerich et al. 2011; Dols et al. 2012). These methods and tools can also account for multiple openings and user-specified airflow patterns through a building (e.g., into the building via a trickle ventilator, through an air transfer grill, then out through a stack). As outlined in CIBSE (2005), driving pressures should be calculated for representative design scenarios when wind speeds and internal/ external temperature differences are relatively low, because these will determine the maximum opening sizes required. Illustrations of how some of these simplified models can be used are given in Axley (2001), CIBSE (2005), and Emmerich et al. (2012). Network Airflow Models. Network airflow models, also referred to as multizone models, are discussed in Chapter 13. They are more complex than the simplified models, because they solve the simultaneous (interdependent) airflows and pressure differences between multiple, interconnected building zones and the outdoors and are typically applied to the building in its entirety. In particular, using airflow network characteristics (e.g., airflow resistances) and climate data (along with assumed temperatures within the building), network airflow models calculate combined buoyancy- and wind-driven pressures and the rate and direction of airflows for all openings in the model. Some models can also be used to perform interzone contaminant mass transport calculations that can be useful in analyzing IAQ (Axley 2007) and simulating contaminant-based ventilation rate control schemes (e.g., CO2-based demand controlled ventilation) (Dols et al. 2016a). Network airflow models can be coupled with dynamic thermal simulation programs to calculate the internal zone temperatures used in the network airflow model (Clarke 2001; Dols et al. 2016a, 2016b; U.S. Department of Energy 1996-2016). A distinct challenge in the use of network airflow models is selecting mathematical representations and parameters that best capture the airflow characteristics of the ventilation components. For example, airflow resistances can be described using a discharge coefficient (or pressure loss factor) and the effective opening area. However, specifying the component characteristics requires a clear understanding of how the model interprets these values. Details of the terminology used can be found in Jones et al. (2016). Computational Fluid Dynamics. Computational fluid dynamics (CFD) is also discussed in Chapter 13. CFD models represent the most complex class of airflow models and can be valuable for simulating natural ventilation, including the effects of wind on the building envelope pressures and internal airflow patterns that develop as a result of various boundary conditions and representative portions of the building interior. CFD is not typically applicable to simulating entire buildings over long-term, transient time frames. However, there are many strategies available for using CFD in the design and analysis of natural ventilation systems; for example, Malkawi et al. (2016) describes using both CFD and a combined network airflow and energy calculation method. The main feature that sets CFD models apart from numerical and network airflow models is the use of a fine mesh comprising many thousands of cells throughout the computational domain. For each cell, nonlinear partial differential equations are solved to predict the air speed, air temperature, turbulence, and pressure in each cell. This provides a very detailed representation of the spatial distribution of the airflow and has the advantage of highlighting drafts, warm zones, and thermal stratification. Boundary conditions are required in CFD models to drive the flow. They include temperature or heat flux values at solid surfaces and conditions of air speed and turbulence at openings. Surface temperatures are often obtained from the results of a dynamic thermal network simulation model. This approach has the advantage of including the effects of radiation without solving a separate radiation model in
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the CFD simulation. Boundary conditions at natural ventilation openings must be specified with care, because neither the flow rate nor the internal temperature are known a priori. This can be overcome by using the orifice flow equation, which simply provides a relationship between the flow rate and pressure difference across openings [e.g., Durrani et al. (2015)]. Another important challenge to be aware of when using CFD is the need for an iterative approach to solve the governing, nonlinear equations. It is therefore important to appreciate that simulation run times can be long (i.e., several hours), and that techniques for controlling convergence are often needed, especially in buoyancy-driven natural ventilation flows where driving pressures are small. Often, this control can be achieved by using a transient simulation or false time-stepping [e.g., Kaye et al. (2009)].
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Hybrid Ventilation When design tools have predicted that natural ventilation is unlikely to be sufficient for providing adequate IAQ and thermal comfort and a hybrid strategy has been proposed, it is useful to carry out some modeling of the system to provide confidence and identify likely energy implications. This can be done by enhancing the techniques described previously for modeling natural ventilation with network airflow and CFD models. Ventilation and energy performance of buildings with natural or mixed-mode ventilation systems can be simulated using wholebuilding network airflow models, including those that may be combined with or directly incorporate a whole-building thermal analysis model (e.g., EnergyPlus) (Dols et al. 2015, 2016b). The energy module sends information about the building, ambient weather conditions, and zone temperatures to the network airflow module. This information is used to calculate infiltration, ventilation, and interzone airflow by the airflow network module; the results will be sent back to the energy module and used in the next time step’s heat balance. In addition to the airflow characteristics through openings such as windows and transfers grilles connecting spaces, models of hybrid systems must also include fan airflow models (e.g., design flow rates and fan performance curves). One of the primary challenges in modeling hybrid ventilation systems is simulating the controls systems that are inherently associated with these systems. Hybrid ventilation schemes may require multiple modes of operation depending on variations in climate, including seasonal, daily, and short-term local variations. These modes of operation can vary in complexity because of the number and type of system components involved. In buildings with automatically controlled ventilation openings or fans, the control system may require parameters such as wind speed and direction, among others, to open, close, or adjust openings; turn fans on/off; or adjust flow rates. This range of flexibility varies among available models. However, it is likely that users will need to input sensors, actuators, and control logic per the input methods of the tool being used, and the model should predict system performance accordingly (e.g., variations in fan airflow, ventilation opening status). Some modeling tools or combinations of tools enable transient simulations to be performed with user-defined control algorithms (e.g., DOE 2015; Dols et al. 2016a; Wetter 2011), and these may be used for performing annual simulations or at least over periods of time that address the various modes of operation.
5.2
supplemental electric lighting is reduced when daylighting is used. In addition, energy simulation programs calculate the thermal heat loss and gain associated with the fenestration for daylighting. Electric lighting level is determined by fraction of lighting area, lighting type, daylight illuminance level, and targeted illuminance. Additional details regarding daylighting and electric lighting control strategies are described in Chapter 15. Historically, daylighting analysis programs have used three components to calculate the amount of daylight coming through a window or skylight: sky component (SC), external reflected component (ERC), and internal reflected component (IRC). Commonly used methods for estimating the IRC include the split-flux, radiosity, and ray-tracing methods. In the 1950s, the split-flux method was used to estimate the IRC using empirical formulas (Hopkinson et al. 1954). In this method, Hopkinson et al. assumed the interior surfaces of a room were a connected spherical shape and were perfectly diffused with no inner obstacles, which worked best in a room shaped as a cube without internal partitions (Winkelmann and Selkowitz 1985). Tregenza (1989) presented a modification to the split-flux method to account for large external obstacles such as overhangs. Figure 13 shows the concept of the split-flux method [ fs = window factor for the light incident on the window from sky, Rfw = average reflectance of the floor and those parts of the walls below the plane of the mid-height of the window (excluding the window-wall), fg = window factor for the light incident on the window from ground, and Rcw = average reflectance of the ceiling and those parts of the walls above the plane of the midheight of the window (excluding the window-wall)] (Oh and Haberl 2016c). The split-flux method is widely used in whole-building energy simulation programs such as DOE-2.1E (Winkelmann et al. 1993), DOE-2.2/eQUEST (LBNL 1998), and EnergyPlus (U.S Department of Energy 1996-2016). The radiosity method more accurately analyzes various geometries and reflectances than the split-flux method but less accurately than the ray-tracing method. The radiosity method provides an accurate method to analyze a zone where object-to-object reflection exists between diffuse surfaces in a zone. This procedure uses the energy balance concept for analyzing radiative heat transfer (i.e., radiant-flux transfer), which has been widely used by thermal engineers (Goral et al. 1984; O’Brien 1955). The radiant-flux transfer analysis between surfaces applies the light-flux transfer analysis in an enclosure using a lumped parameter network (O’Brien 1955; Oppenheim 1954). The radiosity method uses the electrical network approach to account for the initial and interreflected luminous emittances (i.e., the light-flux transfer). The surface reflectances (i.e.,
DAYLIGHTING
Daylighting in buildings is used to reduce the use of electric lighting systems. A proper daylighting design provides improved illumination for occupants and reduces a building’s energy use. A building’s orientation, window size and properties, and shading (e.g., overhangs and fins) affect daylight illuminance at specific points in a space. Whole-building energy simulation programs, such as EnergyPlus and DOE-2.2/eQUEST, can keep track of how much
Fig. 13 Split-Flux Method (Oh and Haberl 2016c)
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Energy Estimating and Modeling Methods
Fig. 14 Forward Ray-Tracing Method (Oh and Haberl 2016c)
19.27 towers are examples of passive solar strategies. Detailed simulation programs such as TRNSYS (Duffie and Beckman 2013) and SUNREL (Deru et al. 2002) can be used to analyze passive solar energy applications and building loads. TRNSYS and SUNREL use a thermal network approach (Paschkis 1942) to analyze the time-varying solar radiation and heat transfer in a building. In addition, simplified methods such as the solar load ratio (SLR) method (Balcomb and Hedstrom 1976) and the unutilizability method (Monsen and Klein 1980; Monsen et al. 1981) can be used to analyze passive solar systems. Computer simulations can be used to calculate the timedependent, short-term, and long-term performance of solar energy systems in detail. Simplified methods that require fewer calculations than hourly solar simulations (Klein 1993) can be used to estimate the long-term performance of solar energy systems. Passive solar design analysis is useful for engineers to select and size solar systems when input data and on-site solar irradiation data may not be available (Evans et al. 1982). Additional details can be found in Oh and Haberl (2016b).
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6. DATA-DRIVEN MODELING
Fig. 15
Backward Ray-Tracing Method (Oh and Haberl 2016c)
the radiant properties) and the view factors (i.e., the relative geometry properties) of the luminous emittances are defined as the resistances in the network (O’Brien 1955). The method is used in Lumen Micro and in one of the daylighting models of EnergyPlus v.1.x and higher. Finally, ray tracing is the most advanced and accurate method for analyzing the interreflections between both diffuse and specular surfaces in complex enclosures (Baker et al. 1993; Ward and Rubinstein 1988; Ward et al. 1988). The method follows the light rays in an enclosure and takes into account all the characteristics of various surface geometries and surface reflectances. The method was originally developed to create high-quality computer graphics in complex scenes (Baker et al. 1993). Ray tracing is typically used to analyze daylight illuminance itself, instead of being combined with whole-building energy simulation programs to account for electric lighting. The ray-tracing method is used in Radiance (Ward 1994) and DIVA (Jakubiec and Reinhart 2011). Additional details can be found in Oh and Haberl (2016c). The forward ray-tracing method traces the rays of light generated from a source of light to the eye of the viewer (Kuchkuda 1988). Figure 14 shows the concept of the forward ray-tracing method (Grantham 2008; Oh and Haberl 2016c). The backward ray-tracing method traces a ray of each point (i.e., each pixel) backwards from the viewer through the image plane to the object (Arvo 1986; Kuchkuda 1988; Ward and Rubinstein 1988; Ward et al. 1988). Figure 15 shows this concept (Grantham 2008; Oh and Haberl 2016c).
5.3
PASSIVE HEATING
Passive buildings use solar heating directly (i.e., without pumps, blowers, etc.) and sometimes include natural passive cooling. Solar direct gain, sunspaces, Trombe walls, and passive downdraft cool
The objective of data-driven (inverse) modeling is to determine a mathematical description of the system and to estimate system parameters when the input and output variables are known and measured. The data-driven approach is relevant only when the system has already been built and actual performance data are available for model development, calibration, and/or identification. There are several important uses of data-driven models. They can be used to verify savings for implemented energy conservation measures, where the data-driven model provides an estimate of baseline energy consumption for comparison to post-retrofit energy consumption. Data-driven models can be used with energy management and control systems to predict energy use (Kreider and Haberl 1994). Hourly or daily comparisons of measured versus predicted energy use can be used to determine whether systems are being left on unnecessarily or are in need of maintenance. Combinations of predicted energy use and a knowledge-based system can indicate above-normal energy use and diagnose the possible cause of the malfunction if sufficient historical information has been previously gathered (Haberl and Claridge 1987). Hourly systems that use artificial neural networks have also been constructed (Kreider and Wang 1991).
6.1
CATEGORIES OF DATA-DRIVEN METHODS
Data-driven methods for energy-use estimation in buildings and related HVAC&R equipment can be classified into two broad categories. These approaches differ widely in data requirements, time and effort needed to develop the associated models, user skill requirements, and sophistication and reliability provided.
Empirical or “Black-Box” Approach With an empirical approach, a simple or multivariate regression model is identified between measured energy use and the various influential parameters (e.g., climatic variables, building occupancy). The form of the regression models can be either purely statistical or loosely based on some basic engineering formulation of energy use in the building. In any case, the identified model coefficients are such that no (or very little) physical meaning can be assigned to them. This approach can be used with any time scale (monthly, daily, hourly or subhourly) if appropriate data are available. Single-variate, multivariate, change point, Fourier series, and artificial neural network (ANN) models fall under this category. Least-squares regression is the most common approach to model identification, although more sophisticated regression techniques
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such as maximum likelihood and two-stage regression schemes can also be used. Statistical regression models are usually adequate for estimating savings in demand-side management (DSM) programs for conventional energy conservation measures (lighting retrofits, air handler retrofits such as CV to VAV retrofits) and for baseline model development in energy conservation measurement and verification (M&V) projects (Claridge 1998; Dhar 1995; Dhar et al. 1998, 1999a, 1999b; Fels 1986; Haberl and Culp 2012; Haberl et al. 1998; Katipamula et al. 1998; Kissock et al. 1998; Krarti et al. 1998; Kreider and Wang 1991; MacDonald and Wasserman 1989; Miller and Seem 1991; Reddy et al. 1997; Ruch and Claridge 1991). Statistical models may also be appropriate for modeling equipment such as pumps and fans, and even more elaborate equipment such as chillers and boilers, if the necessary performance data are available (Braun 1992; Chen et al. 2005; Englander and Norford 1992; Lorenzetti and Norford 1993; Phelan et al. 1996). Although this approach allows detection or flagging of equipment or system faults, it is usually of limited value for diagnosis and online control.
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Gray-Box Approach The gray-box approach first formulates a physical model to represent the structure or physical configuration of the building or HVAC&R equipment or system, and then identifies important parameters representative of certain key and aggregated physical parameters and characteristics by statistical analysis (Rabl and Riahle 1992). This approach requires a high level of user expertise both in setting up the appropriate modeling equations and in estimating these parameters. Often an intrusive experimental protocol is necessary for proper parameter estimation. This approach has great potential, especially for fault detection and diagnosis (FDD) and online control, but its applicability to whole-building energy use is limited. Examples of parameter estimation studies applied to building energy use are Andersen and Brandemuehl (1992), Braun (1990), Gordon and Ng (1995), Guyon and Palomo (1999a), Hammarsten (1984), Rabl (1988), Reddy (1989), Reddy et al. (1999), Sonderegger (1977), and Subbarao (1988).
6.2
TYPES OF DATA-DRIVEN MODELS
Steady-state models do not consider effects such as thermal mass or capacitance that cause short-term temperature transients. Generally, these models are appropriate for monthly, weekly, or daily data and are often used for baseline model development. Dynamic models capture effects such as building warm-up or cooldown periods and peak loads, and are appropriate for building load control, FDD, and equipment control. A simple criterion to determine whether a model is steady-state or dynamic is to look for the presence of time-lagged variables, either in the response or regressor variables. Steady-state models do not contain time-lagged variables. Table 4
Steady-State Models Several types of steady-state models are used for both building and equipment energy use: single-variate, multivariate, polynomial, and physical. Single-Variate Models. Single-variate models (i.e., models with one regressor variable only) are perhaps the most widely used. They formulate energy use in a building as a function of one driving force that affects building energy use. An important aspect in identifying statistical models of baseline energy use is the choice of the functional form and the independent (or regressor) variables. Extensive studies (Fels 1986; Katipamula et al. 1994; Kissock et al. 1993; Reddy et al. 1997) have indicated that the outdoor dry-bulb temperature is the most important regressor variable for typical buildings, especially at monthly time scales but also at daily time scales. The simplest steady-state data-driven model is one developed by regressing monthly utility consumption data against average billing-period temperatures. The model must identify the balancepoint temperatures (or change points) at which energy use switches from weather-dependent to weather-independent behavior. In its simplest form, the 18.3°C degree-day model is a change-point model that has a fixed change point at 18.3°C. Other examples include three- and five-parameter Princeton scorekeeping methods (PRISM) based on the variable-base degree-day concept (Fels 1986). An allied modeling approach for commercial buildings is the four-parameter (4-P) model developed by Ruch and Claridge (1991), which is based on the monthly mean temperature (and not degree-days). Table 4 shows some commonly used model functional forms. The three parameters are a weather-independent base-level use, a change point, and a temperature-dependent energy use, characterized as a slope of a line that is determined by regression. The four parameters include a change point, a slope above the change point, a slope below the change point, and the energy use associated with the change point. A data-driven bin method has also been proposed to handle more than four change points (Thamilseran and Haberl 1995). Figure 16 shows several types of steady-state, single-variate data-driven models. Figure 16A shows a simple one-parameter, or constant, model, and Table 4 gives the equivalent notation for calculating the constant energy use using this model. Figure 16B shows a steady-state two-parameter (2-P) model where b0 is the yaxis intercept and b1 is the slope of the regression line for positive values of x, where x represents the ambient air temperature. The 2P model represents cases when either heating or cooling is always required. Figure 16C shows a three-parameter change-point model, typical of natural gas energy use in a single-family residence that uses gas for space heating and domestic water heating. In the notation of Table 4 for the three-parameter model, b0 represents the baseline energy use and b1 is the slope of the regression line for values of ambient temperature less than the change point b2. In this type of notation, the superscripted plus sign indicates that only positive values of the parenthetical expression are considered. Figure 16D
Single-Variate Models Applied to Utility Billing Data
Model Type
Independent Variable(s)
Form
Examples
One-parameter or constant (1-P) Two-parameter (2-P) Three-parameter (3-P)
None Temperature Degree-days/ Temperature
Non-weather-sensitive demand
Four-parameter change point (4-P)
Temperature
Five-parameter (5-P)
Degree-days/ Monthly mean temperature
E = b0 E = b0 +b1(T ) E = b0 + b1(DDBT) E = b0 + b1(b2 T )+ E = b0 + b1(T b2)+ E = b0 + b1(b3 T )+ b2 (T b3)+ E = b0 – b1(b3 T )+ + b2 (T b3)+ E = b0 b1(DDTH) + b2 (DDTC) E = b0 + b1(b3 T )+ + b (T – b4)+
Note: DD denotes degree-days and T is monthly mean daily outdoor dry-bulb temperature.
Seasonal weather-sensitive use (fuel in winter, electricity in summer for cooling) Energy use in commercial buildings Heating and cooling supplied by same meter
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Energy Estimating and Modeling Methods
Fig. 16 Steady-State, Single-Variate Models for Modeling Energy Use in Residential and Commercial Buildings shows a three-parameter model for cooling energy use, and Table 4 provides the appropriate analytic expression. Figures 16E and 16F show four-parameter models for heating and cooling, respectively. The appropriate expressions for calculating the heating and cooling energy consumption are found in Table 4: b0 represents the baseline energy exactly at the change point b3, and b1 and b2 are the lower and upper region regression slopes for ambient air temperature below and above the change point b3. Figure 16G shows a 5-P model (Fels 1986), which is useful for modeling buildings that are electrically heated and cooled. The 5-P model has two change points and a base level consumption value. The advantage of these steady-state data-driven models is that their use can be easily automated and applied to large numbers of buildings where monthly utility billing data and average daily temperatures for the billing period are available. Steady-state singlevariate data-driven models have also been applied with success to
daily data (Kissock et al. 1998). In such a case, the variable-base degree-day method and monthly mean temperature models described previously for utility billing data analysis become identical in their functional form. Single-variate models can also be applied to daily data to compensate for differences such as weekday and weekend use by separating the data accordingly and identifying models for each period separately. Disadvantages of steady-state single-variate data-driven models include insensitivity to dynamic effects (e.g., thermal mass) and to variables other than temperature (e.g., humidity and solar gain), and inappropriateness for some buildings (e.g., buildings with strong on/off schedule-dependent loads or buildings with multiple change points). Moreover, a single-variable, 3-P model such as the PRISM model (Fels 1986) has a physical basis only when energy use above a base level is linearly proportional to degree-days. This is a good approximation in the case of heating energy use in residential
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19.30 buildings where heating load never exceeds the heating system’s capacity. However, commercial buildings generally have higher internal heat generation with simultaneous heating and cooling energy use and are strongly influenced by HVAC system type and control strategy. This makes energy use in commercial buildings less strongly influenced by outdoor air temperature alone. Therefore, it is not surprising that blind use of single-variate models has had mixed success at modeling energy use in commercial buildings (MacDonald and Wasserman 1989). Change-point regression models work best with heating data from buildings with systems that have few or no part-load nonlinearities (i.e., systems that become less efficient as they begin to cycle on/off with part loads). In general, single-variate change-point regression models do not predict cooling loads as well because outdoor humidity has a large influence on latent loads on the cooling coil. Other factors that decrease the accuracy of change-point models include solar effects, thermal lags, and on/off HVAC schedules. A change point model based on a linear combination of temperature and specific humidity, and in which direct and diffuse irradiance were included as separate explanatory variables, was found to work well for aggregates of buildings (Ali et al. 2011). Four-parameter models are a better statistical fit than threeparameter models in buildings with continuous, year-round cooling or heating (e.g., grocery stores and office buildings with high internal loads). However, every model should be checked to ensure that the regression does not falsely indicate an unreasonable relationship. Paulus et al. (2015) provided an algorithm aiding in the selection of change-point linear regression models. A major advantage of using a steady-state data-driven model to evaluate the effectiveness of energy conservation retrofits is its ability to factor out year-to-year weather variations by using a normalized annual consumption (NAC) (Fels 1986). Basically, annual energy conservation savings can be calculated by comparing the difference obtained by multiplying the pre- and post-retrofit parameters by the weather conditions for the average year. Typically, 10 to 20 years of average daily weather data from a nearby weather service site are used to calculate 365 days of average weather conditions, which are then used to calculate the average pre- and post-retrofit conditions. Utilities and government agencies have found it advantageous to prescreen many buildings against test regression models. These data-driven models can be used to develop comparative figures of merit for buildings in a similar standard industrial code (SIC) classification. A minimum goodness of fit is usually established that determines whether the monthly utility billing data are well fitted by the one-, two-, three-, four-, or five-parameter model being tested. Comparative figures of merit can then be determined by dividing the parameters by the conditioned floor area to yield average daily energy use per unit area of conditioned space. For example, an areanormalized comparison of base-level parameters across residential buildings would be used to analyze weather-independent energy use. This information can be used by energy auditors to focus their efforts on those systems needing assistance (Haberl and Komor 1990a, 1990b). Multivariate Models. Three types of steady-state, multivariate models have been reported: • Standard multiple-linear or change-point regression models, where the set of data observations is treated without retaining the time-series nature of the data (Katipamula et al. 1998). • Fourier series models that retain the time-series nature of building energy use data and capture the diurnal and seasonal cycles according to which buildings are operated (Dhar 1995; Dhar et al. 1998, 1999a, 1999b; Seem and Braun 1991). • Gaussian process (GP) models use a nonparametric, nonlinear regression method that produces estimates of the prediction
2017 ASHRAE Handbook—Fundamentals (SI) uncertainty along with predictions of energy use (Burkhart et al. 2014; Heo and Zavala 2012; Heo et al. 2013). These models are a logical extension of single-variate models, provided that the choice of variables to be included and their functional forms are based on the engineering principles under which HVAC and other building systems operate. The goal of modeling energy use by the multivariate approach is to characterize building energy use with a few readily available and reliable input variables. These input variables should be selected with care. The model should contain variables not affected by the retrofit and thus unlikely to change (e.g., climatic variables) from pre-retrofit to post-retrofit periods. Other less obvious variables, such as changes in operating hours, base load, and occupancy levels, should be included in the model if these are not energy conservation measures (ECMs) but variables that may change during the post-retrofit period. Environmental variables that meet these criteria for modeling heating and cooling energy use include outdoor air dry-bulb temperature, solar radiation, and outdoor specific humidity. Some of these variables are difficult to estimate or measure in an actual building and hence are not good candidates for regressor variables. Further, some of the variables change little over time. Although the effect of these variables on energy use may be important, a datadriven model will implicitly lump their effect into the parameter that represents constant load. In commercial buildings, internally generated loads, such as the heat given off by people, lights, and electrical equipment, also affect heating and cooling energy use. These internal loads are difficult to measure in their entirety. However, monitored electricity used by internal lights and equipment may be a good surrogate for total internal sensible loads (Reddy et al. 1999). For example, when the building is fully occupied, it is also likely to be experiencing high internal electric loads, and vice versa. The effect of environmental variables is important for buildings such as offices but may be less so for buildings with loads dominated by internal heat gain (e.g., data centers) and buildings with loads dominated by occupants (e.g. assembly buildings). Differences in HVAC system behavior during occupied and unoccupied periods can be modeled by a dummy or indicator variable (Draper and Smith 1981). For some office buildings, there seems to be little need to include a dummy variable, but its inclusion in the general functional form adds flexibility. Air-side heating and cooling use in various HVAC system types has been addressed by Reddy et al. (1995) and subsequently applied to monitored data in commercial buildings (Katipamula et al. 1994, 1998). Because quadratic and cross-product terms of engineering equations are not usually picked up by multivariate models, strictly linear energy use models are often the only option. In addition to outdoor temperature To, internal electric equipment and lighting load Eint , solar loads qsol , and latent effects via the outdoor dew-point temperature Tdp are candidate regressor variables. In commercial buildings, a major portion of the latent load derives from fresh air ventilation. However, this load appears only when the outdoor air dew-point temperature exceeds the cooling coil temperature. Hence, the term (Tdp – Ts)+ (where the + sign indicates that the term is to be set to zero if negative, and Ts is the mean surface temperature of the cooling coil, typically about 11 to 13C) is a more realistic descriptor of the latent loads than is Tdp alone. Using (Tdp – Ts)+ as a regressor in the model is a simplification that seems to yield good accuracy. Therefore, a multivariate linear regression model with an engineering basis has the following structure: –
+
Q bldg = b 0 + b 1 T o – b 3 + b 2 T o – b 3 + b 4 T dp – b 6 +
+ b 5 T dp – b 6 + b 7 q sol + b 8 E int
–
(19)
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Based on the preceding discussion, b4 = 0 because outdoor air does not introduce a latent cooling load when Tdp is less than the cooling coil surface temperature. Introducing indicator variable terminology (Draper and Smith 1981), Equation (19) becomes +
Q bldg = a + bT o + cI + dIT o + eT dp + f q sol + gE int
(20)
where the indicator variable I is introduced to handle the change in slope of the energy use due to To. The variable I is set equal to 1 for To values to the right of the change point (i.e., for high To range) and set equal to 0 for low To values. As with the single-variate segmented models (i.e., 3-P and 4-P models), a search method is used to determine the change point that minimizes the total sum of squares of residuals (Fels 1986; Kissock et al. 1993). Katipamula et al. (1994) found that Equation (20), appropriate for VAV systems, could be simplified for constant-volume HVAC systems: +
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Q bldg = a + bT o + eT dp + f q sol + gE int
(21)
Note that instead of using (Tdp – Ts)+, the absolute humidity potential (W0 – Ws)+ could also be used, where W0 is the outdoor absolute humidity, and Ws is the absolute humidity level at the dew point of the cooling coil (typically about 0.009 kg/kg). A final + aspect to keep in mind is that the term T dp should be omitted from the regressor variable set when regressing heating energy use, because there are no latent loads on a heating coil. These multivariate models are very accurate for daily time scales and slightly less so for hourly time scales. The inaccuracy at hourly time scales is because changes in the way the building is operated during the day and the night lead to different relative effects of the various regressors on energy use, which cannot be accurately modeled by one single hourly model. Breaking up energy use data into hourly bins corresponding to each hour of the day and then identifying 24 individual hourly models leads to appreciably greater accuracy (Katipamula et al. 1994). Another accurate method uses a piecewise linear regression with respect to temperature and a time-of-week indicator variable to account for changes in energy use throughout each day of the week. This model (Matthieu 2011) may be applied to daily, hourly, or subhourly data. Gaussian process (GP) models are nonparametric regression models, meaning the relationship between input and output variables is not strictly defined at the outset of the regression. In GP modeling, the measured output data are samples of a multivariate Gaussian distribution with the distribution parameters being the input variables. The distribution is defined by matching the covariance of the distribution to the covariance between the measured input and output data. For any new set of input variables, a Gaussian distribution is defined and the mean of the distribution is interpreted to be the predicted output and the variance of the distribution the uncertainty in the prediction. Thus, GP modeling not only produces an output value given a set of multivariate inputs, it also produces an estimate of the uncertainty of the prediction. Heo and Zavala (2012) used GP models to predict the hourly energy use of a chilled-water system using ambient temperature, occupancy, relative humidity, and chiller supply temperature and showed much better performance than standard multivariate regression models. Hybrid Inverse Change Point Model. ASHRAE RP-1404 (Abushakra et al. 2014) was conducted primarily to develop analysis methodologies by which less than a whole year of field monitoring can be used for whole-building energy use prediction while satisfying current accuracy guidelines. The methodologies were developed to benefit ongoing efforts to spur diffusion of high-performance buildings by actual monitoring and to benefit energy service companies (ESCOs) and energy professionals who need a more
cost-effective and acceptable alternative to year-long energy monitoring. Applications for the hybrid inverse change point model include detailed audits, green-building performance verification, and verification of post-retrofit claims using pre/post monitored data. The modeling method is called hybrid because two regressions are performed at different time scales. The weather-dependent portion uses a year of monthly day-normalized utility bills, and the weather-independent portion uses either daily or hourly data covering a shorter time span, possibly as short as two weeks. The research revealed that (1) two weeks of hourly monitored data can be adequate for producing accurate long-term whole-building energy predictions, if the monitoring is taken at a correct time, and (2) adding more data to the regression model (extending the length of the shortterm monitoring period) can, counterintuitively, make the predictive model poorer. The analysis was completed in detail at two different time scales: daily and hourly. The analysis with both hourly and daily data revealed that the proper monitoring period is highly dependent on the time at which the data are taken, with the best time to collect data being during the spring and fall, or periods where the average temperature is close to the annual average temperature and the temperature has a wide range. The following is an example of the modeling procedure that results in an hourly energy regression model. An example of the method applied to daily data can be found in Singh et al. (2013). The first regression uses one of the steady-state change point models shown in Figure 16 and Table 4, possibly with an additional weather-related independent variable, such as humidity. For example, the form could be a 3-P cooling shape (using the nomenclature of Table 4) such as Edaily = b0 + b1(Tdaily average – b2)+
(22)
where Edaily is day-normalized energy consumption in units of energy consumption per day. For predicting cooling thermal loads, such as chilled-water use, investigate adding the outdoor humidity variable as a regressor in the model to account for latent loads. Equation (22) would become a multivariable linear regression of the form Edaily = b0 + b1(Tdaily average – b2)+ + b3(daily average – 0.009)+ (23) The (daily average – 0.009)+ term in Equation (23) is a surrogate for the latent load existing only when the absolute humidity ratio of air entering the cooling coil is above 0.009. The dependent variable for the second regression is the residuals resulting from using the first equation to predict the temperature dependent energy portion at the hourly time scale, calculated by dividing the temperature dependent portion of the first regression by 24 hours/day. The second regression uses gathered hourly data consisting of weather and an internal driving variable(s), such as Xinternal,hourly, often an occupancy indicator or submetered lighting and equipment load. The second regression is then performed to calculate the coefficients for the internal driving variables and constant term. Using Equation (22), the second regression has the form b1 + E hourly – ------ T hourly – b 2 = b 3 + b 4 X internal hourly 24
(24)
The final equation for prediction at the hourly time scale is a rearrangement of Equation (24). b1 + E hourly = b 3 + ------ T hourly – b 2 + b 4 X internal hourly 24
(25)
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2017 ASHRAE Handbook—Fundamentals (SI)
Polynomial Models. Historically, polynomial models have been widely used as pure statistical models to model the behavior of equipment such as pumps, fans, and chillers (Stoecker and Jones 1982), as discussed in the section on HVAC Component Modeling. The theoretical aspects of calculating pump performance are well understood and documented. Pump capacity and efficiency are calculated from measurements of pump pressure, flow rate, and pump electrical power input. Phelan et al. (1996) studied the predictive ability of linear and quadratic models for electricity consumed by pumps and water mass flow rate, and concluded that quadratic models are superior to linear models. For fans, Phelan et al. (1996) studied the predictive ability of linear and quadratic polynomial single-variate models of fan electricity consumption as a function of supply air mass flow rate, and concluded that, although quadratic models are superior in terms of predicting energy use, the linear model seems to be the better overall predictor of both energy use and demand (i.e., maximum monthly power consumed by the fan). This is a noteworthy conclusion given that a third-order polynomial is warranted analytically as well as from monitored field data presented by previous authors [e.g., Englander and Norford (1992), Lorenzetti and Norford (1993)]. Polynomial models have been used to correlate chiller (or heat pump) capacity Q evap and the electrical power consumed by the chiller (or compressor) E comp with the relevant number of influential physical parameters. For example, based on the functional form of the DOE-2 building simulation software (York and Cappiello 1982), models for part-load performance of energy equipment and plant, Ecomp, can be modeled as the following triquadratic polynomial: in
2
out
E comp = a + bQ evap + cT cond + dT evap + eQ evap +f
in 2 T cond
out 2 gT evap
+
in
out
+
in hQ evap T cond in
+
out iT evap Q evap (26)
out
+ j T cond T evap + kQ evap T cond T evap In this model, there are 11 model parameters to identify. However, because all of them are unlikely to be statistically significant, a step-wise regression to the sample data set yields the optimal set of parameters to retain in a given model. For example, Armstrong et al. (2006b) found the b, i, and k terms, using Todb for Tcond and Tewb for Tevap, to provide an adequate model of three different fixedspeed rooftop units. Other authors, such as Braun (1992) and Hydeman et al. (2002), used slightly different polynomial forms. Over a wide range of capacity and lift a full tricubic polynomial can be justified for VRF heat pump equipment (Gayeski et al. 2011). Physical Models. In contrast to polynomial models, which have no physical basis, physical models are based on fundamental thermodynamic or heat transfer considerations. These types of models are usually associated with the parameter estimation approach. Often, physical models are preferred because they generally have fewer parameters, and their mathematical formulation can be traced to actual physical principles that govern the performance of the building or equipment. Hence, model coefficients tend to be more robust, leading to sounder model predictions. Only a few studies have used steady-state physical models for parameter estimation relating to commercial building energy use [e.g., Reddy et al. (1999)]. Unlike in single-family residences, it is difficult to perform elaborately planned experiments in large buildings and obtain representative values of indoor fluctuations. For example, the generalized Gordon and Ng (GN) model (Gordon and Ng 2000) is a simple, analytical, universal model for chiller performance based on first principles of thermodynamics and linearized heat losses. The model predicts the dependent chiller coefficient of performance (COP) (the ratio of thermal cooling capacity Qch to electrical power E consumed by the chiller) with
easily measurable parameters such as the fluid (water or air) temperature entering the condenser Tcdi, fluid temperature entering the evaporator Tchi , and the thermal cooling capacity of the evaporator. The GN model is a three-parameter model in the following form: T chi T cdi – T chi 1 T chi ----------- + 1 --------- – 1 = a 1 --------- + a 2 -----------------------------Q ch T cdi Q ch COP T cdi 1 COP + 1 Q ch + a 3 -----------------------------------------T cdi
(27a)
where temperatures are in absolute units. Substituting the following, T chi x 1 = --------Q ch and
T cdi – T chi x 2 = -----------------------------T cdi Q ch
1 COP + 1 Q ch x 3 -----------------------------------------T cdi
1 T chi y = ----------- + 1 --------- – 1 COP T cdi
(27b)
the model given by Equation (27a) becomes y = a1 x1 + a2 x2 + a3 x3
(27c)
which is a three-parameter linear model with no intercept term. The parameters of the model in Equation (27c) have the following physical meaning: a1 = S = total internal entropy production in chiller a2 = Qleak = heat losses (or gains) from (or into) chiller a3 = R = total heat exchanger thermal resistance = 1/Ccd + 1/Cch, where C is effective thermal conductance Gordon and Ng (2000) point out that Qleak is typically an order of magnitude smaller than the other terms, but it is not negligible for accurate modeling, and should be retained in the model if the other two parameters identified are to be used for chiller diagnostics. The same linear model structure as Equation (27c) can be used if the fluid temperature leaving the evaporator Tcho is used instead of Tchi. However, the physical interpretation of the term a3 is modified accordingly. Reddy and Anderson (2002) and Sreedharan and Haves (2001) found that the GN and multivariate polynomial (MP) models were comparable in their predictive abilities. The GN model requires much less data if selected judiciously [even four well-chosen data points can yield accurate models, as demonstrated by Corcoran and Reddy (2003)]. Jiang and Reddy (2003) tested the GN model against more than 50 data sets covering various generic types and sizes of water-cooled chillers (single- and double-stage centrifugal chillers with inlet guide vanes and variable-speed drives, screw, scroll), and found excellent predictive ability (coefficient of variation of RMSE in the range of 2 to 5%).
Dynamic Models In contrast to steady-state data-driven models, which are used with monthly and daily data containing one or more independent variables, dynamic data-driven models are typically used with hourly or subhourly data in cases where the building’s thermal mass is significant enough to delay heat gains or losses. Dynamic models traditionally require solving a set of differential equations. Disadvantages of dynamic data-driven models include their complexity and the need for more detailed measurements to tune the model. Unlike steady-state data-driven models, dynamic datadriven models usually require a high degree of user interaction and knowledge of the building or system being modeled.
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Several residential energy studies have used dynamic data-driven models based on parameter estimation approaches, usually involving intrusive data gathering. Rabl (1988) classified the various types of dynamic data-driven models used for whole-building energy use, and identified the common underlying features of these models. There are essentially four different types of model formulations: thermal network, time series, differential equation, and modal, all of which qualify as parameter-estimation approaches. A few studies (Hammarsten 1984; Rabl 1988; Reddy 1989) evaluated these different approaches with the same data set. Several papers reported results of applying different techniques, such as thermal network and autoregressive-moving-average (ARMA) models, to residential and commercial building energy use. Examples of dynamic data-driven models for commercial buildings are found in Andersen and Brandemuehl (1992), Braun (1990), and Rabl (1988), and for apartment buildings in Armstrong et al. (2006a). Successful use of ARMA models for model-predictive control is reported by Gayeski et al. (2012). Dynamic data-driven models based on pure statistical approaches have also been reported. Two examples are machine learning (Miller and Seem 1991) and artificial neural networks (Kreider and Haberl 1994; Kreider and Wang 1991; Miller and Seem 1991).
6.3
MODEL ACCURACY AND GOODNESS OF FIT
Good data-driven models are determined by their overall goodness-of-fit and accuracy metrics, by testing each parameter for significance, and ensuring that the chosen model form and regressors explain the response of the dependent variable as completely as possible. Accuracy of model results may be represented by statistical metrics. Commonly used metrics are the normalized mean bias error (NMBE) and the coefficient of variance of the root mean square error [CV(RMSE)]. Equations for these metrics are provided in the section on Modeling Calibration. CV(RMSE) tests how well the model recreates the data and is also an indication of random error. NMBE tests how biased the model is in predicting outputs over the period the model was developed. A widely used statistic to gage a model’s goodness of fit is the coefficient of determination R2 (Draper and Smith 1981; Neter et al. 1989). A value of R2 = 1 indicates a perfect correlation between actual data and the regression equation; a value of R2 = 0 indicates no correlation. For tuning for a performance contract, a rule of thumb is that the value of R2 should not be less than 0.75. When more than one independent variable is included in the regression, the adjusted R2 should be used to assess the goodness of fit, as the unadjusted R2 increases with the addition of variables, and can be misleading. When additional independent variables are included, each should be tested for significance. The standard error of the estimate of the coefficients on each independent variable should be assessed. The smaller the standard error compared to the coefficient’s magnitude, the more reliable the coefficient estimate. To identify the significance of individual coefficients, t-statistics (or t-values) are used. These are simply the ratio of the coefficient estimate divided by the standard error of the estimate. The coefficient of each variable included in the regression has a t-statistic. For a coefficient to be statistically meaningful, the absolute value of its t-statistic must be at least 2.0. In other words, under no circumstances should a variable be included in a regression if the standard error of its coefficient estimate is greater than half the magnitude of the coefficient (even when including a variable that increases the adjusted R2). Generally, including more variables in a regression increases R2, but the significance of most individual coefficients is likely to decrease. Some independent variables may be linearly correlated. This condition, called multicollinearity, can result in large uncertainty
19.33 in the estimates of the regression coefficients (i.e., error) and can also lead to poorer model prediction accuracy compared to a model where the regressors are not linearly correlated. Residuals should be plotted against each independent variable to check that no relationship of any sort (not just linear correlation) exists. The distribution of residuals should follow a normal distribution. For time-series data, the residuals should be checked for serial correlation (ASHRAE 2010). Several authors recommend using principal component analysis (PCA) to overcome multicollinearity effects. PCA was one of the strongest analysis methods in the ASHRAE Predictor Shootout I and II contests (Haberl and Thamilseran 1996; Kreider and Haberl 1994). Analysis of multiyear monitored daily energy use in a grocery store found a clear superiority of PCA over multivariate regression models (Ruch et al. 1993), but this conclusion is unsupported for commercial building energy use in general. A more general evaluation by Reddy and Claridge (1994) of both analysis techniques using synthetic data from four different U.S. locations found that injudicious use of PCA may exacerbate rather than overcome problems associated with multicollinearity. Draper and Smith (1981) also caution against indiscriminate use of PCA.
6.4
EXAMPLES USING DATA-DRIVEN METHODS
Modeling Utility Bill Data The following example (taken from Sonderegger 1998) illustrates a utility bill analysis. Assume that values of utility bills over an entire year have been measured. To obtain the equation coefficients through regression, the utility bills must be normalized by the length of the time interval between utility bills. This is equivalent to expressing all utility bills, degree-days, and other independent variables by their daily averages. Appropriate modeling software is used in which values are assumed for heating and cooling balance points; from these, the corresponding heating and cooling degree-days for each utility bill period are determined. Repeated regression is done till the regression equation represents the best fit to the meter data. The model coefficients are then assumed to be tuned. Some programs allow direct determination of these optimal model parameters (e.g., balance point temperature) without manual tuning of the parameters by the user. Figure 17 shows how well a regression fit captures measured baseline energy use in a hospital building. Cooling degree-days are found to be a significant variable, with the best fit for a base temperature of 12.2°C. In this analysis, some individual utility bills may be unsuitable to develop a baseline and should be excluded from the regression. For example, a bill may be atypically high because of a one-time equipment malfunction that was subsequently repaired. Given the justification to exclude some data, it is often tempting to look for more reasons to exclude bills that fall far from “the line” and not question those that are close to it. For example, bills for periods containing vacations or production shutdowns may look anomalously low, but excluding them from the regression would result in a chronic overestimate of the future baseline during the same period.
Neural Network Models Figure 18 shows results for a single neural network typical of several hundred networks constructed for an academic engineering center located in central Texas. The cooling load is created by solar gains, internal gains, outdoor air sensible heat, and outdoor air humidity loads. The neural network is used to predict the pre-retrofit energy consumption for comparison with measured consumption of the retrofitted building. Six months of pre-retrofit data were available to train the network. Solid lines show the known building consumption data, and dashed lines show the neural network
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Fig. 17 Variable-Base Degree-Day Model Identification Using Electricity Utility Bills at Hospital
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(Sonderegger 1998)
These measured independent variables were able to predict chilledwater use to an RMS error of less than 4% (JCEM 1992). Choosing an optimal network’s configuration for a given problem remains an art. The number of hidden neurons and layers must be sufficient to meet the requirement of the given application. However, if too many neurons and layers are used, the network tends to memorize data rather than learning (i.e., finding the underlying patterns in the data). Further, choosing an excessively large number of hidden layers significantly increases the required training time for certain learning algorithms. Anstett and Kreider (1993), Krarti et al. (1998), Kreider and Wang (1991), and Wang and Kreider (1992) report additional case studies for commercial buildings.
6.5
Fig. 18 Neural Network Prediction of Whole-Building, Hourly Chilled-Water Consumption for Commercial Building
predictions. This figure shows that a neural network trained for one period (September 1989) can predict energy consumption well into the future (in this case, January 1990). The network used for this prediction had two hidden layers. The input layer contained eight neurons that receive eight different types of input data as listed below. The output layer consisted of one neuron that gave the output datum (chilled-water consumption). Each training fact (i.e., training data set), therefore, contained eight input data (independent variables) and one pattern datum (dependent variable). The following eight hourly input data used in each hour’s data vector were selected on physical bases (Kreider and Rabl 1994): • • • • • • • •
Hour number (0 to 2300) Ambient dry-bulb temperature Horizontal insolation Humidity ratio Wind speed Weekday/weekend binary flag (0, 1) Past hour’s chilled-water consumption Second past hour’s chilled-water consumption
MODEL SELECTION
Table 5 presents a decision diagram for selecting a forward or data-driven model where use of the model, degree of difficulty in understanding and applying the model, time scale for data used by the model, calculation time, and input variables used by the models are the criteria used to choose a particular model. More information on data-driven models can be found in the ASHRAE Inverse Modeling Toolkit (Haberl and Cho 2004; Haberl et al. 2003; Kissock et al. 2003). This toolkit contains FORTRAN 90 and executable code for performing linear and change-point linear regressions, variable-based degree-days, multilinear regression, and combined regressions. It also includes a complete test suite of data sets for testing all models.
7. MODEL CALIBRATION Calibration is the use of known data (e.g., utility bills) on the observed relationship between a dependent variable (e.g., simulation output) and an independent variable (e.g., simulation input) to make estimates of other values of the independent variable from new observations of the dependent variable. For energy simulation models, calibration typically involves observation of changes in simulation output as simulation inputs are modified, with the goal of identifying a set of inputs leading to simulation outputs that match measured building performance. The perception of validity and usefulness of any energy simulation model is largely determined by how closely the simulation output matches actual building performance, usually in terms of energy consumption. The process of calibrating a model to match actual
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Table 5 Capabilities of Different Forward and Data-Driven Modeling Methods Methods
Usea
Difficulty
Time Scaleb
Calc. Time
Variablesc
Accuracy
Simple linear regression Multiple linear regression ASHRAE bin method and data-driven bin method Change-point models ASHRAE TC 4.7 modified bin method Artificial neural networks Thermal network Fourier series analysis ARMA model Modal analysis Differential equation Computer simulation (component-based) (fixed schematic) Computer emulation
ES D, ES ES D, ES ES, DE D, ES, C D, ES, C D, ES, C D, ES, C D, ES, C D, ES, C D, ES, C, DE D, ES, DE D, C
Simple Simple Moderate Simple Moderate Complex Complex Moderate Moderate Complex Complex Very complex Very complex Very complex
D, M D, M H H, D, M H S, H S, H S, H S, H, D S, H S, H S, H H S, H
Very fast Fast Fast Fast Medium Fast Fast Medium Fast Medium Fast Slow Slow Very slow
T T, H, S, W, t T T T, S, tm T, H, S, W, t, tm T, S, tm T, H, S, W, t, tm T, H, S, W, t, tm T, H, S, W, t, tm T, H, S, W, t, tm T, H, S, W, t, tm T, H, S, W, t, tm T, H, S, W, t, tm
Low Medium Medium Medium Medium High High High High High High Medium Medium High
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Notes: a Use shown includes diagnostics (D), energy savings calculations (ES), design (DE), and control (C).
performance can be a complex and time-consuming endeavor. Identifying and isolating sources of discrepancy between the results of a model and actual data are not always possible, as described in the section on Uncertainty in Modeling. In addition, energy model calibration typically involves several input parameters that must be calibrated using a relatively limited amount of measured data; because of combinatorial complexity, calibration is an underdetermined system in which there can exist many unique (and substantially different) models that are within a tolerable error. A typical calibration requires an initial model, set of inputs to be calibrated, weather data, collection of measurements, error metric, and acceptance criteria. Once defined, a manual, semiautomatic, or automatic calibration process is invoked on these requirements. In some cases, there is a validation process afterward to assess how well the calibration process performed. Analytical, mathematical, or statistical techniques can be used for the calibration process or to assess the quality of calibration. Table 6 lists some of the common methods and techniques used for model calibration (Coakley et al. 2014). These methods have varying degrees of complexity and result in varying degrees of calibration accuracy. The quality of a calibration is often evaluated in terms of statistical indicators that quantify discrepancies between the model output and measured output. These metrics are based on time steps, and it may be the case that a model calibrated to low-resolution (e.g., monthly) whole-building data is highly inaccurate when compared to higher-resolution temporal (e.g., hourly) or spatial (e.g., level of a thermal zone) data. Among statistical indicators, the normalized mean bias error (NMBE) and the coefficient of variance of the root mean square error [CV(RMSE)], Equations (28) and (29) respectively, are widely used. In these equations, the values are summed for each time step (e.g., monthly or hourly values) over the course of an evaluation period (e.g., year), and the parameter V is the building performance variable under consideration (usually monthly wholebuilding energy consumption):
Vactual – Vmodeled NMBE = ------------------------------------------------------------ 100% N – 1 Mean V actual
(28)
V actual – V modeled ----------------------------------------------------------2
N–1 CV(RMSE) = ---------------------------------------------------------------- 100% Mean V actual
(29)
bTime
scales shown are hourly (H), daily (D), monthly (M), and subhourly (S).
cVariables include temperature (T ), humidity (H), solar (S), wind (W), time (t),
and thermal mass (tm).
where Vactual = parameter’s measured or metered value for each time step (e.g. month) Vmodeled = parameter’s estimated or modeled value for each time step N = number of time steps being analyzed during period of evaluation
ASHRAE Guideline 14-2014 states that a model can be considered calibrated if NMBE 5% and CV(RMSE) 15% when monthly data are used, or NMBE 10% and CV(RMSE) 30% when hourly data are used. Because CV(RMSE) is the positive average sum-squared error divided by the actual mean, it can be considered the percent error between the simulation and measured data. Because NMBE is a signed error divided by the mean, it indicates bias percent for under- (NMBE 0) or overshooting (NMBE 0) the actual data during the period of evaluation. Additional resources and tools useful for energy model calibration include ASHRAE RP-1051 (Reddy 2006), ASHRAE RP-1093 (Abushakra et al. 2000), and NREL Technical Report 5500-60127 (Robertson 2013). The main challenges of manual model calibration are that it is labor intensive and time consuming, it requires a high level of user skill and knowledge in both simulation and practical building operation, and the results often vary with the individual performing the calibration. Several practical difficulties prevent achieving a calibrated simulation or a simulation that closely reflects actual building performance, including (1) measurement and adaptation of weather data for use by simulation programs (e.g., converting global horizontal solar into beam and diffuse solar radiation), (2) collecting reliable actual meteorological year data for a specific building during the type period in which energy-use data was collected (Bhandari et al. 2012), (3) choice of methods used to calibrate the model, and (4) choice of methods used to measure required input parameters for the simulation (i.e., building mass, infiltration coefficients, and shading coefficients). Calibrated models typically involve a large number of input parameters to be calibrated, a high degree of expertise, multiple iterations, and substantial computing time. Every model, calibrated or not, carries assumptions and simplifications that are often deemed reasonable, but should be reevaluated when the model is used for different purposes. Bou-Saada and Haberl (1995a, 1995b), Bronson et al. (1992), Corson (1992), Garrett and New (2014), Haberl and Bou-Saada (1998), Kaplan et al. (1990), Liu and Liu (2011), Manke et al. (1996), Monfet et al. (2009), Norford et al. (1994), O’Neill et al. (2011), Reddy (2006, 2011), Reddy and Maor (2006), Reddy et al. (2007a, 2007b), Song and Haberl (2008a, 2008b), and Sun et al. (2016) provide examples of different methods used to calibrate
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19.36
2017 ASHRAE Handbook—Fundamentals (SI) Table 6
Method/Technique
Calibration Methods and Techniques
Description
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Detailed audit
Detailed audits are often conducted before building model development to gain a better knowledge of the building systems and characteristics (e.g., geometry, HVAC systems, lighting, equipment, occupancy schedules). Expert knowledge/templates/ Approaches that use model database • Expert knowledge or judgment as a key element of the process • Prior definition of typical building templates • Database of typical building parameters and components to reduce the requirement for user inputs during model development Intrusive testing Intrusive techniques require some intervention in operation of the actual building, such a “blink tests,” whereby groups of end-use loads (e.g., plug loads, lighting) are turned on and off in a controlled sequence to determine their overall effect on the baseline building load. High-resolution (“high-res”) data Data are recorded at hourly (or subhourly) levels as opposed to using daily load profiles or monthly utility bill data. Short-term energy monitoring Metering equipment is used to record on-site measurements for a short period of time (e.g., two weeks). This may be used in identifying typical energy end-use profiles and/or base loads. Graphical comparison Two- or three-dimensional plots are used to aid comparison of measured and simulated data in a way that can inform manual calibration. Whereas 2D scatterplots are frequently used, 3D techniques can allow effective visualization of larger quantities of data. Signature analysis Signature analysis techniques are a specific type of graphical analysis technique, typically used by HVAC simulation engineers to identify faulty parameters in air-handling unit (AHU) simulation. They may also be used to develop optimized operation and control schedules. Signature analysis methods are commonly used for the calibration of models based on the simplified energy analysis procedure (SEAP). Statistical displays Graphical representation of statistical indices and comparisons can facilitate intuitive interpretation for calibration. This includes data comparison techniques such as carpet plots, box-whisker-mean (BWM) plots, and monthly percent difference time-series graphs. Base case modeling The base case model refers to the use of measured base loads to calibrate the building model. Base loads refer to minimum, or weather independent, electrical and gas energy consumption. Calibration is carried out during the base case when heating and cooling loads are minimal and the building is dominated by internal loads, thus minimizing impact of weather dependent variables. Model parameter estimation Deduction of overall aggregate (or lumped) parameters (e.g., U-factors) using nonintrusive, measured data. Parameter reduction This involves reducing the requirement for detailed input for variable schedules (e.g., plug loads, lighting, occupancy, equipment). Day typing is one such approach, which works by analyzing long-term data and reducing this to manageable typical day-type schedules (e.g., weekdays versus weekends, winter versus summer). Zone typing may also be used to reduce large models into similar thermal zones (e.g., core, perimeter, offices, unoccupied spaces). Data disaggregation Data disaggregation refers to the application of nonintrusive load monitoring (NILM) techniques to decouple multiple derived data streams (e.g., lighting, miscellaneous plug loads) from a single measured data stream (e.g. whole-building electrical energy consumption). Evidence-based model Evidence-based approaches may be described as those that implement a procedural approach to model development, development making changes according to source evidence rather than ad hoc intervention. Strictly, this approach should account for adjustments to model parameters in a structured fashion (e.g., using version control software). Sensitivity analysis Sensitivity analysis procedures may be used in some studies to assess the influence of input parameters on model predictions. This information may be used to identify important parameters for measurement, calibration, or detailed investigation. Uncertainty quantification Parameter uncertainty may be used to directly assist in model calibration or provide a basis for quantification of risk propagation to the results (e.g., uncertainty-related risk in energy conservation measure analysis). Bayesian calibration Bayesian calibration is an alternative statistical approach to model calibration. The approach offers the advantage of naturally accounting for uncertainty in model prediction using prior input distributions. Pattern-based approach A process to calibrate energy models using patterns of monthly simulated and measured energy consumption using pattern fit criteria. Multiobjective optimization Most automated calibration techniques use a single- or multiobjective optimization function to reduce the difference between measured and simulated data. An objective function may be used to set a target of minimizing, for example, the mean square error between measured data and simulation output. Conversely, a penalty function may also be used to reduce the likelihood of deviating too far from the default value for a parameter.
simulation models. A methodology for testing calibration methods is described in the section on Model Validation and Testing. Calibration techniques can be roughly classified as either manual or automated methods. Manual calibration methods include graphical analysis and sensitivity analysis. Examples of methods used for automated calibration include Bayesian analysis, pattern matching, and multiobjective optimization. Manual calibration is an iterative approach that can be labor intensive and involves separate manipulation of individual parameters. This approach involves using an existing building simulation model and “tuning” or calibrating the various input parameters so that simulation program output matches with observed energy use. Calibration can be performed using data from any time period (e.g., monthly, or only a few weeks over the year), but the final calibrated
model is likely to be less accurate when fewer data are used during the calibration process. Hourly monitored energy data (most compatible with the time step adopted by most building energy simulation programs) can allow development of more accurate calibrated models, but calibrators often work without hourly data. During the manual calibration process, graphical representations and/or statistics comparing modeled data to measured data are displayed in an attempt to elucidate the value that input parameters could be set to in order to improve the match between simulation output and measured data. Calibrators often use sensitivity analyses to focus calibration efforts on the parameters that make the biggest difference in terms of energy use. In contrast to manual calibration, automated techniques use mathematical, algorithmic techniques implemented as computer software. Bayesian analysis, pattern-based
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Energy Estimating and Modeling Methods calibration, and multiobjective optimization are methods used for automated calibration.
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7.1
BAYESIAN ANALYSIS
This approach uses a theorem originally proposed by Thomas Bayes, an English statistician (1701-1761). Bayesian analysis provides an updated, more refined (more accurate) model based on the model from the prior iteration and a likelihood function derived from a statistical model for the measured data. Individual parameters with varying degrees of uncertainty can be analyzed to predict the range of values for that parameter, and its most likely value. In most Bayesian analyses, the formulation of the likelihood function is based on the work of Kennedy and O’Hagan (2001). Riddle and Muehleisen (2014) stated that, although Bayesian analysis can initially seem overwhelming, this method of calibration does not rely on extensive expertise. They further stated that this method provides an automated process for optimizing parameter values while accounting for multiple sources of uncertainty. Heo (2011) suggested that there are three primary strengths with Bayesian analysis: (1) it improves the reliability of a model by tuning important uncertain parameters in the model to represent actual building operations, (2) it results in calibrated models that are suitable to uncertainty analysis, and (3) the calibration procedure objectively quantifies uncertainty through the determination of a set of calibration parameters prioritized by their importance to the model outcome.
7.2
PATTERN-BASED APPROACH
This approach (Sun et al. 2016) automates a process to calibrate energy models using patterns of monthly simulated and measured energy use. The process contains four key steps: (1) running the original precalibrated energy model to obtain monthly simulated electricity and gas use; (2) establishing a pattern bias, either universal or seasonal, by comparing load shape patterns of simulated and actual monthly energy use; (3) using programmed logic to select which parameter to tune first based on bias pattern, weather, and input parameter interactions; and (4) automatically tuning the calibration parameters and checking the progress using pattern-fit criteria. The automated calibration algorithm was implemented in the Commercial Building Energy Saver (Hong et al. 2015), a web-based building energy retrofit analysis toolkit. The novelty of the developed calibration methodology lies in linking parameter tuning with the underlying logic associated with bias pattern identification. Although there are some limitations (e.g., coverage of building and system types) to the current implementation, the pattern-based automated calibration methodology can be universally adopted as an alternative to manual or hierarchical calibration approaches.
7.3
MULTIOBJECTIVE OPTIMIZATION
This approach uses the well-established field of mathematical optimization to select the best value of an input parameter, from some set of available alternatives, with regard to a fitness function. Standard optimization can be used when a single fitness criterion (e.g., whole-building energy use) is considered, and multiobjective optimization can be used to evaluate any number of building performance criteria (e.g., submetering, temperatures, relative humidity, heat flux) for which a user may have measured data. This class of algorithms is advantageous in that it is fully automated, can calibrate hundreds of (uncertain) input parameters, and is simulation-engineagnostic (i.e., it can apply to any simulation/software that has inputs and generates outputs comparable to measured data). Optimization algorithms can also accommodate more intensive submetering expected in the future, because fitness functions can compare calibration performance to either monthly (12), hourly (8760), or submetered, subhourly (1,000,000+) data points in less than a second.
19.37 Table 7
ANSI/ASHRAE Standard 140 Validation Test Matrix
Test Type/ Building Thermal Fabric Model Type (Basic Building Physics)a
Mechanical Equipment and On-Site Energy Generation Equipmenta
Analytical Ground-Coupled Heat HVAC BESTEST Volume 1 verification Transfer BESTEST for Slab (5.3) On Grade (5.2.4) Gas-Forced Air Furnace BESTEST (5.4.1, 5.4.2) Airside HVAC BESTEST Volume 1b (Neymark et. al. 2016) Comparative Thermal Fabric BESTEST tests (5.2.1-5.2.3) Residential Building BESTEST (7.2) Empirical validation
—
HVAC BESTEST Volume 2 (5.3) Gas-Forced Air Furnace BESTEST (5.4.3) —
aSections
of Standard 140 relevant to a given test suite are indicated in parentheses [e.g., “(5.2.4)” indicates Section 5.2.4]. as Addendum a to Standard 140-2014 in 2017.
bAnticipated
The disadvantage is that these algorithms, by themselves, do not capture the sense-making common of human calibrators and could thus generate physically unrealistic models that may not even simulate. A method of comparing these techniques was defined (New et. al 2012), studied (Edwards 2013), and evaluated (Sanyal 2014), as defined in the section on Testing Model Calibration Techniques Using Synthetic Data, to determine (Garrett and New 2014) the best algorithm, of those tested, for calibration. In terms of robustness, speed, and calibration accuracy over a 20,000 building calibration study, the best algorithm was an evolutionary algorithm known as the nondominated sorting genetic algorithm (NSGA-II) (Deb et al. 2002). Optimal values for NSGA-II, mechanisms for speedup, and methods to mitigate limitations regarding physical realism were identified for the traditional use cases of calibrating to monthly (Garret et al. 2013) and hourly (Garrett and New 2015) wholebuilding electrical data before being made publicly available (New 2016).
8.
VALIDATION AND TESTING
ANSI/ASHRAE Standard 140 was developed to identify and diagnose differences in predictions that may be caused by algorithmic differences, modeling limitations, faulty coding, or input errors. The methodological structure of Standard 140 allows all elements of a complete validation approach to be added as they become available. Table 7 shows the structure and the test suites in Standard 1402014. The table is an abbreviated way of representing the parameter space in which building energy simulation programs operate. Each cell in the matrix represents a large region in the space, so listing a test suite in a cell does not obviate the need for additional suites in that cell, and tests are needed in the empty cells. Class 1 tests in section 5 of the standard are based on procedures developed by the National Renewable Energy Laboratory (NREL) and field-tested by the International Energy Agency (IEA) over three IEA research tasks (Judkoff and Neymark 1995a; Neymark and Judkoff 2002, 2004). An additional section 5 test suite, which follows NREL’s methodology, was developed by Natural Resources Canada (Purdy and Beausoleil-Morrison 2003). Class 2 tests (section 7 of the standard) are based on procedures developed by NREL and field tested by the Home Energy Rating Systems (HERS) Council Technical Committee (Judkoff and Neymark 1995b). Additional tests not yet in
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19.38
2017 ASHRAE Handbook—Fundamentals (SI) Table 8 Validation Techniques
Technique
Advantages
Empirical validation (test of model and solution process)
• Approximate truth standard within experimental • Experimental uncertainties: accuracy • Instrument calibration, spatial/temporal discretization • Instrumentation can alter the effect being measured • Any level of complexity • Imperfect knowledge/specification of experimental object (building) being simulated
Disadvantages
• High-quality, detailed measurements are expensive and time consuming • Only a limited number of test cases are practical • Diagnostics can be difficult • Requires empirical determination of inputs • Only a limited number of sensitivity test cases are practical Analytical verification (test of solution process)
• No input uncertainty • Exact mathematical or secondary mathematical truth standard for given model • Inexpensive
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Comparative testing (relative test of model • No input uncertainty and solution process) • Any level of complexity • Many diagnostic comparisons possible • Inexpensive and quick
• No test of model validity or suitability • Limited to highly constrained cases for which analytical or quasi-analytical solutions can be developed* • No absolute truth standard (only statistically based acceptance ranges are possible)
Source: Judkoff and Neymark (2006). *Use of verified numerical solutions can extend the analytical verification approach to more realistic cases (Neymark et al. 2008).
Standard 140 were developed under ASHRAE research projects (Spitler et al. 2001; Yuill and Haberl 2002; Yuill et al. 2006), and under joint IEA Solar Heating and Cooling Programme/Energy Conservation in Buildings and Community Systems Task 34/Annex 43 (Judkoff and Neymark 2009); these may further populate the validation test matrix in the future. Standard 140 is cited by a number of codes, standards, and regulatory bodies, including ASHRAE Standard 90.1, the DOE Qualified Software List under IRS 179D energy tax credit regulations for commercial buildings (Energy.Gov 2015), the RESNET Home Energy Rating System software approval process (RESNET 2015), the International Energy Conservation Code (IECC 2015), the International Green Construction Code® (IgCC 2015), and ASHRAE Standard 189.1.
be perfect, while the model remains inappropriate for a given physical situation, such as using a one-dimensional conduction model where two-dimensional conduction dominates. The term “truth standard” represents the standard of accuracy for predicting real behavior. An analytical solution is a “mathematical truth standard,” and only tests the solution process for a model, not the appropriateness of the model. An approximate truth standard from an experiment tests both the solution process and appropriateness of the model within experimental uncertainty. The ultimate (or “absolute”) validation truth standard would be comparison of simulation results to a perfectly performed empirical validation experiment, with all simulation inputs perfectly defined.
Empirical Validation 8.1
METHODOLOGICAL BASIS
There are three ways to evaluate a whole-building energy simulation program’s accuracy (Judkoff et al. 1983/2008; Judkoff and Neymark 2006, 2009): • Empirical validation, which compares calculated results from a program, subroutine, algorithm, module, or software object to monitored data from a real building, test cell, or laboratory experiment • Analytical verification, which compares the output from a program, subroutine, algorithm, module, or software object to results from a known analytical solution or to results from a set of closely agreeing quasi-analytical solutions or verified numerical models • Comparative testing, which compares a program to itself or to other programs Neymark and Judkoff (2002) summarize approximately 100 articles and research papers on analytical, empirical, and comparative testing, from 1980 through 2015. Some of these and other works are listed by subject in the Bibliography. Table 8 compares these techniques (Judkoff 1988; Judkoff and Neymark 2006, 2009; Judkoff et al. 1983/2008). In this table, the term “model” is the representation of reality for a given physical behavior. For example, heat transfer may be simulated with one-, two-, or three-dimensional thermal conduction models. The term “solution process” encompasses the mathematics and computer coding to solve a given model. The solution process for a model can
Establishing an absolute truth standard for evaluating a program’s ability to analyze physical behavior requires empirical validation, but this is only possible within the range of measurement uncertainty, including that related to instruments, spatial and temporal discretization, and the overall experimental design. Full-scale test cells and buildings are large, relatively complex experimental objects. The exact design details, material properties, and construction in the field cannot be perfectly known, so there is uncertainty about the simulation model inputs that accurately represent the experimental object. Meticulous care is required to describe the experimental apparatus as clearly as possible to modelers to minimize this uncertainty. This includes experimental determination of as many material properties and other simulation model inputs as possible, including overall building parameters such as overall steady-state heat transmission coefficient (UAo), infiltration (in the form of an effective UA), and thermal capacitance (Judkoff et al. 2000; Neymark et al. 2005; Subbarao 1988). Measuring these overall parameters allows for a closure check on the individual parameters that comprise the overall parameters (e.g., building envelope material properties or individual steady-state envelope component conductances that sum up to the measured overall steady-state heat transmission coefficient). Also required are detailed on-site meteorological measurements. For example, many experiments measure global horizontal solar radiation, but very few experiments measure the splits between direct, diffuse, and ground-reflected radiation, all of which are inputs to many whole-building energy simulation
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Energy Estimating and Modeling Methods
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programs. Small-scale test cells also have shortcomings because of similitude factors for air convection and the increased relative proportion of corner conditions, which increases the importance of 2D and 3D heat transfer. Side-by-side experimental design can help to dampen the effects of input uncertainties, if the effects of the intentional differences between test objects are robust compared to those of the unintentional differences. The National Renewable Energy Laboratory (NREL) divides empirical validation into different levels, because many past validation studies produced inconclusive results. The levels of validation depend on the degree of control over possible sources of error in a simulation. The higher the degree of control over error sources, the better the potential for diagnosing the causes of measured to modeled output differences. The error sources consist of seven types, divided into two groups (Judkoff et al. 1983/2008): External Error Types • Differences between actual building microclimate versus weather input used by the program • Differences between actual schedules, control strategies, effects of occupant behavior, and other effects from the real building versus those assumed by the program user • User error deriving building input files • Differences between actual physical properties of the building (including HVAC systems) versus those input by the user Internal Error Types • Differences between actual thermal transfer mechanisms in the real building and its HVAC systems versus the simplified model of those processes in the simulation (all models, no matter how detailed, are simplifications of reality) • Errors or inaccuracies in the mathematical solution of the models • Coding errors • Ambiguous, incomplete, or faulty documentation The simplest level of empirical validation compares a building’s actual long-term energy use to that calculated by a computer program, with no attempt to eliminate sources of discrepancy. Because this is similar to how a simulation tool is used in practice, it is favored by many in the building industry. However, it is difficult to interpret the results because all possible error sources are acting simultaneously. Even if there is good agreement between measured and calculated performance, possible offsetting errors prevent a definitive conclusion about the model’s accuracy. More informative levels of validation involve controlling or eliminating various combinations of error types and increasing the information density of output-to-data comparisons (e.g., comparing temperature and energy results at time scales ranging from subhourly to annual). At the most detailed level, all known sources of error are controlled to identify and quantify unknown error sources and to reveal causal relationships associated with error sources, thereby facilitating diagnostics. Long-term (ideally annual) overall energy use needed to maintain comfort has extra meaning as a metric, because it is an indicator of how important a fault or simplification in a simulation program is in the context of one of the model’s most important uses. A $15/yr error out of a $1500/yr energy bill is tolerable, but a $500/ yr error is not. Other measurements, such as fluxes through individual surfaces, or short-term temperatures, are important to establish cause and effect but are inherently less indicative by themselves of the overall importance of an error. For example, it is difficult to assess the importance of a 0.3 K absolute average hourly one-week error between simulated and measured zone temperature. These principles also apply to intermodel comparative testing and analytical verification. The more realistic the test case, the more difficult it is to establish causality and diagnose problems; the simpler and more controlled the test case, the easier it is to pinpoint sources of error or inaccuracy. Methodically building up from
19.39 simple, highly controlled cases to realistic cases one parameter at a time is useful for understanding the impact of assumptions and simplifications in building energy simulation programs.
Analytical Verification Analytical verification compares outputs from a program, subroutine, algorithm, or software object to results from a known analytical solution or to results from a set of closely agreeing quasianalytical solutions or verified numerical models. Here, the term analytical solution is the closed-form mathematical solution of a model that has an exact result for a given set of parameters and simplifying assumptions. The term quasi-analytical solution is the mathematical solution of a model for a given set of parameters and simplifying assumptions, which may require minor interpretation differences that cause minor results variations. Such a result may be computed by generally accepted numerical methods or other means, provided that such calculations occur outside the environment of a whole-building energy simulation program and can be scrutinized. The term verified numerical model is a numerical model with solution accuracy verified by close agreement with an analytical solution and/or other quasi-analytical or numerical solutions, according to a process that demonstrates convergence in the space and time domains. Such numerical models may be verified by applying an initial comparison with an analytical solution(s), followed by comparisons with other numerical models for incrementally more realistic cases where analytical solutions are not available. Mathematical Truth Standards. An analytical solution provides an exact mathematical truth standard, limited to highly constrained cases for which exact analytical solutions can be derived. A secondary mathematical truth standard can be established based on the range of disagreement of a set of closely agreeing verified numerical models or other quasi-analytical solutions. Once verified against all available classical analytical solutions, and compared with each other for a number of other diagnostic test cases that do not have exact analytical solutions, the secondary mathematical truth standard can be used to test other models as implemented within whole-building simulation programs. Although a closed-form analytical solution provides the best possible mathematical truth standard, a secondary mathematical truth standard greatly enhances diagnostic capability for identifying software bugs and modeling errors as compared to the purely comparative method. This is because the range of disagreement among the results that comprise the secondary truth standard is typically much narrower than the range of disagreement among whole-building simulations. The secondary mathematical truth standard also allows more realistic (less constrained) boundary conditions to be used in the test cases, extending the analytical verification method beyond the constraints inherent for classical analytical solutions. This extends the usefulness of analytical verification methods by applying comparative techniques methodically, as outlined below. Establishing Secondary Mathematical Truth Standards. The following methodology for verifying numerical models to develop a secondary mathematical truth standard facilitates extension of analytical verification techniques. The methodology applies to both development of test cases and implementation of the numerical models. The logic may be summarized as follows: • Identify or develop exact analytical solutions that may be used as mathematical truth standards for testing detailed numerical models using parameters and simplifying assumptions of the analytical solution. • Apply a numerical solution process that demonstrates convergence in the space and time domains for both the analyticalsolution test cases and additional test cases where numerical models are applied.
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19.40 • Once validated against the analytical solutions, use the numerical models to develop test cases that progress toward more realistic (less idealized) conditions but do not have exact analytical solutions. • Check the numerical models by rigorously comparing their results to each other while developing the more realistic cases. • Tight agreement for the numerical models versus the analytical solution (and versus each other for subsequent test cases) verifies them as a secondary mathematical truth standard based on the range of disagreement among them. • Use the verified numerical-model results as reference results for testing other models of the given behavior, which reside within whole building energy simulation computer programs.
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Example applications of establishing secondary mathematical truth standards are provided in Neymark et al. (2008, 2009). Other Considerations. The following general guidelines are helpful when developing effective empirical validation, analytical verification, and comparative test cases: • Make test cases as simple as possible, to minimize input errors. • Make test cases as robust as possible, to maximize signal to noise ratio for a tested feature. • Vary test cases incrementally (varying just a single parameter when possible) so disagreements among results can be quickly diagnosed. • For numerical models, check sensitivity to spatial and temporal discretization, length of simulation, convergence tolerance, iteration limits, etc., and demonstrate that modeling is at a level of detail where including further detail yields negligible sensitivity in the results; document such work in detailed modeler reports. • Use independently developed and implemented models, and revise the test specification as needed to accommodate various modeling approaches; this reduces bias by ensuring that the test specification clearly addresses different modeling approaches. • For resolving disagreements among results that comprise a secondary truth standard, it is helpful to use an additional, independent expert party not directly involved in developing the models or results being compared. • Corrections to models must have a clear mathematical or physical basis and must be consistently applied across all test cases. Arbitrary alteration of a model solely for the purpose of better matching a given data set is not allowed. • A greater number of results from independently developed quasianalytical solutions or verified numerical solution models may be helpful for diagnosing disagreements among them.
Combining Empirical, Analytical, and Comparative Techniques A comparison between measured and calculated performance represents a small region in an immense N-dimensional parameter space. Investigators are constrained to exploring relatively few regions in this space, yet would like to ensure that the results are not coincidental (e.g., not a result of offsetting errors) and represent the validity of the simulation elsewhere in the parameter space. Analytical and comparative techniques minimize the uncertainty of extrapolations around the limited number of sampled empirical domains. Table 9 classifies these extrapolations. Figure 19 shows one process to combine analytical, empirical, and comparative techniques. These three techniques may also be used together in other ways; for example, intermodel comparisons may be done before an empirical validation exercise, to better define the experiment and to help estimate experimental uncertainty by propagating all known error sources through one or more wholebuilding energy simulation programs (Hunn et al. 1982; Lomas et al. 1994).
2017 ASHRAE Handbook—Fundamentals (SI) Table 9 Obtainable Data Points
Types of Extrapolation Extrapolation
A few climates Short-term total energy use
Many climates Long-term total energy use, or vice versa Short-term (hourly) temperatures Long-term total energy use, or vice and/or fluxes versa A few equipment performance points Many equipment performance points A few buildings representing a few Many buildings representing many sets of variable and parameter sets of variable and parameter combinations combinations, or vice versa* Small-scale: simple test cells, Large-scale complex buildings with buildings, and mechanical systems; complex HVAC systems, or vice laboratory experiments versa Source: Judkoff and Neymark (2006). *Extrapolation can go both ways (e.g., from short- to long-term data and from long- to short-term data). This does not mean that such extrapolations are correct, only that researchers and practitioners have explicitly or implicitly made such inferences in the past.
Fig. 19 Validation Method (Judkoff et al. 1983/2008)
For the path shown in Figure 19, the first step is running the code against analytical verification test cases to check its mathematical solutions. Discrepancies must be corrected before proceeding further. Second, the code is run against high-quality empirical validation data, and errors are corrected. Diagnosing error sources can be quite difficult. Comparative techniques can be used to create diagnostic procedures (Achermann and Zweifel 2003; Judkoff 1988; Judkoff and Neymark 1995a, 1995b; Judkoff et al. 1980, 1983/ 2008; Morck 1986; Neymark and Judkoff 2002, 2004; Purdy and Beausoleil-Morrison 2003; Spitler et al. 2001; Yuill and Haberl 2002) and better define the empirical experiments. The third step is to check agreement of several different programs with different thermal solution and modeling approaches (that have passed through steps 1 and 2) in a variety of representative cases. This uses the comparative technique as an extrapolation tool. Deviations in the program predictions indicate areas for further investigation. When programs successfully complete these three stages, they are considered validated for cases where acceptable agreement was achieved (i.e., for the range of building, climate, and mechanical system types represented by the test cases). Once several detailed simulation programs have satisfactorily completed the procedure, other programs and simplified design tools can be tested against them. A validated code does not necessarily represent truth. It does represent
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Energy Estimating and Modeling Methods a set of algorithms that have been shown, through a repeatable procedure, to perform according to the current state of the art. Similarly, the example results and associated computer programs in the nonnormative sections of Standard 140 do not represent truth, but rather an attempt to identify the range of uncertainty in the current state of the art via a well-defined and repeatable procedure. It is anticipated that, as building energy simulation programs improve, the informative example results in Standard 140 will be periodically updated, and the range of differences among test case results may be reduced. The NREL methodology for validating building energy simulation programs has been generally accepted by the International Energy Agency (Irving 1988), ASHRAE Standard 140, ASHRAE Standard 90.1, and elsewhere, with refinements suggested by other researchers (Bland 1992; Bloomfield 1988, 1999; Guyon and Palomo 1999b; Irving 1988; Lomas 1991; Lomas and Bowman 1987; Lomas and Eppel 1992). Additionally, the Commission of European Communities has conducted considerable work under the PASSYS program (Jensen 1989; Jensen and van de Perre 1991).
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Testing Model Calibration Techniques Using Synthetic Data Calibration is commonly used in conjunction with energy retrofit audit models (Judkoff et al. 2011a, 2011b; Reddy et al. 2006). This test method was initially developed by NREL for testing calibration procedures used with residential retrofit audit software; however, the fundamental concept could also be applied in a commercial building context. Other terms frequently used to describe model calibration include model tuning, model true-up, and model reconciliation. Typically, residential and commercial model calibration has been implemented using monthly energy data collected from utility bills for an existing building that is about to receive an energy retrofit. Sometimes submetered, disaggregated, or higher-frequency data are also available. An audit gathers information about the building needed to construct an input file for a building energy simulation program. A calibration method reconciles model predictions with the data, and then the calibrated model is used to predict energy savings and energy cost savings from various combinations of retrofit measures. Many variations on this approach exist, including some where the savings predictions are subjected to calibration instead of, or along with, the model inputs. Although it is logical to use the building’s actual performance data to tune the model, it is not certain that this results in a model that better predicts post-retrofit energy savings. When calibrating a large number of inputs to a limited number of outputs (in mathematics, an underdetermined or overparameterized problem), there are many combinations of input parameters that result in a close match to the utility bill data, so a close match is not in itself proof of good calibration (Reddy et al. 2006). The lower the frequency or informational content of the building performance data, the lower the probability that the calibration actually improves the model and associated energy savings predictions. Therefore, any method to test calibration techniques should use as many of the following three figures of merit as possible: (1) accuracy of the savings prediction, (2) how closely the calibrated input parameter values match the actual parameter values, and (3) the goodness of fit between the modeled and measured data. A limiting factor in attempting to empirically validate calibration techniques is the lack of high-quality annual monthly pre- and post-retrofit energy data, (higher frequency and submetered data are better), good pre- and post-retrofit building characteristics data, local pre- and post-retrofit weather data, and the dates of the retrofit installations. Until enough such empirical data are available to researchers, an alternative analytical method can be used in which a simulation program is used to generate its own synthetic pre and post-retrofit energy performance data. The synthetic data may be used as a surrogate for actual data. The method follows these general procedures:
19.41 1. Specify a building for a test case and introduce input uncertainty into the test specification (this represents the uncertainty associated with developing inputs from audit survey data): (a) Perform sensitivity tests on inputs with potentially high uncertainties to determine their relative effects on outputs; select inputs that have both substantial uncertainties and effects on outputs as approximate inputs. (b) Specify an uncertainty range (approximate input range) for each approximate input. (c) Select explicit inputs from the approximate input ranges (those who perform the calibrations must not know the explicit inputs). 2. Perform simulations using explicit inputs to create synthetic utility bill data. (Currently, this is typically monthly data, but the method can be used to generate and test against higher- or lowerfrequency synthetic building energy performance data, or end-use data at varying levels of disaggregation, mimicking the availability of submetered data). (a) Perform simulations to generate post-retrofit energy savings results by adjusting appropriate base-case inputs, including explicit inputs, as specified for each retrofit case and combinations of cases. 3. Develop tested program results (those who do this must not know the explicit inputs): (a) Develop the preliminary base-case model for a given calibration scenario. (b) Predict energy savings using one of the following: i. Calibrate the base-case model inputs using synthetic utility bills (from step 2), then apply the specified retrofit cases to the calibrated model. ii. Apply the specified retrofit to the uncirculated base case model and then calibrate or correct energy savings predictions using the synthetic utility bills (without adjustment to base-case model inputs); for example, (Calibrated savings) = (Predicted savings) × (Base-case actual bills)/(Base-case predicted bills). iii. Other calibration methods. The test cases make no recommendation about how to perform calibrations. Any calibration method that seeks to improve energy savings predictions through use of pre-retrofit building energy performance data can be tested by this method. 4. Use the following comparisons as figures of merit to determine the usefulness of the calibration techniques being tested. Note that all three of the listed comparisons are important for assessing the accuracy of the calibration technique. A large disagreement in any one of them indicates the presence of compensating errors, or some other error. (a) Compare the savings predictions from the tested program and any associated calibration techniques, versus the savings predictions from the same program run with the explicit inputs. (b) Compare the goodness of fit between synthetic building energy performance data and the calibrated model output data for the pre-retrofit case(s). (c) For programs where a calibrated base-case model is applied (see step 3bi), compare tested program inputs resulting from the tested program’s calibration techniques to the randomly selected explicit inputs. The preceding method is a pure (isolated) test of the calibration technique: that is, the synthetic utility billing data are generated with the tested program, and the program accuracy related to building physics modeling is not tested. A pure calibration test requires (1) automated calibration where no human judgment is required that would be helped by knowing the explicit inputs, or (2) that the modeler running the calibration test does not know the
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19.42
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 20 Calibration Cases Conceptual Flow (Judkoff et al. 2011a)
explicit inputs used to develop the synthetic utility bills. This method facilitates self-testing of a calibration technique, and is useful in several ways, including (1) testing a single calibration method, (2) testing several calibration methods to determine under what test conditions each is best, and (3) investigating how much and what kind of informational content is needed in the synthetic calibration data to achieve good calibrations with different calibration methods. The pure calibration test, however, may not be practical for a certification test that must be administered by a third-party organization; for this case, a method developed by NREL (Judkoff et al. 2011a, 2011b) ensures that the person performing the test does not know the explicit inputs. The main feature of this test method is that several (preferably at least three) reference programs are used to generate the synthetic utility bills and create the reference energy savings data. The bills and savings are taken as the average of the reference program results. This method tests both the calibration technique, and how closely the physics models in the tested program match the physics models in the reference programs. Example acceptance criteria may be used to facilitate the comparison of energy savings predictions (Judkoff et al. 2011a). Figure 20 shows the overall conceptual approach to testing model calibration techniques.
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Djunaedy, E, J. Hensen, and M. Loomans. 2005. External coupling between CFD and energy simulation: Implementation and validation. ASHRAE Transactions 111(1):612-624. Gunay, H.B., W. O'Brien, I. Beausoleil-Morrison, and B. Huchuk. 2014. On adaptive occupant-learning window blind and lighting controls. Building Research & Information 42(6):1-18. Haberl, J.S., T.A. Reddy, I.E. Figuero, and M. Medina. 1997. Overview of LoanSTAR chiller monitoring—Analysis of in-situ chiller diagnostics using ASHRAE RP-827 test method. Paper presented at Cool Sense National Integrated Chiller Retrofit Forum, Presidio, San Francisco, September. Hensen, J., and R. Lamberts, eds. 2011. Building performance simulation for design and operation. Routledge Publishing, Taylor and Francis, Inc. IBPSA. 2016. References and links. International Building Performance Simulation Association. www.ibpsa.org/?page_id=54. Katipamula, S., and D.E. Claridge. 1993. Use of simplified systems model to measure retrofit energy savings. Transactions of the ASME Journal of Solar Energy Engineering 115(May):57-68. Kavanaugh, S., and S. Lambert. 2004. A bin method energy analysis for ground-coupled heat pumps. ASHRAE Transactions 110:535-542. Kota, S., and Haberl, J. 2009. Historical survey of daylighting calculations methods and their use in energy performance simulations. ESL-IC-0911-07, Energy Systems Laboratory, Texas A&M University, College Station, TX. Kusuda, T., and T. Alereza. 1982. Development of equipment seasonal performance models for simplified energy analysis methods. ASHRAE Transactions 88(2). Labs, K., J. Carmody, R. Sterling, L. Shen, Y. Huang, and D. Parker. 1988. Building foundation design handbook. ORNL Report Sub/86-72143/1. Oak Ridge National Laboratory, Oak Ridge, TN. Lachal, B., W.U. Weber, and O. Guisan. 1992. Simplified methods for the thermal analysis of multifamily and administrative buildings. ASHRAE Transactions 98. Laret, L. 1988. Boiler physical models for use in large scale building simulation. SCS User 1 Conference, Ostend. Landry, R.W., and D.E. Maddox. 1993. Seasonal efficiency and off-cycle flue loss measurements of two boilers. ASHRAE Transactions 99(2). Landry, R.W., D.E. Maddox, and D.L. Bohac. 1994. Field validation of diagnostic techniques for estimating boiler part-load efficiency. ASHRAE Transactions 100(1):859-875. Lee, W.D., M.M. Delichatsios, T.M. Hrycaj, and R.N. Caron. 1983. Review of furnace/boiler field test analysis techniques. ASHRAE Transactions 89(1B):700-705. Liu, M. and D.E. Claridge. 1998. Use of calibrated HVAC system models to optimize system operation. ASME Journal of Solar Energy Engineering 120:131. Lobenstein, M.S. 1994. Application of short-term diagnostic methods for measuring commercial boiler losses. ASHRAE Transactions 100(1):876890. Malkawi, A., and G. Augenbroe. 2004. Advanced building simulation. Routledge Publishing, Taylor and Francis, Inc. McDowell, T., S. Emmerich, J. Thornton, and G. Walton. 2003. Integration of airflow and energy simulation using CONTAM and TRNSYS. ASHRAE Transactions 109(2):757-770. Miller, D.E. 1980. The impact of HVAC process dynamics on energy use. ASHRAE Transactions 86(2):535-556. Mitalas, G.P. 1968. Calculations of transient heat flow through walls and roofs. ASHRAE Transactions 74(2):182-188. Niu, Z., and K.V. Wong. 1998. Adaptive simulation of boiler unit performance. Energy Conversion Management 39(13):1383-1394. Reddy, T.A., K.K. Andersen, and D. Niebur. 2003. Information content of incoming data during field monitoring: Application to chiller modeling. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 9(4):365-383. Scottish Government. 2016. Section six software. www.gov.scot/Topics /Built-Environment/Building/Building-standards/techbooks/sectsixprg. Seem, J., and E. Hancock. 1985. A method for characterizing the thermal performance of a solar storage wall from measured data. Thermal Performance of the Exterior Envelopes of Buildings III: Proceedings of ASHRAE/DOE/BTECC Conference, ASHRAE SP-49, pp. 1304-1315. Sever, F., K. Kissock, D. Brown, and S. Mulqueen. 2011. Estimating industrial building energy savings using inverse simulation. ASHRAE Transactions 117:348-355.
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Energy Estimating and Modeling Methods Shavit, G. 1995. Short-time-step analysis and simulation of homes and buildings during the last 100 years. ASHRAE Transactions 101(1):856-868. Shurcliff, W.A. 1984. Frequency method of analyzing a building’s dynamic thermal performance. Cambridge, MA. Sowell, E.F., and G.N. Walton. 1980. Efficient computation of zone loads. ASHRAE Transactions 86(1):49-72. Sowell, E.F., and D.C. Hittle. 1995. Evolution of building energy simulation methodology. ASHRAE Transactions 101(1):850-855. Spitler, J.D. 1996. Annotated guide to load calculation models and algorithms. ASHRAE. Strand, R., Fisher, D., Liesen, R., and Pedersen, C. 2002. Modular HVAC simulation and the future integration of alternative cooling systems in a new building energy simulation program. ASHRAE Transactions 108(2): 1107-1117. Subbarao, K. 1986. Thermal parameters for single and multi-zone buildings and their determination from performance data. SERI Report SERI/TR253-2617. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Subbarao, K., Lei, Y., and Reddy, T. 2011. The nearest neighborhood method to improve uncertainty estimates in statistical building energy models. ASHRAE Transactions 117:459-471. Thevenard, D. 2011. Methods for estimating heating and cooling degreedays to any base temperature. ASHRAE Transactions 117:884-891. Tierney, T.M., and C.J. Fishman. 1994. Filed study of “real world” gas steam boiler seasonal efficiency ASHRAE Transactions 100(1):891-897. U.K. Government. 2014. Department for Communities and Local Government approved software for the production of non domestic energy performance certificates. www.gov.uk/government/publications/department -for-communities-and-local-government-approved-software-for-the -production-of-non-domestic-energy-performance-certificates-epc. U.S. Army. 1979. BLAST, the building loads analysis and system thermodynamics program—Users manual. U.S. Army Construction Engineering Research Laboratory Report E-153. U.S. Department of Energy. 2001. International performance measurement & verification protocol, concepts and options for determining energy and water savings, vol. I. U.S. Department of Energy, Washington, D.C. U.S. Department of Energy. 2001. International Performance Measurement and Verification Protocol (IPMVP): Vol. I: Concepts and options for determining energy and water savings. DOE/GO-102001-1187. U.S. Department of Energy. 2001. International Performance Measurement and Verification Protocol (IPMVP): Vol. II: Concepts and practices for improved indoor environmental quality. DOE/GO-102001-1188. U.S. Department of Energy. 2003. International Performance Measurement and Verification Protocol (IPMVP): Vol. III: Concepts and practices for determining energy savings in new construction. U.S. Department of Energy. 2016. Energy-efficient new homes tax credit for home builders. energy.gov/savings/energy-efficient-new-homes-tax -credit-home-builders. Wasserman, P.D. 1989. Neural computing, theory and practice. Van Nostrand Reinhold, New York. Yuill, G.K. 1990. An annotated guide to models and algorithms for energy calculations relating to HVAC equipment. ASHRAE.
Analytical Verification Bland, B. 1993. Conduction tests for the validation of dynamic thermal models of buildings. Building Research Establishment, Garston, U.K. Bland, B.H., and D.P. Bloomfield. 1986. Validation of conduction algorithms in dynamic thermal models. Proceedings of the CIBSE 5th International Symposium on the Use of Computers for Environmental Engineering Related to Buildings, Bath, U.K. CEN. 2004. PrEN ISO 13791. Thermal performance of buildings—Calculation of internal temperatures of a room in summer without mechanical cooling—General criteria and validation procedures. Final draft. Comité Européen de la Normalisation, Brussels. Judkoff, R., D. Wortman, and B. O’Doherty. 1981. A comparative study of four building energy simulations, Phase II: DOE-2.1, BLAST-3.0, SUNCAT-2.4, and DEROB-4. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Neymark, J., R. Judkoff, G. Knabe, H. Le, M. Durig, A. Glass, and G. Zweifel. 2002. An analytical verification procedure for testing the ability of whole-building energy simulation programs to model space conditioning equipment at typical design conditions. ASHRAE Transactions 108.2:1144-1154.
19.51 Pinney, A., and M. Bean. 1988. A set of analytical tests for internal longwave radiation and view factor calculations. Final Report of the BRE/ SERC Collaboration, vol. II, Appendix II.2. Building Research Establishment, Garston, U.K. Rees, S., D. Xiao, and J. Spitler. 2002. An analytical verification test suite for building fabric models in whole building energy simulation programs. ASHRAE Transactions 108:30-41. Rodriguez, E., and S. Alvarez. 1991. Solar shading analytical tests (I). Universidad de Savilla, Seville. San Isidro, M. 2000. Validating the solar shading test of IEA. Centro de Investigaciones Energeticas Medioambientales y Tecnologicas, Madrid. Stefanizzi, P., A. Wilson, and A. Pinney. 1988. The internal longwave radiation exchange in thermal models, vol. II, Ch. 9. Final Report of the BRE/ SERC Collaboration. Building Research Establishment, Garston, U.K. Tuomaala, P., ed. 1999. IEA task 22: A working document of subtask A.1, analytical tests. VTT Building Technology, Espoo, Finland. Tuomaala, P., K. Piira, J. Piippo, and C. Simonson. 1999. A validation test set for building energy simulation tools results obtained by BUS++. VTT Building Technology, Espoo, Finland. Walton, G. 1989. AIRNET—A computer program for building airflow network modeling. Appendix B: AIRNET Validation Tests. NISTIR 894072. National Institute of Standards and Technology, Gaithersburg, MD. Wortman, D., B. O’Doherty, and R. Judkoff. 1981. The implementation of an analytical verification technique on three building energy analysis codes: SUNCAT 2.4, DOE 2.1, and DEROB III. SERI/TP-721-1008, UL-59c. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Zhai, J., M. Krarti, and M-H. Johnson. 2010. Assess and implement natural and hybrid ventilation models in whole-building energy simulations. ASHRAE Research Project RP-1456. See also Bland (1992), Guyon and Palomo (1999b), Neymark and Judkoff (2002), Neymark et al. (2008), Purdy and Beausoleil-Morrison (2003), Spitler et al. (2001), and Yuill and Haberl (2002) in the References.
Empirical Validation Ahmad, Q., and S. Szokolay. 1993. Thermal design tools in Australia: A comparative study of TEMPER, CHEETAH, ARCHIPAK and QUICK. Building Simulation ’93, Adelaide, Australia. International Building Performance Simulation Association. Barakat, S. 1983. Passive solar heating studies at the Division of Building Research. Building Research Note 181. Division of Building Research, Ottawa, ON. Beausoleil-Morrison, I., and P. Strachan. 1999. On the significance of modeling internal surface convection in dynamic whole-building simulation programs. ASHRAE Transactions 105(2). Bloomfield, D., Y. Candau, P. Dalicieux, S. DeLille, S. Hammond, K. Lomas, C. Martin, F. Parand, J. Patronis, and N. Ramdani. 1995. New techniques for validating building energy simulation programs. Proceedings of Building Simulation ’95, Madison, WI. International Building Performance Simulation Association. Boulkroune, K., Y. Candau, G. Piar, and A. Jeandel. 1993. Modeling and simulation of the thermal behavior of a dwelling under ALLAN. Building Simulation ’93, Adelaide, Australia. International Building Performance Simulation Association. Bowman, N., and K. Lomas. 1985. Empirical validation of dynamic thermal computer models of buildings. Building Service Engineering Research and Technology 6(4):153-162. Bowman, N., and K. Lomas. 1985. Building energy evaluation. Proceedings of the CICA Conference on Computers in Building Services Design, Nottingham, pp. 99-110. Construction Industry Computer Association. Burch, J., D. Wortman, R. Judkoff, and B. Hunn. 1985. Solar Energy Research Institute validation test house site handbook. LA-10333-MS and SERI/PR-254-2028. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO, and Los Alamos National Laboratory, NM. Chantrasrisalai, C., D. Fisher, I. Iu, and D. Eldridge. 2003. Experimental validation of design cooling load procedures: The heat balance method. ASHRAE Transactions 109(2):160-173. Chantrasrisalai, C., V. Ghatti, D. Fisher, and D. Scheatzle. 2003. Experimental validation of the EnergyPlus low-temperature radiant simulation. ASHRAE Transactions 109(2):614-623.
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19.52 David, G. 1991. Sensitivity analysis and empirical validation of HLITE using data from the NIST indoor test cell. Proceedings of Building Simulation ’91, Nice, France. International Building Performance Simulation Association. Eppel, H., and K. Lomas. 1995. Empirical validation of three thermal simulation programs using data from a passive solar building. Proceedings of Building Simulation ’95, Madison, WI. International Building Performance Simulation Association. Fisher, D.E., and C.O. Pedersen. 1997. Convective heat transfer in building energy and thermal load calculations. ASHRAE Transactions 103(2): 137-148. Felsmann, C. 2008. Mechanical equipment and control strategies for a chilled water and a hot water system. Dresden, Germany: Technical University of Dresden. archive.iea-shc.org/publications/downloads/task34 -subtaskd.pdf. Gong, X., D. Claridge, and D. Archer. 2010. Infiltration investigation of a radiantly heated and cooled office. ASHRAE Transactions 116:437-446. Guyon, G., and N. Rahni. 1997. Validation of a building thermal model in CLIM2000 simulation software using full-scale experimental data, sensitivity analysis and uncertainty analysis. Proceedings of Building Simulation ’97, Prague. International Building Performance Simulation Association. Guyon, G., S. Moinard, and N. Ramdani. 1999. Empirical validation of building energy analysis tools by using tests carried out in small cells. Proceedings of Building Simulation ’99, Kyoto. International Building Performance Simulation Association. Haddad, K. 2004. An air-conditioning model validation and implementation into a building energy analysis software. ASHRAE Transactions 110:4654. Izquierdo, M., G. LeFebvre, E. Palomo, F. Boudaud, and A. Jeandel. 1995. A statistical methodology for model validation in the ALLAN simulation environment. Proceedings of Building Simulation ’95, Madison, WI. International Building Performance Simulation Association. Jensen, S. 1993. Empirical whole model validation case study: The PASSYS reference wall. Proceedings of Building Simulation ’93, Adelaide, Australia. International Building Performance Simulation Association. Judkoff, R., and D. Wortman. 1984. Validation of building energy analysis simulations using 1983 data from the SERI Class A test house (draft). SERI/TR-253-2806. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Judkoff, R., D. Wortman, and J. Burch. 1983. Measured versus predicted performance of the SERI test house: A validation study. SERI/TP-2541953. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Kalyanova, O., and P. Heiselberg. 2009. Empirical validation of building simulation software: Modeling of double facades, final report. DCE Technical Report No. 027. Aalborg, Denmark: Aalborg University. archive.iea-shc .org/publications/downloads/DSF_EMPIRICAL REPORT_18Jun2009 .pdf. LeRoy, J., E. Groll, and J. Braun. 1997. Capacity and power demand of unitary air conditioners and heat pumps under extreme temperature and humidity conditions. ASHRAE Research Project RP-859, Final Report. LeRoy, J., E. Groll, and J. Braun. 1998. Computer model predictions of dehumidification performance of unitary air conditioners and heat pumps under extreme operating conditions. ASHRAE Transactions 104(2). Lomas, K., and N. Bowman. 1986. The evaluation and use of existing data sets for validating dynamic thermal models of buildings. Proceedings of the CIBSE 5th International Symposium on the Use of Computers for Environmental Engineering Related to Buildings, Bath, U.K. Loutzenhiser, P., H. Manz, and G. Maxwell. 2007. Empirical Validations of Shading/Daylighting/Load Interactions in Building Energy Simulation Tools. Swiss Federal Laboratories for Materials Testing and Research; Iowa Energy Center. Available at archive.iea-shc.org/task34 /publications/index.html. Martin, C. 1991. Detailed model comparisons: An empirical validation exercise using SERI-RES. Contractor Report to U.K. Department of Energy, ETSU S 1197-p9. Maxwell, G., P. Loutzenhiser, and C. Klaassen. 2003. Daylighting—HVAC interaction tests for the empirical validation of building energy analysis tools. Iowa State University, Department of Mechanical Engineering, Ames. task22.iea-shc.org/data/sites/1/publications/IEA%20Daylighting %20Report%20Final.pdf.
2017 ASHRAE Handbook—Fundamentals (SI) Maxwell, G., P. Loutzenhiser, and C. Klaassen. 2004. Economizer control tests for the empirical validation of building energy analysis tools. Iowa State University, Department of Mechanical Engineering. Iowa Energy Center. task22.iea-shc.org/data/sites/1/publications/IEA%20Economizer% 20Report.pdf. McFarland, R. 1982. Passive test cell data for the solar laboratory winter 1980-81. LA-9300-MS. Los Alamos National Laboratory, NM. Moinard, S., and G. Guyon. 1999. Empirical validation of EDF ETNA and GENEC test-cell models. Final Report, IEA SHC Task 22, Building Energy Analysis Tools, Project A.3. Electricité de France, Moret sur Loing. archive.iea-shc.org/publications/downloads/Final_report.pdf. Neymark, J., P. Girault, G. Guyon, R. Judkoff, R. LeBerre, J. Ojalvo, and P. Reimer. 2005. The “ETNA BESTEST” empirical validation data set. Proceedings of Building Simulation 2005, Montreal. International Building Performance Simulation Association. Nishitani, Y., M. Zheng, H. Niwa, and N. Nakahara. 1999. A comparative study of HVAC dynamic behavior between actual measurements and simulated results by HVACSIM+(J). Proceedings of Building Simulation ’99, Kyoto. International Building Performance Simulation Association. Rahni, N., N. Ramdani, Y. Candau, and G. Guyon. 1999. New experimental validation and model improvement tools for the CLIM2000 energy simulation software program. Proceedings of Building Simulation ’99, Kyoto. International Building Performance Simulation Association. Sullivan, R. 1998. Validation studies of the DOE-2 building energy simulation program. LBNL-42241, Final Report. Lawrence Berkeley National Laboratory, CA. Travesi, J., G. Maxwell, C. Klaassen, M. Holtz, G. Knabe, C. Felsmann, M. Achermann, and M. Behne. 2001. Empirical validation of Iowa Energy Resource Station building energy analysis simulation models. Report, IEA SHC Task 22, Subtask A, Building Energy Analysis Tools, Project A.1 Empirical Validation. Centro de Investigaciones Energeticas, Medioambientales y Technologicas, Madrid. archive.iea-shc.org /publications/downloads/Iowa_Energy_Report.pdf. Trombe, A., L. Serres, and A. Mavroulakis. 1993. Simulation study of coupled energy saving systems included in real site building. Proceedings of Building Simulation ’93, Adelaide, Australia. International Building Performance Simulation Association. Walker, I., J. Siegel, and G. Degenetais. 2001. Simulation of residential HVAC system performance. Proceedings of eSim 2001, Natural Resources Canada, Ottawa, ON. Yazdanian, M., and J. Klems. 1994. Measurement of the exterior convective film coefficient for windows in low-rise buildings. ASHRAE Transactions 100(1):1087-1096. Zheng, M., Y. Nishitani, S. Hayashi, and N. Nakahara. 1999. Comparison of reproducibility of a real CAV system by dynamic simulation HVACSIM+ and TRNSYS. Proceedings of Building Simulation ’99, Kyoto. International Building Performance Simulation Association. See also Guyon and Palomo (1999a), Jensen (1989), Jensen and van de Perre (1991), Lomas (1991), Lomas and Bowman (1987), and Spitler et al. (1991) in the References.
Intermodel Comparative Testing Deru, M., R. Judkoff, and J. Neymark. 2003. Proposed IEA BESTEST ground-coupled cases. International Energy Agency, Solar Heating and Cooling Programme Task 22, Working Document. Fairey, P., M. Anello, L. Gu, D. Parker, M. Swami, and R. Vieira. 1998. Comparison of EnGauge 2.0 heating and cooling load predictions with the HERS BESTEST criteria. FSEC-CR-983-98. Florida Solar Energy Center, Cocoa. Haddad, K., and I. Beausoleil-Morrison. 2001. Results of the HERS BESTEST on an energy simulation computer program. ASHRAE Transactions 107(2). Haltrecht, D., and K. Fraser. 1997. Validation of HOT2000 using HERS BESTEST. Proceedings of Building Simulation ’97, Prague. International Building Performance Simulation Association. ISSO. 2003. Energie Diagnose Referentie Versie 3.0. Institut voor Studie en Stimulering van Onderzoekop Het Gebied van Gebouwinstallaties, Rotterdam, The Netherlands. Judkoff, R. 1985. A comparative validation study of the BLAST-3.0, SERIRES-1.0, and DOE-2.1 A computer programs using the Canadian direct gain test building (draft). SERI/TR-253-2652. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO.
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Energy Estimating and Modeling Methods Judkoff, R. 1985. International Energy Agency building simulation comparison and validation study. Proceedings of the Building Energy Simulation Conference, Seattle. Judkoff, R. 1986. International Energy Agency sunspace intermodel comparison (draft). SERI/TR-254-2977. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Judkoff, R., and J. Neymark. 1997. Home Energy Rating System Building Energy Simulation Test for Florida (Florida-HERS BESTEST), vols. 1 and 2. NREL/TP-550-23124. National Renewable Energy Laboratory, Golden, CO. www.nrel.gov/docs/legosti/fy97/23124a.pdf and www.nrel .gov/docs/legosti/fy97/23124b.pdf. Judkoff, R., and J. Neymark. 1998. The BESTEST method for evaluating and diagnosing building energy software. Proceedings of the ACEEE Summer Study 1998, Washington, D.C. American Council for an EnergyEfficient Economy. Judkoff, R., and J. Neymark. 1999. Adaptation of the BESTEST intermodel comparison method for proposed ASHRAE Standard 140P: Method of test for building energy simulation programs. ASHRAE Transactions 105(2). Judkoff, R., and J. Neymark. 2013. Twenty years on!: Updating the IEA BESTEST building thermal fabric test cases for ASHRAE Standard 140. Proceedings Building Simulation 2013, Chambéry, France, August 2528, 2013. International Building Performance Simulation Association. Also see the preprinted version, NREL CP-5500-58487. Golden, CO: National Renewable Energy Laboratory. www.nrel.gov/docs/ fy13osti /58487.pdf. Mathew, P., and A. Mahdavi. 1998. High-resolution thermal modeling for computational building design assistance. Proceedings of the International Computing Congress, Computing in Civil Engineering, Boston. Neymark, J., and R. Judkoff. 1997. A comparative validation based certification test for home energy rating system software. Proceedings of Building Simulation ’97, Prague. International Building Performance Simulation Association. Neymark, J., R. Judkoff, D. Alexander, C. Felsmann, P. Strachan, and A. Wijsman. 2008. International Energy Agency Building Energy Simulation Test and Diagnostic Method (IEA BESTEST) multi-zone non-airflow in-depth diagnostic cases: MZ320-MZ360. NREL Report No. TP-55043827. National Renewable Energy Laboratory, Golden, CO. www.nrel .gov/docs/fy08osti/43827.pdf. Neymark, J., R. Judkoff, D. Alexander, C. Felsmann, P. Strachan, and A. Wijsman, A. 2011. IEA BESTEST multi-zone non-airflow in-depth diagnostic cases. Proceedings of Building Simulation 2011, Sydney. International Building Performance Simulation Association. Preprint version, NREL CP-5500-51589, National Renewable Energy Laboratory, Golden, CO. www.nrel.gov/docs/fy12osti/51589.pdf. NRC. 2000. Benchmark test for the evaluation of building energy analysis computer programs. Natural Resources Canada, Ottawa, ON. (Translation of original Japanese version, approved by the Japanese Ministry of Construction.) RESNET. 2007. Procedures for verification of International Energy Conservation Code performance path calculation tools. RESNET Publication 07-003. Residential Energy Services Network, Inc., Oceanside, CA. www1.resnet.us/standards/iecc/procedures.pdf.
19.53 Sakamoto, Y. 2000. Determination of standard values of benchmark test to evaluate annual heating and cooling load computer program. Natural Resources Canada, Ottawa, ON. Soubdhan, T., T. Mara, H. Boyer, and A. Younes. 1999. Use of BESTEST procedure to improve a building thermal simulation program. Université de la Réunion, St Denis, La Reunion, France. Strachan, P., G. Kokogiannakis, L. MacDonald, and I. Beausoleil-Morrison. 2006. Integrated comparative validation tests as an aid for building simulation tool users and developers. ASHRAE Transactions 112:395-408. See also Achermann and Zweifel (2003), Judkoff et al. (1980, 1983), Judkoff and Neymark (1995a, 1995b), Judkoff et al. (2011a, 2011b), Morck (1986), Neymark and Judkoff (2004), and Purdy and Beausoleil-Morrison (2003) in the References.
General Testing and Validation Abushakra, B., and D. Claridge. 2008. Modeling office building occupancy in hourly data-driven and detailed energy simulation programs. ASHRAE Transactions 114(2):472-481. Allen, E., D. Bloomfield, N. Bowman, K. Lomas, J. Allen, J. Whittle, and A. Irving. 1985. Analytical and empirical validation of dynamic thermal building models. Proceedings of the First Building Energy Simulation Conference, Seattle, pp. 274-280. Beausoleil-Morrison, I. 2000. The adaptive coupling of heat and air flow modelling within dynamic whole-building simulation. Ph.D. dissertation. Energy Systems Research Unit, Department of Mechanical Engineering, University of Strathclyde, Glasgow. Bloomfield, D. 1985. Appraisal techniques for methods of calculating the thermal performance of buildings. Building Services Engineering Research & Technology. 6(1):13-20. Bloomfield, D., ed. 1989. Design tool evaluation: Benchmark cases. IEA T8B4. Solar Heating and Cooling Program, Task VIII: Passive and Hybrid Solar Low-Energy Buildings. Building Research Establishment, Garston, U.K. Bloomfield, D., K. Lomas, and C. Martin. 1992. Assessing programs which predict the thermal performance of buildings. BRE Information Paper IP7/92. Building Research Establishment, Garston, U.K. Gough, M. 1999. A review of new techniques in building energy and environmental modelling. Final Report, BRE Contract BREA-42. Building Research Establishment, Garston, U.K. Judkoff, R., S. Barakat, D. Bloomfield, B. Poel, R. Stricker, P. van Haaster, and D. Wortman. 1988. International Energy Agency design tool evaluation procedure. SERI/TP-254-3371. Solar Energy Research Institute (now National Renewable Energy Laboratory), Golden, CO. Palomo, E., and G. Guyon. 2002. Using parameters space analysis techniques for diagnostic purposes in the framework of empirical model validation. LEPT-ENSAM, Talance, France. Electricité de France, Moret sur Loing. See also Bloomfield (1988, 1999), Irving (1988), Judkoff et al. (1983), Judkoff (1988), Judkoff and Neymark (2006), and Lomas and Eppel (1992) in the References.
Related Commercial Resources
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Related Commercial Resources CHAPTER 20
SPACE AIR DIFFUSION Indoor Air Quality and Sustainability........................................................................................... Terminology .................................................................................................................................. Principles of Jet Behavior ............................................................................................................. Symbols .........................................................................................................................................
OOM air distribution systems are intended to provide thermal comfort and ventilation for space occupants and processes. Although air terminals (inlets and outlets), terminal units, fan-coil units, local ducts, and rooms themselves may affect room air diffusion, this chapter addresses only air inlets and outlets and their direct effect on occupant comfort. This chapter is intended to present HVAC designers the fundamental characteristics of air distribution devices. For information on naturally ventilated spaces, see Chapter 16. For a discussion of various air distribution strategies, tools, and guidelines for design and application, see Chapter 57 in the 2015 ASHRAE Handbook—HVAC Applications. Chapter 20 in the 2016 ASHRAE Handbook—HVAC Systems and Equipment describes the characteristics of various air inlets, outlets, fan-coil units, chilled beams, air curtain units, and terminal units, as well as selection tools and guidelines. Room air diffusion methods can be classified as one of the following as shown in Figure 1:
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R
• Mixed systems produce little or no thermal stratification of air within the space. Overhead air distribution is an example of this type of system. • Fully (thermally) stratified systems produce little or no mixing of air within the occupied space. Thermal displacement ventilation is an example of this type of system. • Partially mixed systems provide some mixing within the occupied and/or process space while creating stratified conditions in the volume above. Most underfloor air distribution and task/ambient conditioning designs are examples of this type of system. The preparation of this chapter is assigned to TC 5.3, Room Air Distribution.
Local temperature and carbon dioxide (CO2) concentration have similar stratification profiles. Air distribution systems, such as thermal displacement ventilation (TDV) and underfloor air distribution (UFAD), that deliver air in cooling mode at or near floor level and return air at or near ceiling level produce varying amounts of room air stratification. For floor-level supply, thermal plumes that develop over heat sources in the room play a major role in driving overall floor-to-ceiling air motion. The amount of stratification in the room is primarily determined by the balance between total room airflow and heat load. In practice, the actual temperature and concentration profile depends on the combined effects of various factors, but is largely driven by the characteristics of the room supply airflow and heat load configuration. For room supply airflow, the major factors are • • • •
Total room supply airflow quantity Room supply air temperature Diffuser type Diffuser throw height (or outlet velocity); this is associated with the amount of mixing provided by a floor diffuser (or room conditions near a low-sidewall TDV diffuser) For room heat loads, the major factors are
• Magnitude and number of loads in space • Load type (point or distributed source) • Elevation of load (e.g., overhead lighting, person standing on floor, floor-to-ceiling glazing) • Radiative/convective split • Whether pollutants are associated with heat sources
Fig. 1 Classification of Air Diffusion Methods
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20.2 20.2 20.3 20.8
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2017 ASHRAE Handbook—Fundamentals (SI)
INDOOR AIR QUALITY AND SUSTAINABILITY
Air diffusion methods affect not only indoor air quality (IAQ) and thermal comfort, but also energy consumption over the building’s life. Choices made early in the design process are important. Programs such as U.S. Green Building Council’s (USGBC 2013) Leadership in Energy and Environmental Design (LEED®) v4 rating system, which was originally created in response to indoor air quality concerns, now include prerequisites and credits for increasing ventilation rates and improving indoor environmental quality. These program requirements are sometimes achievable by following good room air diffusion design principles, methods, and standards (see Chapter 57 of the 2015 ASHRAE Handbook—HVAC Applications). ANSI/ASHRAE Standard 62.1 provides a table of typical values to help predict zone air distribution effectiveness. For example, well-designed ceiling-based air distribution systems produce nearperfect air mixing in cooling mode, and yield an air distribution effectiveness of 1.0. Displacement ventilation and underfloor air distribution (UFAD) systems have the potential for values greater than 1.0. More information on ceiling- and wall-mounted air inlets and outlets can be found in Rock and Zhu (2002). Displacement system performance is described in Chen and Glicksman (2003). ASHRAE’s (2013) UFAD Design Guide discusses UFAD in detail. More information on ANSI/ASHRAE Standard 62.1 is available in its user’s manual (ASHRAE 2010).
2.
TERMINOLOGY
Aspect ratio. Ratio of length to width of opening or core of a grille. Attached jet. A supply air jet drawn to a surface, parallel to the direction of airflow and caused by the Coanda effect. Axial jet. A supply air jet with a conical discharge profile. Centerline velocity. Maximum velocity of an air jet at any given cross section perpendicular to the direction of airflow. Coanda effect. Effect of a moving jet attaching to a parallel surface because of negative pressure developed between jet and surface. Coefficient of discharge. Ratio of area at vena contracta to free area of opening. Core area. Area of a register, grille, or linear slot diffuser pertaining to the inside of the frame or border. Diffusion. Distribution of air into a space. Distribution. Moving air to or in a space by an outlet discharging supply air. Draft. Current of air, when referring to localized effect (generally, the unwanted local cooling of the body caused by air movement) caused by one or more factors of high air velocity, low ambient temperature, or direction of airflow whereby more heat is withdrawn from a person’s skin than is normally dissipated. Drop. Vertical distance that the lower edge of a horizontally projected airstream descends between the outlet and the end of its throw. Effective area. Net area of an outlet or inlet device through which air can pass; equal to the free area times the coefficient of discharge. Entrainment. Air drawn into an air jet because of the pressure differential caused by the airstream discharged from the outlet. Entrainment ratio. Volumetric flow rate of total air (supply air plus entrained air) at a given distance from an outlet divided by the volumetric flow rate of supply air. Free area. Total minimum area of openings in an air outlet or inlet through which air can pass. Free jet. An air jet not obstructed or affected by walls, ceiling, or other surfaces.
Induction. Movement of space air into an air device. Induction ratio. Volumetric flow rate of induced air divided by volumetric flow rate of primary air. Inlet. A device that allows air to exit the zone (e.g., grilles, registers, diffusers) Isothermal jet. An air jet in which supply air temperature equals surrounding room air temperature. Linear jet. A supply air jet with a relatively high aspect ratio. Neck area. Nominal area of duct connection to air outlet or inlet. Nonisothermal jet. An air jet in which supply air temperature does not equal surrounding room air temperature. Occupied zone. The volume of space intended to be comfort conditioned for occupants (see ANSI/ASHRAE Standard 55). Outlet. A device discharging supply air into the space (e.g., grilles, registers, diffusers). Classified according to location and type of discharge. Outlet velocity. Average velocity of air discharging from an outlet. Primary air. Air delivered to an outlet or terminal device. Radial jet. A supply air jet that discharges 360° and expands uniformly. Spread. Divergence of an airstream in a horizontal and/or vertical plane after it leaves an outlet. Stratification height. Vertical distance from floor to horizontal plane that defines lower boundary of upper mixed zone in a fully stratified or partially mixed system. Stratified zone. Zone in which air movement is entirely driven by buoyancy caused by convective heat sources. Typically found in fully stratified or partially mixed systems. Supply Air. Air delivered into a zone from an outlet. Terminal velocity. An arbitrary specified centerline air velocity at a distance from an outlet. Throw. The distance from the centerline of an outlet perpendicular to a point in the mixed airstream where the velocity has been reduced to a specified terminal velocity (e.g., 0.25, 0.5, 0.75, or 1.0 m/s), defined by ASHRAE Standard 70. Total air. Combination of supply air and entrained air at a given distance from an outlet. Vena contracta. Smallest cross-sectional area of a fluid stream leaving an orifice.
Outlet Types and Characteristics Straub and Chen (1957) and Straub et al. (1956) classified outlets into five major groups (the subgrouping was added in 2017 and was not part of the original research): Group A1. Outlets mounted in or near the ceiling that discharge air horizontally (Figures 2 and 3). Group A2. Outlets discharging horizontally that are not influenced by an adjacent surface (free jet; Figure 4). Group B. Outlets mounted in or near the floor that discharge air vertically in a linear jet (Figure 5). Group C. Outlets mounted in or near the floor that discharge air vertically in a spreading jet (Figure 6). Group D. Outlets mounted in or near the floor that discharge air horizontally (Figure 7 and 8). When used in fully stratified systems (TDV), these outlets use low discharge velocities; in mixed systems, they use higher discharge velocities. Group E. Outlets that project supply air vertically downward (Figures 9 and 10). These outlets When used in partially stratified systems (e.g., laminar flow outlets, TDV), these outlets use low discharge velocities; in mixed systems (e.g., air curtain units, other downward directed ceiling devices, etc.), they use higher discharge velocities.
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Space Air Diffusion
20.3 3.
PRINCIPLES OF JET BEHAVIOR
Air Jet Fundamentals
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Fig. 2 Example Airflow Patterns of Outlet Group A1
Air supplied to rooms through various types of outlets can be distributed by turbulent air jets (mixed and partially mixed systems) or in a low-velocity, unidirectional manner (stratified systems). The air jet discharged from an outlet is a primary factor affecting room air motion. The jet boundary contours are not well defined and are easily affected by external influences. Baturin (1972), Christianson (1989), and Murakami (1992) have further information on the relationship between the air jet and occupied zone. If the supply air temperature is equal to the ambient room air temperature, the air jet is called an isothermal jet. A jet with an initial temperature different from the ambient air temperature is called a nonisothermal jet. The air temperature differential between supplied and ambient room air generates thermal forces (buoyancy) in jets, affecting the jet’s (1) trajectory, (2) location at which it attaches to and separates from the ceiling/floor, and (3) throw. The significance of these effects depends on the ratio between the thermal buoyancy of the air and jet momentum. If an air jet is not obstructed or affected by walls, ceiling, or other surfaces, it is considered a free jet. When outlet area is small compared to the dimensions of the space normal to the jet, the jet may be considered free as long as X 1.5 A R
Fig. 3 Example Airflow Patterns (Nonisothermal) of Outlet Group A1
where X = distance from face of outlet, m AR = cross-sectional area of confined space normal to jet, m2
Fig. 4
Fig. 5
Example Airflow Patterns (Isothermal) of Outlet Group A2
Example Airflow Patterns (Nonisothermal) of Outlet Group B
(1)
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2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 6
Example Airflow Patterns (Nonisothermal) of Outlet Group C
Fig. 7
Example Airflow Patterns (Nonisothermal) of Outlet Group D (High Velocity)
Fig. 8
Example Airflow Patterns (Nonisothermal) of Outlet Group D (Low Velocity)
• Zone 1 extends from the outlet face, in which the velocity and temperature of the airstream remains practically unchanged. • Zone 2 is a transition zone, with its length determined by the type of outlet, aspect ratio of the outlet, initial airflow turbulence, etc. • Zone 3 is a zone of jet degradation, where centerline air velocity and temperature differential decrease rapidly. Turbulent flow is fully established and may be 25 to 100 equivalent air outlet diameters long. The angle of divergence is well defined. Typically, free air jets diverge at a constant angle, usually ranging from 20 to 24°, with an average of 22°. Coalescing jets for closely spaced multiple outlets expand at smaller angles, averaging 18°, and jets discharging into relatively small spaces show even smaller angles of expansion (McElroy 1943). The angle of divergence is easily affected by external influences, such as local eddies, vortices, and surges. Internal forces governing this air motion are extremely delicate (Nottage et al. 1952a). • Zone 4 is important because, in most cases, the jet enters the occupied area in this zone. Distance to this zone and its length depend on the velocities and turbulence characteristics of ambient air. In a few diameters or widths, air velocity becomes less than 0.25 m/s. Centerline Velocities in Zones 1 and 2. In zone 1, the ratio Vx /Vo is constant for a given outlet and ranges between 1.0 and 1.2, equal to the ratio of the centerline velocity of the jet at the start of expansion to the average initial velocity. The ratio Vx /Vo varies from approximately 1.0 for rounded entrance nozzles to about 1.2 for straight pipe discharges; it has higher values for diverging discharge outlets. The aspect ratio (Tuve 1953) and turbulence (Nottage et al. 1952a) primarily affect centerline velocities in zones 1 and 2. Aspect ratio has little effect on the terminal zone of the jet when Ho is greater than 100 mm. This is particularly true of nonisothermal jets. When Ho is very small, induced air can penetrate the core of the jet, thus reducing centerline velocities. The difference in performance between a radial outlet with small Ho and an axial outlet with large Ho shows the importance of jet thickness. When air is discharged from relatively large perforated panels, the constant-velocity core formed by coalescence of individual jets extends a considerable distance from the panel face. In zone 1, when the aspect ratio is less than 5, use the following equation for estimating centerline velocities (Koestel et al. 1949): Vx = 1.2Vo C d R fa
(2)
In zone 2, the ratio Vx /Vo begins to decrease. Experimental evidence indicates that, in zone 2, Fig. 9
Example Airflow Patterns (Nonisothermal) of Outlet Group E (High Velocity)
Vx ------ = Vo
K c2 Ho ---------------X
(3)
where Vx Vo Vc Cd Rfa
Fig. 10
Example Airflow Patterns (Nonisothermal) of Outlet Group E (Low Velocity)
Jet Expansion Zones. The full length of an air jet, in terms of the maximum or centerline velocity and temperature differential at the cross section, can be divided into four zones:
= = = = =
centerline velocity at distance X from outlet, m/s Vc /Cd Rf a = average initial velocity at discharge, m/s nominal velocity of discharge based on core area, m/s coefficient of discharge (usually between 0.65 and 0.90) ratio of free area to core area
Ho = width of jet at outlet or at vena contracta, m Kc2 = centerline velocity constant, depending on outlet type and discharge pattern X (1/Kc2Ho )1/2 = distance from outlet to measurement of centerline velocity Vx, m
Centerline Velocity in Zone 3. In zone 3, centerline velocities of radial and axial isothermal jets can be determined accurately from the following equation:
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Table 1 Generic Values for Centerline Velocity Constant Kc3a for Commercial Supply Outlets for Fully and Partially Mixed Systems, Except UFAD Outlet Type
Discharge Pattern
High sidewall grilles (Figure 4) High sidewall linear
deflectionb
0° Wide deflection Core less than 100 mm highc Core more than 100 mm high Low sidewall Up and on wall, no spread (Figure 7) Wide spreadc Baseboard Up and on wall, no spread Wide spread Floor grille (Figure 5) No spreadc Wide spread 360° horizontald Ceiling (Figure 2) Four-way; little spread Horizontal/vertical along surfacec Ceiling linear slot (Figure 3) Horizontal/vertical free jetc Free jet (air curtain units)
Ao
Kc3a
Free Free Free Free Free Free Core Core Free Free Neck Neck Free Free Free
5.7 4.2 4.4 5.0 4.5 3.0 4.0 2.0 4.7 1.6 1.1 3.8 5.5 3.9 6.0
aGeneric
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values shown for example purposes bFree area is about 80% of core area. only. See manufacturer’s data for specific cFree area is about 50% of core area. dCone free area is greater than duct area. Kc3 values.
K c3 Q o K c3 Vo A o Vx = -------------------------- = ---------------X X Ao
(4)
Fig. 11
Zones of Expansion for Axial or Radial Air Jets
where Kc3 = centerline velocity constant (see Table 1 for generic values) Vo = Vc /CdRfa = average initial velocity at discharge, m/s Ao = free area, core area, or neck area as shown in Table 1 (obtained from outlet manufacturer), m2 Qo = volumetric flow rate of supply air, m3/s X = distance from face of outlet, m
For centerline velocities of linear jets, where Kc3 = Kc2, use Equation (3). The effective area, according to ASHRAE Standard 70, can be used in place of Ao in Equation (4) with the appropriate value of Kc3. Centerline Velocity in Zone 4. In zone 4, centerline velocities can be difficult to predict, based on the large dispersal pattern. Determining Centerline Velocities. To correlate data from all four zones, plot centerline velocity ratios against distance from the outlet in Figures 11 and 12. Airflow patterns of diffusers are related to the centerline velocity constants and throw distance. In general, diffusers with a circular airflow pattern (radial jet) have a shorter throw than those with a directional or cross-flow pattern (axial jet). During cooling, the circular pattern tends to curl back from the end of the throw toward the diffuser, reducing the drop and ensuring that the cool air remains near the ceiling. In cross-flow airflow patterns, the airflow does not roll back to the diffuser at the end of the throw, but continues to move away from the diffuser at low velocities. Throw. At a given supply airflow and centerline velocity, Equation (4) can be transposed into Equation (5) to determine the throw X of an outlet. The centerline velocity constant and appropriate outlet area Ao should be available from the outlet manufacturer. According to Figures 11 and 12, 0.25 m/s terminal velocity can occur in zone 4. When this occurs, an accepted practice to approximate throw in zone 4 is to reduce the calculated throw in zone 3 by 30%. K c3 Q o X = ----------------Vx Ao
(5)
Fig. 12 Zones of Expansion for Linear Air Jets See Informative Appendix B of ASHRAE Standard 70-2006 for the application of this methodology. The following example shows the use of Table 1 and Figures 11 and 12. Example 1. A 300 by 450 mm high sidewall grille with an 280 by 430 mm core area is selected. From Table 1, Kc3 = 5 for zone 3, and Ao should be 80% of the core area, in square metres. If the airflow is 0.3 m3/s, what is the throw to 0.25, 0.5, and 0.75 m/s? Solution: From Equation (5), K c3 Q o 5.7 0.3 1.71 X = ----------------- = --------------------------------------------------------------- = ---------------------V x 0.31 6 Vx Ao V x 0.8 280 430 10
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20.6
2017 ASHRAE Handbook—Fundamentals (SI) Solving for 0.25 m/s throw,
Qx 2X ------ = ----------------Qo Kc Ao
X = 1.71/(0.25 0.31) = 22 m However, according to Figures 11 and 12, 0.25 m/s is in zone 4, which is typically 30% less than calculated in Equation (4), or
By substituting from Equation (4), Qx Vo ------ = 2 -----Vx Qo
X = 22 0.70 = 15 m Solving for 0.5 m/s throw,
Solving for 0.75 m/s throw,
Qx 2 ------ = --------K c3 Qo
X = 1.71/(0.75 0.31) = 7 m
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Velocity Profiles of Jets. In zone 3 of both axial and radial jets, the velocity distribution may be expressed by a single curve (Figures 11 and 12) in terms of dimensionless coordinates; this same curve can be used as a good approximation for adjacent portions of zones 2 and 4. Temperature and density differences have little effect on cross-sectional velocity profiles. Velocity distribution in zone 3 can be expressed by the Gauss error function or probability curve, which is approximated by the following equation: 2
Vx = 3.3 log ----V
(8)
For a continuous slot with active sections up to 3 m and separated by 0.6 m,
X = 1.71/(0.5 0.31) = 11 m
r r----------- 0.5V
(7)
(6)
where r = radial distance of point under consideration from centerline of jet r0.5V = radial distance in same cross-sectional plane from axis to point where velocity is one-half centerline velocity (i.e., V = 0.5Vx) Vx = centerline velocity in same cross-sectional plane V = actual velocity at point being considered
Experiments show that the conical angle for r0.5V is approximately one-half the total angle of divergence of a jet. The velocity profile curve for one-half of a straight-flow turbulent jet (the other half being a symmetrical duplicate) is shown in Figure 13. For multiple-opening outlets, such as grilles or perforated panels, the velocity profiles are similar, but the angles of divergence are smaller. Entrainment Ratios. The following equations are for entrainment of circular jets and of jets from long slots. For third-zone expansion of circular jets,
X -----Ho
(9)
or, substituting from Equation (3), Qx ------ = Qo
Vo 2 -----Vx
(10)
where Qx Qo X Kc Ao
= = = = =
total volumetric flow rate at distance X from face of outlet, m3/s discharge from outlet, m3/s distance from face of outlet, m centerline velocity constant core area or neck area free (see Table 1), m2
The entrainment ratio Qx /Qo is important in determining total air movement at a given distance from an outlet. For a given outlet, the entrainment ratio is proportional to the distance X [Equation (7)] or to the square root of the distance X [Equation (9)] from the outlet. Equations (8) and (10) show that, for a fixed centerline velocity Vx, the entrainment ratio is proportional to outlet velocity. Equations (8) and (10) also show that, at a given centerline and outlet velocity, a circular jet has greater entrainment and total air movement than a long slot. Comparing Equations (7) and (9), the long slot should have a greater rate of entrainment. The entrainment ratio at a given distance is less with a large Kc3 than with a small Kc3.
Isothermal Radial Flow Jets In a radial jet, as with an axial jet, the cross-sectional area at any distance from the outlet varies as the square of this distance. Centerline velocity gradients and cross-sectional velocity profiles are similar to those of zone 3 of axial jets, and the angles of divergence are about the same.
Nonisothermal Jets When the temperature of introduced air is different from the room air temperature, the diffuser air jet is affected by thermal buoyancy caused by air density difference. The trajectory of a nonisothermal jet introduced horizontally is determined by the Archimedes number (Baturin 1972): gL o T o – T A Ar = -------------------------------2 Vo TA
(11)
where g = gravitational acceleration rate, m/s2 Lo = length scale of diffuser outlet equal to hydraulic diameter of outlet, m (To – TA) = initial temperature of jet – temperature of ambient air, °C Vo = initial air velocity of jet, m/s TA = room air temperature, K
Fig. 13
Cross-Sectional Velocity Profiles for Straight-Flow Turbulent Jets
The influence of buoyant forces on horizontally projected heated and chilled jets is significant in heating and cooling with wall outlets. Koestel’s (1955) equation describes the behavior of these jets. Helander and Jakowatz (1948), Helander et al. (1953, 1954, 1957), Knaak (1957), and Yen et al. (1956) developed equations for outlet characteristics that affect the downward throw of heated
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20.7
air. Koestel (1954, 1955) developed equations for temperatures and velocities in heated and chilled jets. Kirkpatrick and Elleson (1996) and Li et al. (1993) provide additional information on nonisothermal jets.
Nonisothermal Horizontal Free Jet A horizontal free jet rises or falls according to the temperature difference between it and the ambient environment. The horizontal jet throw to a given distance follows an arc, rising for heated air and falling for cooled air. Therefore, whether the equivalent temperature difference is positive or negative, the distance from the diffuser to a given terminal velocity along the discharge jet remains essentially the same.
Comparison of Free Jet to Attached Jet
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An attached jet entrains air along the exposed side of the jet, whereas a free jet can entrain air on all its surfaces. Because a free jet’s entrainment rate is larger compared to that of an attached jet, a free jet’s throw distance will be shorter. To calculate the throw distance X for a noncircular free jet from catalog data for an attached jet, the following estimate can be used. (12) Xfree = Xattached × 0.707 Jets from ceiling diffusers initially tend to attach to the ceiling surface, because of the force exerted by the Coanda effect. However, air jets detach from the ceiling if the airstream’s buoyancy forces are greater than the inertia of the moving airstream. With separation, a cold jet may enter the occupied space, and can result in thermal discomfort. The thermal discomfort is caused by two factors: the cold draft caused by the separated jet in the occupied space, and areas of the room not reached by the separated jet. The separation distance parameter xs is the distance from the diffuser at which a jet separates from the ceiling. Separation distance correlates with outlet jet conditions (Kirkpatrick and Elleson 1996). Separation distance depends on the velocity constant Kc , outlet temperature, flow rate, and static pressure drop. For slot and round diffusers, xs = (48.04)Cs Kc 1/2 (T/T )–1/2 Qo 1/4 P 3/8
(13)
where xs Cs Kc T T Qo P
= = = = = = =
jet separation distance, m separation coefficient, 1.2 centerline velocity constant room-jet temperature difference, K average absolute room temperature, K outlet flow rate, m3/s diffuser static pressure drop, Pa
Attached jets travel at a higher velocity and entrain less air than a free jet. Values of centerline velocity constant Kc are approximately those for a free jet multiplied by 2 . When a jet is discharged parallel to but at some distance from a solid surface (wall, ceiling, or floor), its expansion in the direction of the surface is reduced, and entrained air must be obtained by recirculation from the jet instead of from ambient air (McElroy 1943; Nottage et al. 1952b; Zhang et al. 1990). The restriction to entrainment caused by the solid surface induces the Coanda effect, which makes the jet attach to a surface after it leaves the diffuser outlet. The jet then remains attached to the surface for some distance before separating again. In nonisothermal cases, the jet’s trajectory is determined by the balance between thermal buoyancy and the Coanda effect, which depends on jet momentum and distance between the jet exit and solid surface. The behavior of such nonisothermal surface jets has been studied by Kirkpatrick et al. (1991), Oakes (1987), Wilson et al. (1970), and Zhang et al. (1990), each addressing different factors. More systematic study of these jets in room ventilation
flows is needed to provide reliable guidelines for designing air distribution systems.
Air Curtain Units Non-recirculating air curtain units operate in zones 1 to 3 where velocity degradation is at a minimum. The air curtain unit is designed such that the jet strikes the floor, comparable, surface or another jet in zone 3 at a minimum of 2 m/s, to generate a stable split to resist minimal thermal and pressure differentials. Recirculating air curtain units also operate in zones 1 to 3, where velocity degradation is at a minimum. The target distance is designed for the jet to be captured by the low-pressure return and maintain a minimum of 3 m/s velocity while in zone 3, to create a stable barrier to resist minimal thermal and pressure differentials.
Multiple Jets Twin parallel air jets act independently until they interfere. The point of interference and its distance from outlets vary with the distance between outlets. From outlets to the point of interference, maximum velocity, as for a single jet, is on the centerline of each jet. After interference, velocity on a line midway between and parallel to the two jet centerlines increases until it equals jet centerline velocity. From this point, maximum velocity of the combined jet stream is on the midway line, and the profile seems to emanate from a single outlet of twice the area of one of the two outlets.
Air Movement in Occupied Zone Zhang et al. (1990) found that, for a given heat load and room air supply rate, air velocity in the occupied zone increases when outlet discharge velocity increases. Therefore, the design supply air velocity should be high enough to maintain the jet traveling in the desired direction, to ensure adequate mixing before it reaches the occupied zone. Excessively high outlet air velocity produces high air velocities in the occupied zone and may result in thermal discomfort. Air turbulence in a room is mainly produced at the diffuser jet region by interaction of supply air with room air and with solid surfaces in the vicinity. It is then transported to other parts of the room, including the occupied zone (Zhang et al. 1992). Air in the occupied zone usually contains very small amounts of turbulent kinetic energy compared to the jet region. Because turbulence may cause thermal discomfort (Fanger et al. 1988), air distribution systems should be designed to avoid air turbulence in the occupied zone (except in specialized applications such as task ambient or spot-conditioning systems). Thermal Plumes. As a thermal plume rises because of natural convection above a heat source, it entrains surrounding air and therefore increases in size and volume, and decreases in velocity (Figure 14). The maximum height to which a plume rises depends primarily on the heat source’s strength (relative to the air turbulence surrounding the heat source), and secondarily on stratification in the room (which decreases the rising plume’s buoyancy). The stratified zone has little or no recirculation. In fully stratified and partially mixed applications, cool supply air introduced at or near the floor gradually flows across the lower level of the space. In the case of fully stratified applications (e.g., TDV systems), this layer is typically 100 to 150 mm thick. In partially mixed systems, this layer typically ranges from 0.3 to 2.4 m thick, depending on the upward vertical projection of the outlets’ supply air jets. It is drawn horizontally toward convective heat sources located within or close to it, where it is entrained upward by the associated heat plume. Partially mixed systems are characterized by relatively well-mixed conditions from the floor up to the height where their supply air jet velocities decay to 0.25 m/s or less. This area is referred to as the lower mixed zone. These plumes expand and rise until they encounter equally warm air in the upper
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20.8
2017 ASHRAE Handbook—Fundamentals (SI) 4. Ac Ao AR Ar c Cd cR g H Ho Kc2 Kc3 Lo
= = = = = = = = = = = = =
P Qo Qx r
= = = = =
r0.5V
Thermal Plume from Point Source
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Fig. 14
= = = = = = = = = = = = = = = = Xattached = Xfree = XH = XVT = Rfa SH T T TA TE TF TH TO TS V Vc Vo VT Vx X
SYMBOLS
measured gross (core) area of outlet, m2 core area or neck area, m2 cross-sectional area of confined space normal to jet, m2 Archimedes number [Equation (11)] pollutant concentration discharge coefficient (usually between 0.65 and 0.90) concentration of pollutant at return grille near ceiling level gravitational acceleration rate, m/s2 height or width of slot [Equation (2)], or of room width of jet at outlet or at vena contracta or width of slot, m centerline velocity constant in zone 2 centerline velocity constant in zone 3 length scale of diffuser outlet equal to hydraulic diameter of outlet, m diffuser static pressure drop, Pa discharge from outlet, m3/s total volumetric flow rate at distance X from face of outlet, m3/s radial distance of point under consideration from centerline of jet radial distance in same cross-sectional plane from axis to point where velocity is one-half centerline velocity (i.e., V = 0.5Vx) ratio of free area to gross (core) area stratification height average absolute room temperature, K room/jet temperature difference, K temperature of ambient air, °C temperature at ceiling, °C temperature near floor, °C temperature at given height, °C initial temperature of jet, °C supply temperature, °C actual velocity at point being considered nominal velocity of discharge based on core area, m/s initial air velocity of jet, m/s terminal velocity, m/s centerline velocity, m/s distance from face of outlet to location of centerline velocity VX, m throw distance of attached jet, m throw distance of free jet, m throw height from floor outlet, m distance to given terminal velocity, m
REFERENCES Fig. 15 Schematic Diagram of Major Flow Elements in Room with Displacement Ventilation
ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore.
regions of the space. The upper mixed zone above the stratification height is characterized by low-velocity recirculation, which produces a fairly well-mixed layer of warm air with greater contaminant concentration than that in the lower levels of the space. Typically, warmer, more polluted air will not reenter the stratified zone. This principle is the basis for the improved ventilation effectiveness and heat removal efficiency of TDV systems. In some situations (e.g., morning start-up, winter), there are also sources of cooling in the space, such as cold perimeter windows. The resulting cold downdraft may transport some air from the upper zone back down to the stratified zone. Figure 15 shows basic elements in a simplified schematic of a TDV system. In the figure, q0 represents the supply airflow into the room from a low sidewall diffuser, and q1, q2, and q3 are the upward-moving airflows in thermal plumes that form above heat sources. In this simplified configuration, the stratification height occurs at a height SH, where the net upward-moving flow q1 + q2 + q3 = q0. An important objective in designing and operating a TDV system is to maintain stratification above the occupied zone.
ASHRAE. 2013. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2013. ASHRAE. 2013. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2013. ASHRAE. 2011. Method of testing for rating the performance of air outlets and inlets. ANSI/ASHRAE Standard 70-2006 (RA 2011). ASHRAE. 2010. Standard 62.1 user’s manual. ASHRAE. 2013. UFAD guide: Design, construction and operation of underfloor air distribution systems. Baturin, V.V. 1972. Fundamentals of industrial ventilation, 3rd ed. Translated by O.M. Blunn. Pergamon Press, New York. Chen, Q., and L. Glicksman. 2003. System performance evaluation and design guidelines for displacement ventilation. ASHRAE. Christianson, L.L., ed. 1989. Building systems: Room air and air contaminant distribution. ASHRAE. Fanger, P.O., A.K. Melikov, H. Hanzawa, and J. Ring. 1988. Air turbulence and sensation of draft. Energy and Buildings 12:21-39. Helander, L., and C.V. Jakowatz. 1948. Downward projection of heated air. ASHVE Transactions 54:71. Helander, L., S.M. Yen, and R.E. Crank. 1953. Maximum downward travel of heated jets from standard long radius ASME nozzles. ASHVE Transactions 59:241.
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Licensed for single user. © 2017 ASHRAE, Inc.
Space Air Diffusion Helander, L., S.M. Yen, and L.B. Knee. 1954. Characteristics of downward jets of heated air from a vertical delivery discharge unit heater. ASHVE Transactions 60:359. Helander, L., S.M. Yen, and W. Tripp. 1957. Outlet characteristics that affect the downthrow of heated air jets. ASHAE Transactions 63:255. Kirkpatrick, A., and J. Elleson. 1996. Design guide for cold air distribution systems. ASHRAE. Kirkpatrick, A., T. Malmstrom, P. Miller, and V. Hassani. 1991. Use of low temperature air for cooling of buildings. Proceedings of Building Simulation. Knaak, R. 1957. Velocities and temperatures on axis of downward heated jet from 4-inch long-radius ASME nozzle. ASHAE Transactions 63:527. Koestel, A. 1954. Computing temperatures and velocities in vertical jets of hot or cold air. ASHVE Transactions 60:385. Koestel, A. 1955. Paths of horizontally projected heated and chilled air jets. ASHAE Transactions 61:213. Koestel, A., P. Hermann, and G.L. Tuve. 1949. Air streams from perforated panels. ASHVE Transactions 55:283. Li, Z., J.S. Zhang, A.M. Zhivov, and L.L. Christianson. 1993. Characteristics of diffuser air jets and airflow in the occupied regions of mechanically ventilated rooms: A literature review. ASHRAE Transactions 99(1): 1119-1127. McElroy, G.E. 1943. Air flow at discharge of fan-pipe lines in mines. U.S. Bureau of Mines Report of Investigations 19. Murakami, S. 1992. New scales for ventilation efficiency and their application based on numerical simulation of room airflow. International Symposium on Room Air Convection and Ventilation Effectiveness. Nottage, H.B., J.G. Slaby, and W.P. Gojsza. 1952a. Outlet turbulence intensity as a factor in isothermal-jet flow. ASHVE Transactions 58:343. Nottage, H.B., J.G. Slaby, and W.P. Gojsza. 1952b. Isothermal ventilation jet fundamentals. ASHVE Transactions 58:107. Oakes, W.C. 1987. Experimental investigation of Coanda jet. M.S. thesis, Michigan State University, East Lansing. Rock, B.A., and D. Zhu. 2002. Designer’s guide to ceiling-based air diffusion (RP-1065). ASHRAE. Tuve, G.L. 1953. Air velocities in ventilating jets. ASHVE Transactions 59:261. USGBC. 2013. LEED v4 for building design and construction. U.S. Green Building Council, Washington, D.C. www.usgbc.org/resources /leed-reference-guide-building-design-and-construction. Wilson, J.D., M.L. Esmay, and S. Persson. 1970. Wall-jet velocity and temperature profiles resulting from a ventilation inlet. ASAE Transactions. Yen, S.M., L. Helander, and L.B. Knee. 1956. Characteristics of downward jets from a vertical discharge unit heater. ASHAE Transactions 62:123. Zhang, J.S., L.L. Christianson, and G.L. Riskowski. 1990. Regional airflow characteristics in a mechanically ventilated room under nonisothermal conditions. ASHRAE Transactions 96(1):751-759. Zhang, J.S., L.L. Christianson, G.J. Wu, and G.L. Riskowski. 1992. Detailed measurements of room air distribution for evaluating numerical simulation models. ASHRAE Transactions 98(1):58-65.
BIBLIOGRAPHY Arens, E.A., F. Bauman, L. Johnston, and H. Zhang. 1991. Testing of localized ventilation systems in a new controlled environment chamber. Indoor Air 3:263-281. Arens, E., S. Turner, H. Zhang, and G. Paliaga. 2009. Moving air for comfort. ASHRAE Journal 51(5):18-29. ASHRAE. 2013. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2013. ASHRAE. 2013. Method of testing for room air diffusion. ANSI/ASHRAE Standard 113-2013. ASHRAE. 2002. Measuring air-change effectiveness. ANSI/ASHRAE Standard 129-1997 (RA 2002). Ball, H.D., R.G. Nevins, and H.E. Straub. 1971. Thermal analysis of heat removal troffers. ASHRAE Transactions 77(2). Bauman, F., and E. Arens. 1996. Task/ambient conditioning systems: Engineering and application guidelines. Center for Environmental Design Research, University of California, Berkeley. Bauman, F., P. Pecora, and T. Webster. 1999. How low can you go? Air flow performance of low-height underfloor plenums. Center for the Built Environment, University of California, Berkeley.
20.9 Bauman, F.S., L.P. Johnston, H. Zhang, and E.A. Arens. 1991. Performance testing of a floor-based, occupant-controlled office ventilation system. ASHRAE Transactions 97(1):553-565. Bauman, F.S., H. Zhang, E. Arens, and C. Benton. 1993. Localized comfort control with a desktop task conditioning system: Laboratory and field measurements. ASHRAE Transactions 99(2):733-749. Bauman, F.S., E.A. Arens, S. Tanabe, H. Zhang, and A. Baharlo. 1995. Testing and optimizing the performance of a floor-based task conditioning system. Energy and Buildings 22(3):173-186. Bauman, F.S., T.G. Carter, A.V. Baughman, and E.A. Arens. 1998. Field study of the impact of a desktop task/ambient conditioning system in office buildings. ASHRAE Transactions 104(1):1153-1171. Chen, Q., and L. Glicksman. 1999. Performance evaluation and development of design guidelines for displacement ventilation (RP-949). ASHRAE Research Project, Final Report. de Dear, R.J., and G.S. Brager. 1999. Developing an adaptive model of thermal comfort and preference. ASHRAE Transactions 104(1A):145-167. Faulkner, D., W.J. Fisk, and D.P. Sullivan. 1993. Indoor air flow and pollutant removal in a room with desktop ventilation. ASHRAE Transactions 99(2):750-758. Faulkner, D., W.J. Fisk, D.P. Sullivan, and D.P. Wyon. 1999. Ventilation efficiencies of task/ambient conditioning systems with desk-mounted air supplies. Proceedings of Indoor Air ’99, Edinburgh, Scotland, 8-13 August. Fisk, W.J., D. Faulkner, D. Pih, P. McNeel, F. Bauman, and E. Arens. 1991. Indoor air flow and pollutant removal in a room with task ventilation. Indoor Air 3:247-262. Hanzawa, H., and Y. Nagasawa. 1990. Thermal comfort with underfloor airconditioning systems. ASHRAE Transactions 96(2). Hart, G.H., and D. Int-Hout. 1980. The performance of a continuous linear diffuser in the perimeter zone of an office environment. ASHRAE Transactions 86(2). Hart, G.H., and D. Int-Hout. 1981. The performance of a continuous linear diffuser in the interior zone of an open office environment. ASHRAE Transactions 87(2). Heiselberg, P., and M. Sandberg. 1990. Convection from a slender cylinder in a ventilated room. Proceedings of ROOMVENT ’90, Oslo. Houghton, D. 1995. Turning air conditioning on its head: Underfloor air distribution offers flexibility, comfort, and efficiency. E Source TU-95-8. E Source, Inc., Boulder, CO. Houghten, F.C., C. Gutberlet, and E. Witkowski. 1938. Draft temperatures and velocities in relation to skin temperatures and feelings of warmth. ASHVE Transactions 44:289. Howe, M., D. Holland, and A. Livchak. 2003. Displacement ventilation— Smart way to deal with increased heat gains in the telecommunication equipment room. ASHRAE Transactions 109(1):323-327. Int-Hout, D. 1981. Measurement of room air diffusion in actual office environments to predict occupant thermal comfort. ASHRAE Transactions 87(2). Int-Hout, D. 2007. Overhead heating: Revisiting a lost art. ASHRAE Journal 49(3):56-61. Jackman, P.J. 1991. Displacement ventilation. CIBSE National Conference. Chartered Institution of Building Services Engineers, London. Jackman, P.J., and P.A. Appleby. 1990. Displacement flow ventilation. BSRIA Project Report. Building Services Research and Information Association, Berkshire, U.K. Kegel, B., and U.W. Schulz. 1989. Displacement ventilation for office buildings. Proceedings of the 10th AIVC Conference, Helsinki. Koestel, A. 1957. Jet velocities from radial flow outlets. ASHAE Transactions 63:505. Koestel, A., and J.B. Austin, Jr. 1956. Air velocities in two parallel ventilating jets. ASHAE Transactions 62:425. Koestel, A., and G.L. Tuve. 1955. Performance and evaluation of room air distribution systems. ASHAE Transactions 61:533. Koestel, A., P. Hermann, and G.L. Tuve. 1950. Comparative study of ventilating jets from various types of outlets. ASHVE Transactions 56:459. Livchak, A., and D. Nall. 2001. Displacement ventilation—Application for hot and humid climate. Proceedings of CLIMA 2000, Napoli. Loudermilk, K. 1999. Underfloor air distribution solutions for open office applications. ASHRAE Transactions 105(1):605-613. Lorch, F.A., and H.E. Straub. 1983. Performance of overhead slot diffusers with simulated heating and cooling conditions. ASHRAE Transactions 89(1).
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2017 ASHRAE Handbook—Fundamentals (SI)
Matsunawa, K., H. Iizuka, and S. Tanabe. 1995. Development and application of an underfloor air-conditioning system with improved outlets for a “smart” building in Tokyo. ASHRAE Transactions 101(2):887-901. Mattsson, M. 2000. A note on the thermal comfort in displacement ventilated classrooms. ROOMVENT 2000, Proceedings of the 7th International Conference on Air Distribution in Rooms. McCarry, B.T. 1995. Underfloor air distribution systems: Benefits and when to use the system in building design. ASHRAE Transactions 101(2):902911. McCarry, B.T. 1998. Innovative underfloor system. ASHRAE Journal 40(3). Melikov, A.K., and J.B. Nielsen. 1989. Local thermal discomfort due to draft and vertical temperature difference in rooms with displacement ventilation. ASHRAE Transactions 95(2):1050-1057. Miller, P.L. 1971. Room air distribution performance of four selected outlets. ASHRAE Transactions 77(2):194. Miller, P.L. 1979. Design of room air diffusion systems using the air diffusion performance index (ADPI). ASHRAE Journal 10:85. Miller, P.L. 1989. Descriptive methods. In Building systems: Room air and air contaminant distribution, L.L. Christianson, ed. ASHRAE. Miller, P.L., and R.T. Nash. 1971. A further analysis of room air distribution performance. ASHRAE Transactions 77(2):205. Miller, P.L., and R.G. Nevins. 1969. Room air distribution with an air distributing ceiling—Part II. ASHRAE Transactions 75:118. Miller, P.L., and R.G. Nevins. 1970. Room air distribution performance of ventilating ceilings and cone-type circular ceiling diffusers. ASHRAE Transactions 76(1):186. Miller, P.L., and R.G. Nevins. 1972. An analysis of the performance of room air distribution systems. ASHRAE Transactions 78(2):191. Nelson, D.W., and G.E. Smedberg. 1943. Performance of side outlets on horizontal ducts. ASHVE Transactions 49:58. Nelson, D.W., H. Krans, and A.F. Tuthill. 1940. The performance of stack heads. ASHVE Transactions 46:205. Nelson, D.W., D.H. Lamb, and G.E. Smedberg. 1942. Performance of stack heads equipped with grilles. ASHVE Transactions 48:279. Nevins, R.G., and P.L. Miller. 1972. Analysis, evaluation and comparison of room air distribution performance. ASHRAE Transactions 78(2):235. Nevins, R.G., and E.D. Ward. 1968. Room air distribution with an air distributing ceiling. ASHRAE Transactions 74:VI.2.1. Nielsen, P.V. 1996. Temperature distribution in a displacement ventilated room. ROOMVENT 1996, Proceedings of the 5th International Conference on Air Distribution in Rooms, Yokohama. Poz, M.Y. 1991. Theoretical investigation and practical applications of nonisothermal jets for the rooms ventilating. Current East/West HVAC Developments. IEI/CIBSE/ABOK Joint Conference. Reinmann, J.J., A. Koestel, and G.L. Tuve. 1959. Evaluation of three room air distribution systems for summer cooling. ASHRAE Transactions 65:717. Rock, B.A. 2006. Ventilation for environmental tobacco smoke—Controlling ETS irritants where smoking is allowed. ASHRAE and Elsevier. Rousseau, W.H. 1983. Perimeter air diffusion performance index tests for heating with a ceiling slot diffuser. ASHRAE Transactions 89(1). Rydberg, J., and P. Norback. 1949. Air distribution and draft. ASHVE Transactions 55:225.
Scaret, E. 1985. Ventilation by displacement: Characterization and design implications. Elsevier Science, New York. Seppanen, O.A., W.J. Fisk, J. Eto, and D.T. Grimsrud. 1989. Comparison of conventional mixing and displacement air-conditioning and ventilating systems in U.S. commercial buildings. ASHRAE Transactions 95(2): 1028-1040. Sandberg, M., and C. Blomqvist. 1989. Displacement ventilation in office rooms. ASHRAE Transactions 95(2):1041-1049. Shilkrot, E., and A. Zhivov. 1992. Room ventilation with designed vertical air temperature stratification. ROOMVENT ’92, Proceedings of the 3rd International Conference on Engineering Aero- and Thermodynamics of Ventilated Rooms. Shute, R.W. 1992. Integrating access floor plenums for HVAC air distribution. ASHRAE Journal 34(10). Shute, R.W. 1995. Integrated access floor HVAC: Lessons learned. ASHRAE Transactions 101(2):877-886. Skistad, H. 1994. Displacement ventilation. Research Studies Press, John Wiley & Sons, West Sussex, U.K. Skistad, H., E. Mundt, P. Nielsen, K. Hagstrom, and J. Railio. 2002. Displacement ventilation in non-industrial premises. REHVA Guidebook 1. Sodec, F., and R. Craig. 1990. The underfloor air supply system—The European experience. ASHRAE Transactions 96(2). Spoormaker, H.J. 1990. Low-pressure underfloor HVAC system. ASHRAE Transactions 96(2). Straub, H.E., and M.M. Chen. 1957. Distribution of air within a room for year-round air conditioning—Part II. University of Illinois Engineering Experiment Station Bulletin 442. Straub, H.E., S.F. Gilman, and S. Konzo. 1956. Distribution of air within a room for year-round air conditioning—Part I. University of Illinois Engineering Experiment Station Bulletin 435. Stymne, H., M. Sandberg, and M. Mattsson. 1991. Dispersion pattern of contaminants in a displacement ventilation room—Implications for demand control. Proceedings of the 12th Air Movement and Ventilation Control Within Buildings, Ottawa. Svensson, A.G.L. 1989. Nordic experiences of displacement ventilation systems. ASHRAE Transactions 95(2):1013-1017. Tan, H., T. Murata, K. Aoki, and T. Kurabuchi. 1998. Cooled ceilings/displacement ventilation hybrid air conditioning system—Design criteria. Proceedings of ROOMVENT ’98, Stockholm. Tanabe, S., and K. Kimura. 1996. Comparisons of ventilation performance and thermal comfort among displacement, underfloor and ceiling based air distribution systems by experiments in a real sized office chamber. ROOMVENT ’96, Proceedings of the 5th International Conference on Air Distribution in Rooms. Tse, W.L., and A.T.P. So. 2006. The importance of human productivity to air-conditioned control in office environments. HVAC&R Research (now Science and Technology for the Built Environment) 13:3-21. Tsuzuki, K., E.A. Arens, F.S. Bauman, and D.P. Wyon. 1999. Individual thermal comfort control with desk-mounted and floor-mounted task/ ambient conditioning (TAC) systems. Proceedings of Indoor Air ’99, Edinburgh, vol. 2, pp. 368-373.
Related Commercial Resources
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Related Commercial Resources CHAPTER 21
DUCT DESIGN BERNOULLI EQUATION ....................................................... 21.1 Head and Pressure................................................................... 21.2 SYSTEM ANALYSIS................................................................. 21.2 Pressure Changes in System .................................................... 21.5 FLUID RESISTANCE .............................................................. 21.6 Friction Losses......................................................................... 21.6 Dynamic Losses ....................................................................... 21.8 Ductwork Sectional Losses .................................................... 21.13
OMMERCIAL, industrial, and residential air duct system design must consider (1) space availability, (2) noise levels, (3) air leakage, (4) balancing, (5) fire and smoke control, (6) initial investment cost, and (7) system operating cost. Deficiencies in duct design can result in systems that operate incorrectly or are expensive (increased energy) to own and operate. Poor design or lack of system sealing can produce inadequate airflow rates at the terminals, leading to discomfort, loss of productivity, and even adverse health effects. Lack of sound attenuation may lead to objectionable noise levels. Proper duct insulation eliminates excessive heat gain or loss. In this chapter, system design and calculation of a system’s frictional and dynamic resistance (total pressure) to airflow are considered. Chapter 19 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment examines duct construction and presents construction standards for residential, commercial, and industrial HVAC and exhaust systems. For design guidance specific to residential systems, refer to Manual D by ACCA (2014).
FAN/SYSTEM INTERFACE................................................... MECHANICAL EQUIPMENT ROOMS............................................................................... DUCT DESIGN ...................................................................... Design Considerations ........................................................... Design Recommendations ...................................................... Design Methods...................................................................... Industrial Exhaust Systems ....................................................
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C
1.
BERNOULLI EQUATION
The Bernoulli equation can be developed by equating the forces on an element of a stream tube in a frictionless fluid flow to the rate of momentum change. On integrating this relationship for steady flow, the following expression (Osborne 1966) results:
2
v p ----- + ------- + gz = constant, N·m/kg 2
(1)
where v p g z
= = = = =
streamline (local) velocity, m/s absolute pressure, Pa (N/m2) density, kg/m3 acceleration caused by gravity, m/s2 elevation, m
Assuming constant fluid density in the system, Equation (1) reduces to 2
p v ----- + --- + gz = constant, N·m/kg 2
(2)
Although Equation (2) was derived for steady, ideal frictionless flow along a stream tube, it can be extended to analyze flow through ducts in real systems. In terms of pressure, the relationship for fluid resistance between two sections is The preparation of this chapter is assigned to TC 5.2, Duct Design.
21.1 Copyright © 2017, ASHRAE
2
21.13 21.15 21.15 21.15 21.21 21.22 21.28
2
1V 1 2V 2 ------------ + p1 + g1z1 = ------------ + p2 + g2z2 + pt, 1–2 2 2
(3)
where V = average duct velocity, m/s pt,1–2 = total pressure loss caused by friction and dynamic losses between sections 1 and 2, Pa
In Equation (3), V (section average velocity) replaces v (streamline velocity) because experimentally determined loss coefficients allow for errors in calculating v 2/2 (velocity pressure) across streamlines. On the left side of Equation (3), add and subtract pz1; on the right side, add and subtract pz2, where pz1 and pz2 are the values of atmospheric air at heights z1 and z2. Thus, 2
1V 1 ------------ + p 1 + p z1 – p z1 + g 1 z 1 2 2
(4)
2V 2 = ------------ + p 2 + p z2 – p z2 + g 2 z 2 + p t, 1-2 2 Atmospheric pressure at any elevation ( pz1 and pz2) expressed in terms of the atmospheric pressure pa at the same datum elevation is given by (5) pz1 = pa – ga z1 pz2 = pa – ga z2
(6)
Substituting Equations (5) and (6) into Equation (4) and simplifying yields the total pressure change between sections 1 and 2. Assume no temperature change between sections 1 and 2 (no heat exchanger within the section); therefore, 1 = 2. When a heat exchanger is located in the section, the average of the inlet and outlet temperatures is generally used. Let = 1 = 2, and ( p1 – pz1) and ( p2 – pz2) are gage pressures at elevations z1 and z2. 2 2 V V pt,1–2 = p s ,1 + ---------1 – p s ,2 + ---------2 + g(a – )(z2 – z1) (7a) 2 2 (7b) pt,1–2 = pt + pse
Rearranging Equation (7b) yields pt = pt,1-2 – pse where ps,1 ps,2 V1 V2 a pse
= = = = = = =
static pressure, gage at elevation z1, Pa static pressure, gage at elevation z2, Pa average velocity at section 1, m/s average velocity at section 2, m/s density of ambient air, kg/m3 density of air or gas in duct, kg/m3 thermal gravity effect, Pa
(7c)
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21.2
2017 ASHRAE Handbook—Fundamentals (SI)
pt = total pressure change between sections 1 and 2, Pa pt,1-2 = total pressure loss caused by friction and dynamic losses between sections 1 and 2, Pa
1.1
pt = pf + i
The terms head and pressure are often used interchangeably; however, head is the height of a fluid column supported by fluid flow, whereas pressure is the normal force per unit area. For liquids, it is convenient to measure head in terms of the flowing fluid. With a gas or air, however, it is customary to measure pressure exerted by the gas on a column of liquid.
Static Pressure The term p/g is static head; p is static pressure.
Velocity Pressure The term V 2/2g refers to velocity head, and V 2/2gc refers to velocity pressure. Although velocity head is independent of fluid density, velocity pressure [Equation (8)] is not.
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pij + pik – pse m
n
j =1
k =1
r =1
ir
(8)
where p t = net total pressure change for i sections, Pa i p f = pressure loss caused by friction for i sections, Pa i pij = total pressure loss caused by j fittings, including fan system effect (FSE), for i sections, Pa pik = pressure loss caused by k equipment for i sections, Pa pse = thermal gravity effect caused by r stacks for i sections, Pa ir m = number of fittings within i sections n = number of equipment within i sections = number of stacks within i sections nup = number of duct sections upstream of fan (exhaust/return air subsystems) ndn = number of duct sections downstream of fan (supply air subsystems)
From Equation (7), the thermal gravity effect for each nonhorizontal duct with a density other than that of ambient air is determined by the following equation: pse = g(a – )(z2 – z1)
where pv = velocity pressure, Pa V = fluid mean velocity, m/s
(9)
Velocity is calculated by V = Q/1000A
(14)
where
For air at standard conditions (1.204 kg/m3), Equation (8) becomes pv = 0.602 V 2
(13)
for i = 1, 2, , n up + n dn
HEAD AND PRESSURE
pv = V 2/2
i
(10)
where
pse z1 and z2 a g
= thermal gravity effect, Pa = elevation from datum in direction of airflow (Figure 1), m = density of ambient air, kg/m3 = density of air or gas within duct, kg/m3 = 9.81 = gravitational acceleration, m/s2
Example 1. For Figure 1, calculate the thermal gravity effect for two cases: (a) air cooled to –34°C, and (b) air heated to 540°C. Density of air at –34°C is 1.477 kg/m3 and at 540°C is 0.434 kg/m3. Density of ambient air is 1.204 kg/m3. Stack height is 15 m. Solution:
Q = airflow rate, L/s A = cross-sectional area of duct, m2
pse = 9.81(a – )z (a) For a (Figure 1A),
Total Pressure
pse = 9.81(1.204 – 1.477)15 = –40 Pa
Total pressure is the sum of static pressure and velocity pressure: pt = ps + V 2/2
(11)
pt = ps + pv
(12)
(b) For a (Figure 1B), pse = 9.81(1.204 – 0.434)15 = +113 Pa
or
where pt = total pressure, Pa ps = static pressure, Pa
Pressure Measurement The range, precision, and limitations of instruments for measuring pressure and velocity are discussed in Chapter 36. The manometer is a simple and useful means for measuring partial vacuum and low pressure. Static, velocity, and total pressures in a duct system relative to ambient space pressure can be measured with a pitot tube connected to a manometer. Pitot tube construction and locations for traversing round and rectangular ducts are presented in Chapter 37.
2. SYSTEM ANALYSIS The total pressure change caused by friction, fittings, equipment, and net thermal gravity effect for each section of a duct system is calculated by the following equation:
Fig. 1
Thermal Gravity Effect for Example 1
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Duct Design
21.3
Example 2. Calculate the thermal gravity effect for the two-stack system shown in Figure 2, where the air is 120°C and stack heights are 15 and 30 m. Density of 120°C air is 0.898 kg/m3; ambient air is 1.204 kg/m3. Solution: pse = 9.81(a – )(z2 – z1) = 9.81(1.204 – 0.898)(30 – 15) = 45 Pa
For the system shown in Figure 3, the direction of air movement created by the thermal gravity effect depends on the initiating force
(e.g., fans, wind, opening and closing doors, turning equipment on and off). If for any reason air starts to enter the left stack (Figure 3A), it creates a buoyancy effect in the right stack. On the other hand, if flow starts to enter the right stack (Figure 3B), it creates a buoyancy effect in the left stack. In both cases, the produced thermal gravity effect is stable and depends on stack height and magnitude of heating. The starting direction of flow is important when using natural convection for ventilation. To determine the fan total pressure requirement for a system, use the following equation: Pt =
pt + pt i
iF up
i
for i = 1, 2, …, nup + ndn
(15)
iF dn
where Fup and Fdn=sets of duct sections upstream and downstream of fan Pt = fan total pressure, Pa
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= symbol that ties duct sections into system paths from exhaust/ return air terminals to supply terminals
Fig. 2 Multiple Stacks for Example 2
Figure 4 shows the use of Equation (15). This system has three supply and two return terminals consisting of nine sections connected in six paths: 1-3-4-9-7-5, 1-3-4-9-7-6, 1-3-4-9-8, 2-4-9-7-5, 2-4-97-6, and 2-4-9-8. Sections 1 and 3 are unequal area; thus, they are assigned separate numbers in accordance with the rules for identifying sections (see step 5 in the section on HVAC Duct Design Procedures). To determine the fan pressure requirement, apply the following six equations, derived from Equation (15). These equations must be satisfied to attain pressure balancing for design airflow. Relying entirely on dampers is not economical and may create objectionable flow-generated noise. P t Pt Pt Pt Pt Pt
= p1 + p3 + p4 + p9 + p7 + p5 = p1 + p3 + p4 + p9 + p7 + p6 = p1 + p3 + p4 + p9 + p8 = p2 + p4 + p9 + p7 + p5
(16)
= p2 + p4 + p9 + p7 + p6 = p2 + p4 + p9 + p8
Example 3. For Figures 5A and 5C, calculate the thermal gravity effect and fan total pressure required when the air is cooled to –34°C. The heat exchanger and ductwork (section 1 to 2) total pressure losses are 170 and 70 Pa respectively. The density of –34°C air is 1.477 kg/m3; ambient air is 1.204 kg/m3. Elevations are 21 and 3 m. Solution: (a) For Figure 5A (downward flow), p se = 9.81 a – z 2 – z 1 = 9.81 1.204 – 1.477 3 – 21 = 48 Pa
Fig. 3 Multiple Stack Analysis
Fig. 4 Illustrative 6-Path, 9-Section System
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21.4
2017 ASHRAE Handbook—Fundamentals (SI) P t = p t,3–2 – p se = 170 + 70 – 48 = 192 Pa
Solution: (a) For Figure 5B (downward flow), p se = 9.81 a – z 2 – z 1 = 9.81 1.204 – 0.898 3 – 21
(b) For Figure 5C (upward flow),
= – 54 Pa
p se = 9.81 a – z 2 – z 1
P t = p t,3–2 – p se
= 9.81 1.204 – 1.477 21 – 3
= 170 + 70 – – 54
= – 48 Pa
= 294 Pa (b) For Figure 5D (upward flow),
P t = p t,3-2 – p se = 170 + 70 – – 48 = 288 Pa
p se = 9.81 a – z 2 – z 1 = 9.81 1.204 – 0.898 21 – 3 = 54 Pa P t = p t,3-2 – p se
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Example 4. For Figures 5B and 5D, calculate the thermal gravity effect and fan total pressure required when air is heated to 120°C. Heat exchanger and ductwork (section 1 to 2) total pressure losses are 170 and 70 Pa respectively. Density of 120°C air is 0.898 kg/m3; ambient air is 1.204 kg/m3. Elevations are 21 and 3 m.
Fig. 5 Single Stack with Fan for Examples 3 and 4
= 170 + 70 – 54 = 186 Pa
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Duct Design
21.5
Example 5. Calculate the thermal gravity effect for each section of the system shown in Figure 6, and the systems’ net thermal gravity effect. Density of ambient air is 1.204 kg/m3, and the lengths are as follows: z1 = 15 m, z2 = 27 m, z4 = 30 m, z5 = 8 m, and z9 = 60 m. Pressure required at section 3 is 25 Pa. Write the equation to determine the fan total pressure requirement. Solution: The following table summarizes the thermal gravity effect for each section of the system as calculated by Equation (14). The net thermal gravity effect for the system is 118 Pa. To select a fan, use the following equation: P t = 25 + p t ,1-7 + p t ,8-9 – p se = 25 + p t ,1-7
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+ p t ,8-9 – 118 = p t ,1-7 + p t ,8-9 – 93
Fig. 6 Triple Stack System for Example 5
Path (x–x)
Temp., °C
, kg/m3
pse , z Pa (zx – zx), (a – xx), kg/m3 m [Eq. (14)]
1-2 3-4 4-5 6-7 8-9
815 540 540 120 120
0.324 0.434 0.434 0.898 0.898
(27 – 15) 0 (8 – 30) 0 (60 – 0)
+0.880 +0.770 +0.770 +0.306 +0.306
+104 0 –166 0 +180
Net Thermal Gravity Effect
118
2.1 PRESSURE CHANGES IN SYSTEM Figure 7 shows total and static pressure changes in a fan/duct system consisting of a fan with both supply and return air ductwork. Also shown are total and static pressure gradients referenced to atmospheric pressure. For all constant-area sections, total and static pressure losses are equal. At diverging transitions, velocity pressure decreases, absolute total pressure decreases, and absolute static pressure can increase. The static pressure increase at these sections is known as static regain. At converging transitions, velocity pressure increases in the direction of airflow, and absolute total and absolute static pressures decrease. At the exit, total pressure loss depends on the shape of the fitting and the flow characteristics. Exit loss coefficients Co can be greater than, less than, or equal to one. Total and static pressure grade lines for the various coefficients are shown in Figure 7. Note that, for a loss coefficient less than one, static pressure upstream of the exit is less than atmospheric pressure (negative). Static pressure just upstream of the discharge fitting can be calculated by subtracting the upstream velocity pressure from the upstream total pressure. At section 1, total pressure loss depends on the shape of the entry. Total pressure immediately downstream of the entrance equals the difference between the upstream pressure, which is zero (atmospheric pressure), and loss through the fitting. Static pressure of ambient air is zero; several diameters downstream, static pressure is negative, equal to the sum of the total pressure (negative) and the velocity pressure (always positive). System resistance to airflow is noted by the total pressure grade line in Figure 7. Sections 3 and 4 include fan system effect pressure
Fig. 7 Pressure Changes During Flow in Ducts
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21.6
2017 ASHRAE Handbook—Fundamentals (SI)
losses. To obtain the fan static pressure requirement for fan selection where fan total pressure is known, use Ps = Pt – pv,o (17) where Ps = fan static pressure, Pa Pt = fan total pressure, Pa pv,o = fan outlet velocity pressure, Pa
3.
FLUID RESISTANCE
Duct system losses are the irreversible transformation of mechanical energy into heat. The two types of losses are (1) friction and (2) dynamic.
3.1
FRICTION LOSSES
Friction losses are caused by fluid viscosity and result from momentum exchange between molecules (in laminar flow) or between individual particles of adjacent fluid layers moving at different velocities (in turbulent flow). Friction losses occur along the entire duct length. Fig. 8 Pressure Loss Correction Factor for Flexible Duct Not Fully Extended
Darcy and Colebrook Equations
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For fluid flow in conduits, friction loss can be calculated by the Darcy equation: 1000 f L V 2 pf = ------------------ --------2 Dh
(18)
where pf f L Dh V
= = = = = =
friction losses in terms of total pressure, Pa friction factor, dimensionless duct length, m hydraulic diameter [Equation (24)], mm velocity, m/s density, kg/m3
In the region of laminar flow (Reynolds numbers less than 2300), the friction factor is a function of Reynolds number only. For completely turbulent flow (fully rough), the friction factor depends on duct surface roughness and internal protuberances (e.g., joints). Between the bounding limits of hydraulically smooth behavior (laminar flow) and fully rough behavior is a transitional zone where the friction factor depends on both roughness and Reynolds number. In both the transitional and fully rough (see Moody diagram: Figure 13 in Chapter 3) regions, the friction factor f is calculated by Colebrook’s equation (Colebrook 1938-1939). Because Colebrook’s equation cannot be solved explicitly for f, use iterative techniques (Behls 1971). 1 2.51 ------- = – 2 log -------------- + -------------- f 3.7D h Re f
(19)
where = material absolute roughness factor, mm Re = Reynolds number
Reynolds number (Re) may be calculated by using the following equation. Dh V Re = ----------------(20) 1000 where = kinematic viscosity, m2/s. For standard air and temperature between 4 and 38°C, Re can be calculated by (21) Re = 66.4 DhV
Roughness Factors Roughness factors listed in Table 1, column 3, are recommended for use with Equation (19). For increased calculation accuracy, use an absolute roughness factor from column 2.
Flexible Duct. For fully stretched and compressed flexible duct, use Equation (22) or Figure 8 (Abushakra et al. 2004; Culp 2011), where the multiplier PDCF is based on flexible duct with an absolute roughness = 0.9 mm. The resistance of flexible duct can be calculated using Fitting CD11-2 in the ASHRAE Duct Fitting Database (ASHRAE 2016). Flexible duct should be installed fully extended; its resistance when fully extended is increased approximately 50%. See Example 6 for the increase in resistance of flexible duct fully stretched and compressed relative to rigid spiral duct. For commercial systems, flexible ducts should be • Limited to connections of rigid ducts to diffusers. For diffuser installation suggestions, see Figure 9. The purpose of limiting the offset to D/8 in Figure 9B is to minimize noise generation. Flexible duct should not be installed upstream of variable-air-volume (VAV) boxes. Commentary: The loss coefficient or pressure loss for flexible duct elbows, such as those in Figure 9A, can be obtained from the ASHRAE Duct Fitting Database (ASHRAE 2016), Fitting CD322 (r/D = 1.0) or CD3-23 (r/D = 1.5). Loss coefficients are for fully stretched elbows. • Limited to 2 m maximum, fully stretched. • Installed without any radial compression. Example 6. Compare the total pressure resistance of 250 mm, 1.8 m installed length, galvanized steel spiral and flexible duct, 0% compressed (fully stretched), 4%, 15%, and 30% compressed. Airflow is 470 L/s, air density is 1.204 kg/m3, and absolute roughnesses of spiral round and flexible ducts are 0.12 mm and 0.9 mm. Calculate using the ASHRAE Duct Fitting Database [DFDB; ASHRAE (2016)]. Solution: See Table 2 for results.
with
PDCF = 1 + 0.58 Kc e– 0.00496D
(22)
L FE – L Kc = ------------------ 100 L FE
(23)
where PDCF = pressure drop correction factor Kc = flexible duct compressed, percent D = flexible duct diameter, mm
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Duct Design
21.7
Table 1 Duct Roughness Factors 1
2
3 Absolute Roughness , mm
Duct Type/Material
Range
Drawn tubing (Madison and Elliot 1946)
0.00046
Smooth 0.00046
PVC plastic pipe (Swim 1982) Commercial steel or wrought iron (Moody 1944)
0.009 to 0.046 0.046
Medium smooth 0.046
Aluminum, round, longitudinal seams, crimped slip joints, 0.91 m spacing (Hutchinson 1953)
0.037 to 0.061
Friction chart: Galvanized steel, round, longitudinal seams, variable joints (Vanstone, drawband, welded. Primarily beaded coupling), 1.22 m joint spacing (Griggs et al. 1987) Galvanized steel, spiral seams, 3.05 m joint spacing (Jones 1979) Galvanized steel, spiral seam with 1, 2, and 3 ribs, beaded couplings, 3.66 m joint spacing (Griggs et al. 1987) Galvanized steel, rectangular, various type joints (Vanstone, drawband, welded. Beaded coupling), 1.22 m spacinga (Griggs and Khodabakhsh-Sharifabad 1992)
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Wright Friction Chart: Galvanized steel, round, longitudinal seams, 0.76 m joint spacing, = 0.15 mm
Roughness Category
0.049 to 0.098
Average 0.09
0.061 to 0.12 0.088 to 0.116 0.082 to 0.15
Retained for historical purposes [See Wright (1945) for development of friction chart]
Flexible duct, nonmetallic and wire, fully extended (Abushakra et al. 2004; Culp 2011) Galvanized steel, spiral, corrugated,b Beaded slip couplings, 3.05 m spacing (Kulkarni et al. 2009) Fibrous glass duct, rigid (tentative)c Fibrous glass duct liner, air side with facing material (Swim 1978)
0.09 to 0.9 0.54 to 0.91 — 1.52
Medium rough 0.9
Fibrous glass duct liner, air side spray coated (Swim 1978) Flexible duct, metallic corrugated, fully extended Concrete (Moody 1944)
4.57 1.2 to 2.1 0.30 to 3.0
Rough 3.0
aGriggs
and Khodabakhsh-Sharifabad (1992) showed that values for rectangular duct construction combine effects of surface condition, joint spacing, joint type, and duct construction (cross breaks, etc.), and that the -value range listed is representative. bSpiral seam spacing was 119 mm with two corrugations between seams. Corrugations were 19 mm wide by 6 mm high (semicircle). cSubject duct classified “tentatively medium rough” because no data available.
Table 2
Solution for Example 6
Duct
DFDB Fitting
, mm
Airflow, L/s
Diameter, mm
Velocity, m/s
Compression, %
Flexible
CD11-2
0.9
470
250
9.6
Flexible Flexible Flexible Galvanized steel, spiral
CD11-2 CD11-2 CD11-2 CD11-1
0.9 0.9 0.9 0.12
470 470 470 470
250 250 250 250
9.6 9.6 9.6 9.6
(fully stretched) 4 15 30 NA
*See Equation (22)
Fig. 9
Diffuser Installation Suggestions
PDCF*
pt, Pa
% pt Increased
1.0
11.4
48
1.7 3.5 5.9 NA
19.5 40.0 68.9 7.7
153 419 795 Base
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21.8
2017 ASHRAE Handbook—Fundamentals (SI)
L = installed duct length, m LFE = duct length fully extended, m
Licensed for single user. © 2017 ASHRAE, Inc.
Friction Chart The friction chart (Figure 10) is a plot of the Darcy and Colebrook equations [Equations (18) and (19), respectively), where the absolute roughness is 0.09 mm and the air is standard air (density = 1.204 kg/m3)]. Figure 10 can be used for (1) duct construction/materials categorized as “average” in Table 1, (2) temperature variations of ±15 K from 20°C, (3) elevations to 500 m, and (4) duct pressures from –5 to +5 kPa relative to ambient pressure. These individual variations in temperature, elevation, and duct pressure result in duct losses within ±5% of the standard air friction chart. The friction chart was changed in 1985 from an absolute roughness of 0.15 mm to 0.09 mm based on research by Griggs et al. (1987), who found that the roughness factor is affected by the material surface, joint spacing, and type of joint. The Wright friction chart appeared in the Handbook from 1946 to 1981. This chart was based on an absolute roughness = 0.15 mm, primarily because of the 760 mm joint spacing. In 1985 the friction chart was changed to = 0.09 mm because joint spacing was increasing. For the relative effect of straight duct resistance between charts, see Figure 11. For a 250 mm diameter duct at 10 m/s (491 L/s), the resistance decreased 5 to 6%.
Noncircular Ducts A momentum analysis can relate average wall shear stress to pressure drop per unit length for fully developed turbulent flow in a passage of arbitrary shape but uniform longitudinal cross-sectional area. This analysis leads to the definition of hydraulic diameter: Dh = 4A/P (24) where Dh = hydraulic diameter, mm A = duct area, mm2 P = perimeter of cross section, mm
Although hydraulic diameter is often used to correlate noncircular data, exact solutions for laminar flow in noncircular passages show that this causes some inconsistencies. No exact solutions exist for turbulent flow. Tests over a limited range of turbulent flow indicated that fluid resistance is the same for equal lengths of duct for equal mean velocities of flow if the ducts have the same ratio of crosssectional area to perimeter. From experiments using round, square, and rectangular ducts having essentially the same hydraulic diameter, Huebscher (1948) found that each, for most purposes, had the same flow resistance at equal mean velocities. Tests by Griggs and Khodabakhsh-Sharifabad (1992) also indicated that experimental rectangular duct data for airflow over the range typical of HVAC systems can be correlated satisfactorily using Equation (19) together with hydraulic diameter, particularly when a realistic experimental uncertainty is accepted. These tests support using hydraulic diameter to correlate noncircular duct data. Rectangular Ducts. Huebscher (1948) developed the relationship between rectangular and round ducts that is used to determine size equivalency based on equal flow, resistance, and length. This relationship, Equation (25), is the basis for Table 3. 0.625
1.30 ab De = -------------------------------0.250 a + b
(25)
where De = circular equivalent of rectangular duct for equal length, fluid resistance, and airflow, mm a = length one side of duct, mm b = length adjacent side of duct, mm
To determine equivalent round duct diameter, use Table 3. Equations (18) and (19) must be used to determine pressure loss. Flat Oval Ducts. To convert round ducts to flat oval sizes, use Table 4, which is based on Equation (26) (Heyt and Diaz 1975), the circular equivalent of a flat oval duct for equal airflow, resistance, and length. Equations (18) and (19) must be used to determine friction loss. 0.625
1.55AR De = ---------------------------0.250 P
(26)
where AR is the cross-sectional area of flat oval duct defined as AR = (a2/4) + a(A – a)
(27)
and the perimeter P is calculated by P = a + 2(A – a)
(28)
where P = perimeter of flat oval duct, mm A = major axis of flat oval duct, mm a = minor axis of flat oval duct, mm
3.2 DYNAMIC LOSSES Dynamic losses result from flow disturbances caused by ductmounted equipment and fittings that change flow direction (elbows), area changes (transitions), and converging/diverging junctions. For a detailed discussion of hydraulic networks, consult Idelchik et al. (1994).
Local Loss Coefficients The dimensionless coefficient C is used for fluid resistance because this coefficient has the same value in dynamically similar streams (i.e., streams with geometrically similar stretches, equal Reynolds numbers, and equal values of other criteria necessary for dynamic similarity). The fluid resistance coefficient represents the ratio of total pressure loss to velocity pressure at the referenced cross section: pt pt C = --------------------- = --------(29) 2 pv V 2 where C pt V pv
= = = = =
local loss coefficient, dimensionless total pressure loss, Pa density, kg/m3 velocity, m/s velocity pressure, Pa
For all fittings, except junctions, total pressure loss is calculated by Equation (30): pt = Co pv,o
(30)
where pt = total pressure loss of fitting, Pa Co = local loss coefficient of fitting, dimensionless pv,o = velocity pressure at section o of fitting, Pa
Dynamic loss is based on the actual velocity in the duct, not the velocity in an equivalent circular duct. For the cross section to reference a fitting loss coefficient, see step 5 in the section on HVAC Duct Design Procedures. Where necessary (e.g., unequal-area fittings), convert a loss coefficient from section o to section 1 using Equation (31), where V is the velocity at the respective sections. Co C1 = -----------------------2 V1 Vo
(31)
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Duct Design 21.9
Fig. 10 Friction Chart for Round Duct ( = 1.20 kg/m3 and = 0.09 mm)
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21.10
2017 ASHRAE Handbook—Fundamentals (SI) Table 3 Circular Equivalents of Rectangular Duct for Equal Friction and Airflow* Length of One Side of Rectangular Duct a, mm 100
150
200
250
300
Side Lengths, mm 200 250 300 400 500 600 800 1 000 1 200 1 400 1 600 1 800 2 000 *Table
400
500
600
800
1 000
1 200
656 755 840 914 980 1 041 1 096 1 147
875 976 1 066 1 146 1 219 1 286 1 348
1 093 1 196 1 289 1 373 1 451 1 523
1 312 1 416 1 511 1 598 1 680
Circular Duct Diameter, mm 152 169 183 207
189 210 229 260 287 310
219 244 266 305 337 365 414
273 299 343 381 414 470 517
328 378 420 457 520 574 620
437 488 533 609 674 731 781 827
547 598 687 762 827 886 939 988 1 034
based on Equation (25).
Table 4 Equivalent Flat Oval Duct Dimensions* Licensed for single user. © 2017 ASHRAE, Inc.
Minor Axis a, mm Circular Duct Diameter, mm 125 140 160 180 200 224 250 280 315 355 400 450 500 560 630 710 800 900 1 000 1 120 1 250
70
100
125
150
175
200
250
275
300
325
350
375
400
450
500
550
600
Major Axis A, mm 205 265 360 475
180 235 300 380 490
190 235 290 375 475
200 245 305 385 485 635 840 1 115 1 490
215 — 325 410 525 — — —
240 290 360 — 580 760 995 1 275 1 680
— — 460 — — — —
285 345 425 530 675 845 1 085 1 425
325 395 490 — — — —
375 460 435 570 535 505 700 655 615 580 890 820 765 720 1 150 1 050 970 905 810 1 505 1 370 1 260 1 165 1 025 1 800 1 645 1 515 1 315 2 165 1 985 1 705 2 170
1 170 1 500 1 895 2 455
1 065 1 350 1 690 2 170 1 950 2 795 2 495
*Table based on Equation (26).
For converging and diverging flow junctions, total pressure loss through the straight (main) section is calculated by pt,s = Cs pv, s
(32)
where pt,s = total pressure loss across straight-through section s of junction, Pa Cs = local loss coefficient referenced to s section of junction, dimensionless pv,s = velocity pressure at section s, Pa
For total pressure loss through the branch section, pt, b = Cb pv, b where pt,b = total pressure loss across branch b section of junction, Pa Cb = local loss coefficient referenced to b section of junction, dimensionless
(33)
pv,b = velocity pressure at section b, Pa
The junction of two parallel streams moving at different velocities is characterized by turbulent mixing of the streams, accompanied by pressure losses. In the course of this mixing, particles moving at different velocities exchange momentum, resulting in equalization of the velocity distributions in the common stream. The jet with higher velocity loses part of its kinetic energy by transmitting it to the slower jet. The loss in total pressure before and after mixing is always large and positive for the higher-velocity jet, and increases with an increase in the amount of energy transmitted to the lower-velocity jet. Consequently, the local loss coefficient [Equation (29)] is always positive. Energy stored in the lower-velocity jet can increase because of mixing. The loss in total pressure and the local loss coefficient can, therefore, also have negative values for the lower-velocity jet (Idelchik et al. 1994).
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Duct Design
21.11 Table 5 Duct Fitting Codes
Fitting Function S: Supply E: Exhaust/Return C: Common (supply and return)
Geometry
Category
D: round (Diameter) 1. 2. R: Rectangular 3. 4. F: Flat oval 5. 6. 7. 8.
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9. 10. 11.
Table 6 200 mm VAV Box Data Sequential Number
Entries 1,2,3...n Exits Elbows Transitions Junctions Obstructions Fan and system interactions Duct-mounted equipment Dampers Hoods Straight duct
Inlet Outlet Velocity, Velocity, pv,in, m/s m/s Pa
Airflow, L/s
ps , Pa
1
2
3
4
165 236 330 425
13.2 27.1 53.0 87.8
5.1 7.3 10.2 13.1
2.1 3.0 4.3 5.5
pv,out , Pa
pv , Pa
pt , Pa
5
6
7
8
15.6 31.9 62.5 103.3
2.7 5.6 11.0 18.1
12.9 26.3 51.5 85.1
26.1 53.4 104.5 173.0
ps = static pressure difference pt = total pressure difference at standard air conditions pv,in = inlet velocity pressure at standard air conditions pv,out = outlet velocity pressure at standard air conditions pv = velocity pressure difference at standard air conditions
Fig. 12 VAV Box Loss Coefficient Plot
Fig. 11 Plot Illustrating Relative Resistance of Roughness Categories
Duct Fitting Database Loss coefficients for more than 220 round, flat oval, and rectangular fittings are available in the ASHRAE Duct Fitting Database [DFDB; ASHRAE (2016)]. Also included are the pressure loss for round duct (CD11-1), flexible duct (CD11-2), rectangular duct (CR11-1), and flat oval (CF11-1) duct, as well as the following design tools: • CD11-3, Straight Duct, Round, Velocity Limited • CD11-4, Straight Duct, Round, Friction Rate Constant • CD11-5, Straight Duct, Round, Minimum Velocity Commentary: CD11-3 determines the size of a duct knowing airflow such that the design velocity is not exceeded. CD11-4 is for sizing duct systems by the equal friction method, knowing the design (target) friction rate and airflow. CD4-11 determines the duct size for industrial systems that must maintain a minimum velocity to convey particulates. Example 8 uses CD11-3 in the equal friction and static regain designs, and CD11-4 for the equal friction design. The fittings are numbered (coded) as shown in Table 5. Entries and converging junctions are only in the exhaust/return portion of systems. Exits and diverging junctions are only in supply systems. Equal-area elbows, obstructions, and duct-mounted equipment are common to both supply and exhaust systems. Transitions and unequal-area elbows can be either supply or exhaust fittings. Fitting
ED5-1 is an Exhaust fitting with a round shape (Diameter). The number 5 indicates that the fitting is a junction, and 1 is its sequential number. Fittings SR31 and ER3-1 are Supply and Exhaust fittings, respectively. The R indicates that the fitting is Rectangular, and the 3 identifies the fitting as an elbow. Note that the crosssectional areas at sections 0 and 1 are not equal. Otherwise, the elbow would be a Common fitting such as CR3-6. Terminal Unit Loss Coefficients. Manufacturers’ data for terminal units are not useful for duct design because they are given in terms of static pressure resistance and velocity pressure at standard air conditions (1.204 kg/m3). The total pressure loss coefficient for the 200 mm terminal used in Example 7 is calculated from manufacturer’s published data (Table 6: Columns 1, 2, and 7) and plotting pv,in versus pt (Figure 12). The loss coefficient is 1.67 and is applicable for any elevation (air density). The total pressure loss (pt) in Table 6 is only useful for projects at sea level. ASHRAE Standard 130-2016, Section 5.2, covers the laboratory test for total pressure loss and the calculation of the loss coefficient. In the ASHRAE Duct Fitting Database (ASHRAE 2016), the single-duct reheat VAV box is the CD8-series.
3.3
DUCTWORK SECTIONAL LOSSES
Darcy-Weisbach Equation Total pressure loss in a duct section is calculated by combining Equations (18) and (29) in terms of p, where C is the summation of local loss coefficients in the duct section. Each fitting loss coefficient must be referenced to that section’s velocity pressure. 1000 f L V 2 p = ------------------- + C --------- Dh 2
(34)
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21.12
2017 ASHRAE Handbook—Fundamentals (SI) Table 7 ED7-2 Loss Coefficients (see Figure 16) L/Do L/Do 0.5 0.75 1.0 1.5 2.0 4.0
0
2
5
10
1.80 1.40 1.20 1.10 1.00 0.67
1.00 0.80 0.67 0.60 0.53 0.40
0.53 0.40 0.33 0.33 0.33 0.22
0.53 0.40 0.33 0.33 0.33 0.22
Fan Outlet Conditions. Fans intended primarily for duct systems are usually tested with an outlet duct in place (AMCA Standard 210). Figure 14 shows the changes in velocity profiles at various distances from the fan outlet. For 100% recovery, the duct, including transition, must meet the requirements for 100% effective duct length [Le (Figure 14)], which is calculated as follows: For Vo > 13 m/s,
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Fig. 13
4.
Deficient System Performance with System Effect Ignored
Fan performance data measured in the field may show lower performance capacity than manufacturers’ ratings. The most common causes of deficient performance of the fan/system combination are poor outlet connections, nonuniform inlet flow, and swirl at the fan inlet. These conditions alter the fan’s aerodynamic characteristics so that its full flow potential is not realized. One bad connection can reduce fan performance below its rating. Normally, a fan is tested with open inlets and a section of straight duct attached to the outlet (AMCA Standard 210). This setup results in uniform flow into the fan and efficient static pressure recovery on the fan outlet. If good inlet and outlet conditions are not provided in the design of duct systems, fan performance suffers. Figure 13 shows deficient fan performance resulting from poor inlet and outlet connections to ductwork. Point 1 is the fan/system operating point without taking into account poor inlet and/or outlet conditions. Point 2 is the system operating point when the apparent resistance of poor connections is included in the calculations. Point 4 is the operating point on the original fan performance curve, taking into consideration the apparent system resistance of poor fan/system connections. Point 3 is the fan operating point on the original system curve when the apparent resistance of poor inlet and/or fan connections is not taken into account. The airflow difference between points 2 and 4 represents the deficiency in airflow from design airflow.
Fan System Effect Coefficients The system effect concept was formulated by Farquhar (1973) and Meyer (1973); the magnitudes of the system effect, called system effect factors, were determined experimentally by the Air Movement and Control Association International (AMCA 2011a; Brown 1973; Clarke et al. 1978). The system effect factors, converted to local loss coefficients, are in the ASHRAE Duct Fitting Database (ASHRAE 2016) for both centrifugal and axial fans. Fan system effect coefficients are only an approximation. Fans of different types and even fans of the same type, but supplied by different manufacturers, do not necessarily react to a system in the same way. Therefore, judgment based on experience must be applied to any design.
(35)
Ao Le = ----------350
(36)
For Vo 13 m/s,
FAN/SYSTEM INTERFACE
Fan Inlet and Outlet Conditions
Vo Ao Le = -----------------4500
where Vo = duct velocity, m/s Le = effective duct length, m Ao = duct area, mm2
Centrifugal fans should not abruptly discharge to the atmosphere. A diffuser design should be selected from Fitting SR7-2 or SR7-3. Consult the SR7-series fittings of the ASHRAE Duct Fitting Database (ASHRAE 2016) for guidance in the design of fan/ductwork connections. Fan Inlet Conditions. For rated performance, air must enter the fan uniformly over the inlet area in an axial direction without prerotation. Nonuniform flow into the inlet is the most common cause of reduced fan performance. Such inlet conditions are not equivalent to a simple increase in system resistance; therefore, they cannot be treated as a percentage decrease in the flow and pressure from the fan. A poor inlet condition results in an entirely new fan performance. Inlet spin may arise from many different approach conditions, and sometimes the cause is not obvious. Figure 15 shows some common duct connections that cause inlet spin. Inlet spin can be avoided by providing an adequate length of duct between the elbow and the fan inlet, as shown by Figure 16. Two L/Do duct lengths upstream of the fan inlet reduces swirl and pressure loss (loss coefficient) by approximately 45% (Table 7). Fans within plenums and cabinets or next to walls should be located so that air may flow unobstructed into the inlets. Fan performance is reduced if the space between the fan inlet and the enclosure is too restrictive. System effect coefficients for fans in an enclosure or adjacent to walls are listed under Fitting ED7-1. How the airstream enters an enclosure in relation to the fan inlets also affects fan performance. Plenum or enclosure inlets or walls that are not symmetrical with the fan inlets cause uneven flow and/or inlet spin.
5.
MECHANICAL EQUIPMENT ROOMS
In the initial phase of building design, the design engineer seldom has sufficient information to render the optimum HVAC design
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Duct Design
21.13
Fig. 14
Establishment of Uniform Velocity Profile in Straight Fan Outlet Duct (Adapted by permission from AMCA Publication 201)
Fig. 16
Fitting ED7-2 (Fan Inlet, Centrifugal Fan, SISW, with 4-Gore Elbow) [ASHRAE Duct Fitting Database (ASHRAE 2016)]
Fig. 15
Inlet Duct Connections Causing Inlet Spin
(Adapted by permission from AMCA Publication 201)
for the project, and its space requirements are often based on percentage of total area or other rule of thumb. The final design is usually a compromise between what the engineer recommends and what the architect can accommodate. Total mechanical and electrical space requirements range between 4 and 9% of gross building area, with most buildings in the 6 to 9% range. This range includes space for HVAC, electrical, plumbing, and fire protection equipment, as well as vertical shaft space for mechanical and electrical distribution through the building.
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21.14
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Outdoor Air Intake and Exhaust Air Discharge Locations A key factor in the location of mechanical equipment rooms is the source of outdoor air. If the air intake or exhaust system is not well designed, contaminants from nearby outdoor sources (e.g., vehicle exhaust) or from the building itself (e.g., laboratory fume hood exhaust) can enter the building with insufficient dilution. Poorly diluted contaminants may cause odors, health impacts, and reduced indoor air quality. Examples are toxic stack exhausts, automobile and truck traffic, kitchen cooking hoods, evaporative cooling towers, building general exhaust air, trash dumpsters, stagnant water bodies, snow and leaves, rain and fog, plumbing vents, vandalism, and terrorism. Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications discusses proper design of exhaust stacks and placement of air intakes to avoid adverse air quality impacts. Experience provides some general guidelines on air intake placement. Unless dispersion modeling analysis is conducted, air intakes should never be located on the roof in the same architectural screen enclosure as exhaust outlets. If exhaust is discharged from several locations on the roof, intakes should be located to minimize contamination. Typically, this means maximizing separation distance. Where all exhausts of concern are emitted from a single, relatively tall stack or tight cluster of stacks, a possible intake location might be close to the base of this tall stack, if this location is not adversely affected by other exhaust locations, or is not influenced by tall adjacent structures creating downwash. Architectural screens placed around rooftop equipment to reduce noise or hide equipment interact with the windflow patterns on the roof and can adversely affect exhaust dilution. Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications describes a method to account for these screens by modifying the physical stack height. When wind is perpendicular to the upwind wall, air flows up and down the wall, dividing at about two-thirds up the wall. The downward flow creates ground-level swirl that stirs up dust and debris. To take advantage of the natural separation of wind over the upper and lower halves of a building, toxic or nuisance exhausts should be located on the roof and intakes on the lower one-third of the building, but high enough to avoid wind-blown dust, debris, and vehicle exhaust. If ground-level sources are major sources of contaminants, rooftop intake is desirable. Buildings over three stories usually require vertical shafts to consolidate mechanical, electrical, and telecommunication distribution through the facility. Vertical shafts should be located in or adjacent to mechanical/fan rooms and as far as possible from noisesensitive areas. In general, duct shafts with an aspect ratio of 2:1 to 4:1 are easier to develop than large square shafts. The rectangular shaft also facilitates transition from equipment in the fan rooms to the shaft. Fan rooms in a basement or at street level should be avoided. These locations are a security concern because harmful substances could easily be introduced. Using louvers at these locations is also a concern because debris, leaves, and snow may fill the area, resulting in safety, health, and fan performance concerns. Loading docks and nearby parking areas may also compromise ventilation air quality.
Equipment Room Locations Mechanical equipment rooms, including air-handling units, should be centrally located to centralize maintenance and operation. But, for many reasons, not all equipment rooms can be centrally located in the building. In any case, equipment should be kept together whenever possible to minimize space requirement, centralize maintenance and operation, and simplify electrical systems. All HVAC air system equipment rooms should have space for maintaining equipment and the replacement of fans, coils and other key equipment.
2017 ASHRAE Handbook—Fundamentals (SI) High-rise buildings may opt for decentralized fan rooms for each floor, or for more centralized service with one mechanical/fan room serving the lower 10 to 20 floors, one serving the middle floors of the building, and one at the roof serving the top floors. Decentralized Equipment Rooms. Locate decentralized mechanical equipment rooms as far as possible from noise-sensitive areas, and surround equipment rooms with buffer zones such as toilet and storage rooms, as well as elevator, stair, and duct shafts. Figure 17 shows various core locations from poor to best. The number of decentralized fan rooms required depends largely on total floor area and any fan system power limitation imposed by codes or standards (e.g., ASHRAE Standard 90.1-2016, section 6.5.3.1). When decentralized air systems are located centrally to the spaces served, duct systems are shorter, occupy less volume, use less power, are quieter, and are less expensive. Pointing out these advantages can often help to convince architects to make desirable decentralized locations available.
6. DUCT DESIGN 6.1
DESIGN CONSIDERATIONS
HVAC System Air Leakage See Chapter 19 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment for (1) sealant specifications, and (2) the rationale for HVAC system sealing and leakage testing. System Sealing. All ductwork and plenum transverse joints, longitudinal seams, and duct penetrations should be sealed. Openings for rotating shafts should be sealed with bushings or other devices that minimize air leakage but that do not interfere with shaft rotation or prevent thermal expansion. Spiral lock seams need not be sealed. Duct-mounted equipment, such as terminal units, reheat coils, access doors, sound attenuators, balancing dampers, control dampers, and fire dampers, should be specified as low leakage so that the system can meet the air leakage acceptance criteria set by the designer, standards, and codes. Recommended specifications for duct-mounted equipment are provided later in this section. Sealing that would void product listings (e.g., for fire/smoke or volume control dampers) is not required. It is, however, recommended that the design engineer specify low-leakage duct-mounted components. For example, some UL-listed and -labeled fire/smoke dampers allow sealing and gasketing of breakaway duct/sleeve connections; all can provide sealed non-breakaway duct/sleeve connections. Scope. It is recommended that supply air (both upstream and downstream of the VAV box primary air inlet damper when used) and independent exhaust air systems be tested for air leakage after construction at operating conditions using ASHRAE Proposed Standard 215 (under development, and expected to be published in 2018) to verify (1) good workmanship, and (2) the use of low-leakage components as required to achieve the design allowable system air leakage. To ensure that a system passes its air leakage test at operating conditions, sufficient ductwork sections should be leak tested during construction. Leakage of duct-mounted components should be determined by specification and certified leakage data provided with equipment submittals. Leakage of air-handling units should be determined by specification and verification leakage tests Independent exhaust air system: air discharged from a space to the outdoors by a system not coupled to supply or return air systems. Acceptance Criteria. To enable proper accounting of leakagerelated impacts on fan energy and space conditioning loads, the
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Duct Design
21.15
Fig. 17 Comparison of Various Mechanical Equipment Room Locations (Schaffer 2011)
allowable system air leakage for each fan system should be established by the design engineer as a percentage of fan airflow at the maximum system operating conditions. The recommended maximum system leakage is 5% of design airflow. Exceptions: supply and return ductwork sections that leak directly to/from outdoors and exhaust system ductwork sections that draw in indoor air through leaks should be limited to 2%. Equation (37) is for use in a leakage test during construction to ensure that ductwork will meet the leakage specification. This equation translates system fractional air leakage to test section leakage class, as specified by ASHRAE Standard 90.1-2016, section 6.4.4.2.2. Q leak frac 100 Q fan A system CL, section = -------------------------------------------------------------------------------0.65 p sec tion
(37)
where CL,section=
test section leakage class, L/s per Pa0.65 per square metre of duct
surface area Qfan = maximum fan airflow that would occur during operation, L/s} Qleak, frac= system leakage fraction corresponding to maximum fan airflow that would occur during operation, % Asystem = total system duct surface area, m2 psection = test section static pressure difference corresponding to maximum fan airflow that would occur during operation, Pa
Equation (37) shows that leakage class depends on fractional air leakage and normalized fan airflow (Qfan/Asystem), and varies inversely with the pressure difference raised to the 0.65 power. For example, to achieve 3% leakage for ductwork with Qfan/Asystem = 10 L/s per m2 of duct surface area and psection = 750 Pa, the required leakage class is 0.004 L/s per Pa0.65 per square metre of duct
surface area. With psection =125 Pa instead (six times less), the required leakage class is about three times greater (0.13). The maximum acceptable air leakage for a test section corresponding to the leakage class determined by Equation (37) can be expressed by Equation (38). 0.65
Qleak, section = CL, section A sec tion p sec tion
(38)
where Qleak,section=test section air leakage, L/s Asection = test section duct surface area, m2
Equations (37) and (38) can be combined so that the maximum acceptable air leakage for a test section during construction is simply a function of system fractional leakage, normalized fan airflow (Qfan/Asystem), and section duct surface area: Q leak, frac Qleak, section = ------------------------- (Qfan/Asystem) Asection 100
(39)
Thus, for air leakage tests during construction, the maximum acceptable leakage for a ductwork section is given by Equation (38) or (39). Duct surface area should be calculated in accordance with European Standard EN 14239. The test pressure for each section should be specified by the design engineer based on the maximum static pressure for that section that would occur during operation at maximum fan airflow. Example 7. The system depicted by Figure 18 has the characteristics summarized by Table 8. For a maximum acceptable leakage of 3% of the 2360 L/s maximum fan airflow, which is 71 L/s total, what is the maximum allowable ductwork leakage in (1) each section that is to be tested at the pressures noted, and (2) sections 4 and 5 when leak tested together?
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21.16
2017 ASHRAE Handbook—Fundamentals (SI)
Solution: The calculations and maximum allowable leakage are also summarized in Table 8. Note that adjacent sections with the same static pressure can be grouped for leakage testing. In this case, Sections 4 and 5 can be grouped. The allowable leakage for Sections 4 and 5 when tested individually is 4 L/s and 20 L/s respectively. The allowable leakage for Sections 4 and 5 when tested together is 24 L/s.
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Recommended Specification for Duct-Mounted Equipment. Duct-mounted component leakage is controlled by specification and certified leakage data provided with equipment submittals. The following are recommended leakage-related specifications for ductmounted equipment. • Terminal Units. Seal longitudinal seams of casings, inlet face of casings, and inlet collars with mastic. Seal damper shaft penetrations of casings. Casing leakage for the basic terminal unit should not exceed 2.1 L/s at 2.49 Pa static pressure differential. Terminal unit leakage with an access door should not exceed 2.2 L/s at 2.49 Pa. Testing should be by an accredited laboratory. Leakage tests should comply with ASHRAE Standard 130. Access doors in terminal unit casings should comply with AMCA Standard 500-D, and leakage rates should be certified per AMCA (2013) Publication 511. Access door (frame not included) leakage should not exceed 0.24 L/s at 2.49 kPa static pressure differential. • Electric Reheat Coils. Flange-mount electric coils with a gasket. Coil leakage should not exceed 0.24 L/s at 249 Pa static pressure differential. Tests should be by an accredited laboratory. Leakage tests should be in compliance with ASHRAE Standard 126. • Hot-Water Reheat Coils. Flange-mount hot-water coils with a gasket in an insulated plenum or casing. Seal all seams and casing penetrations for supply and return water tubes. Coil casing leakage (not counting transverse joints) should not exceed 0.24 L/s at 249 Pa static pressure differential. An accredited laboratory should perform leakage tests in compliance with ASHRAE Standard 126. • Access Doors. Access door leakage (excluding the frame) should not exceed 0.24 L/s at 2.49 kPa static pressure differential. Leakage tests should comply with AMCA Standard 500-D, and leakage rates certified per AMCA (2013) Publication 511. Table 8 Solution for Example 7 Section Section Section Section Leakage Class, Allowable Static Inlet Flow, Asection, Pressure, L/s per m2 per Leakage, L/s Pa0.65 m2 Pa L/s Equation (38) Section Input Equation (37) or (39) 1 2 3 4 5
2 360 944 1 416 472 944
Total 4 and 5
46.5 62.0 69.7 18.6 92.9
750 250 500 250 250
0.003 0.006 0.004 0.006 0.006
250
0.006
289.6 1 416
111.5
10 14 15 4 20 63 24
• Attenuators. Sound attenuator casing seams should be sealed at the factory if used in-line with the ductwork. Attenuators stacked in plenums do not have to be sealed, but plenums should be sealed. • Fire/Smoke Dampers. These dampers should be installed in accordance with the manufacturer’s UL installation instructions. Each fire damper should be furnished with a UL-approved sleeve. Sleeve seams should be continuously welded or sealed, and the transverse joint should be a sealed UL-approved flanged duct sleeve connection (break-away or non-break-away). • Balancing Dampers. Balancing damper casing seams should be continuously welded or sealed, and the shaft penetrating the casing should have seals. • Control Dampers. Control damper shafts penetrating ducts should have seals. Recommended Specification for Air-Handling Units. Refer to Chapter 19 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment for leakage-related specifications for air-handling units. Responsibilities. The engineer should • Specify HVAC system components, duct-mounted equipment, sealants, and sealing procedures that together will meet the system airtightness design objective. • Inspect the system during construction for quality of workmanship and to verify that correct duct-mounted components and airhandling units are installed. • Specify the construction-stage ductwork leakage test standard or procedures. • Specify the construction-stage test pressures to the nearest 25 Pa expected during operation at design conditions and the maximum allowable air leakage. • Review and approve the sheet metal and test contractors’ leakage test reports. If any system has a leakage failure, the engineer should discuss remedies with the sheet metal contractor, vendor, and/or owner’s representative. The sheet metal contractor should • Construct the system using quality workmanship and correct duct-mounted components and air-handling units. If any installed duct-mounted equipment appears to be leakage suspect, the contractor should discuss remedies with the engineer and/or owner’s representative. • Conduct ductwork leakage pressurization tests during construction. As a minimum, 25% of the ductwork system (based on duct surface area) should be tested during construction, and another 25% if any of the initial sections fail. If any section of the second 25% fails, the entire ductwork system should be leak tested. • Provide connections for test apparatus, and separate test sections from each other as needed so that the test apparatus capacity is not exceeded. • Report test results and, where required, take corrective action to seal ductwork and absorb the cost for conducting related additional leak tests. The test contractor should • Conduct the operating system leakage test in compliance with ASHRAE Proposed Standard 215 (expected to publish in 2018). • Report test results, including reasons for any failures. It should be the responsibility of the owner or owner’s representative to provide direction upon request.
Fire and Smoke Control
Fig. 18
Duct Layout for Example 7
Because duct systems can convey smoke, hot gases, and fire from one area to another and can accelerate a fire within the system, fire protection is an essential part of air-conditioning and ventilation system design. Generally, fire safety codes require compliance with the standards of national organizations. NFPA
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Duct Design Standard 90A examines fire safety requirements for (1) ducts, connectors, and appurtenances; (2) plenums and corridors; (3) air outlets, air inlets, and fresh air intakes; (4) air filters; (5) fans; (6) electric wiring and equipment; (7) air-cooling and -heating equipment; (8) building construction, including protection of penetrations; and (9) controls, including smoke control. Fire safety codes often refer to the testing and labeling practices of nationally recognized laboratories, such as Factory Mutual and Underwriters Laboratories (UL). UL’s annual Building Materials Directory lists fire and smoke dampers that have been tested and meet the requirements of UL Standards 555 and 555S. This directory also summarizes maximum allowable sizes for individual dampers and assemblies of these dampers. Fire dampers are 1.5 h or 3 h fire-rated. Smoke dampers are classified by (1) temperature degradation [ambient air or high temperature (120°C minimum)] and (2) leakage at 250 and 1000 Pa pressure difference (2 and 3 kPa classification optional). Smoke dampers are tested under conditions of maximum airflow. UL’s annual Fire Resistance Directory lists fire resistances of floor/roof and ceiling assemblies with and without ceiling fire dampers. For a more detailed presentation of fire protection, see the NFPA (2008) Fire Protection Handbook, Chapter 53 of the 2015 ASHRAE Handbook—HVAC Applications, and Klote et al. (2012).
Duct Insulation In all new construction (except low-rise residential buildings), airhandling ducts and plenums that are part of an HVAC air distribution system should be thermally insulated in accordance with ASHRAE Standard 90.1. Duct insulation for new low-rise residential buildings should comply with ASHRAE Standard 90.2. Existing buildings should meet requirements of ASHRAE Standard 100. In all cases, thermal insulation should meet local code requirements. Insulation thicknesses in these standards are minimum values; economic and thermal considerations may justify higher insulation levels. Additional insulation, vapor retarders, or both may be required to limit vapor transmission and condensation. Duct heat gains or losses must be known to calculate supply air quantities, supply air temperatures, and coil loads. To estimate duct heat transfer and entering or leaving air temperatures, refer to Chapters 4 and 23.
21.17 Louvers Use Figure 19 for preliminary sizing of air intake and exhaust louvers. For air quantities greater than 3300 L/s per louver, the air intake gross louver openings are based on 2 m/s; for exhaust louvers, 2.5 m/s is used for air quantities of 2400 L/s per louver and greater. For smaller air quantities, see Figure 19. These criteria are presented on a per-louver basis (i.e., each louver in a bank of louvers) to include each louver frame. Representative productionrun louvers were used in establishing Figure 19, and all data used were based on AMCA Standard 500-L tests. For louvers larger than 1.5 m2, the free areas are greater than 45%; for louvers less than 1.5 m2, free areas are less than 45%. Unless specific louver data are analyzed, no louver should have a face area less than 0.4 m2. If debris can collect on the screen of an intake louver, or if louvers are located at grade with adjacent pedestrian traffic, louver face velocity should not exceed 0.5 m/s. Louvers require special treatment because the blade shapes, angles, and spacing cause significant variations in louver-free area and performance (pressure drop and water penetration). Selection and analysis should be based on test data obtained from the manufacturer in accordance with AMCA Standard 500-L, which presents both pressure drop and water penetration test procedures and a uniform method for calculating the free area of a louver. Tests are conducted on a 1220 mm square louver with the frame mounted flush in the wall. For water penetration tests, rainfall is 100 mm/h, no wind, and the water flow down the wall is 0.05 L/s per linear metre of louver width. AMCA Standard 500-L also includes a method for measuring water rejection performance of louvers. These louvers are subjected to simulated rain and wind pressure and tested at a rainfall of 76 mm/h falling on the louver’s face with a predetermined wind velocity directed at the face of the louver (typically 13 or 20 m/s). Effectiveness ratings are assigned at various airflow rates through the louver.
Physical Security Ducts entering into spaces considered secured or sensitive should contain measures to detect or inhibit forced entry into those spaces through the duct system. Inhibiting measures should be based on a delay or resistance time set by the facility owner or user. The following security measures should be considered: • Barrier duct bars: placement should be at locations deemed appropriate by the facility owner and user, or where secured or sensitive area boundaries exist (e.g., duct penetration through the wall of a secured room). Bar diameter, material, and spacing should be established by the facility owner or user. Ensure that barriers do not inhibit proper operation and maintenance of dampers, detectors, sensors, and other devices in the duct system. • Welded diffuser and return grilles: grille material and spacing should be established by the facility owner or user. • Sensors on access doors: sensors should be connected to a facility security system or panel, and notify security personnel of a duct system breach. Communicating HVAC maintenance schedules to security personnel is necessary to ensure that an intrusive duct system breach is not confused with HVAC system maintenance.
Fig. 19 Criteria for Louver Sizing
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21.18
2017 ASHRAE Handbook—Fundamentals (SI)
Duct Shape Selection No Space Constraints. Round ductwork is preferable to rectangular or flat oval ductwork when adequate space is available for the following reasons.
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• Mass of round ductwork is less than rectangular. Figure 20 shows the relative weight of rectangular duct to round duct for duct pressures from ±125 to ±2500 Pa when the equivalent diameter of the rectangular duct is the same as the round duct diameter. Equivalent diameter is defined as the diameter of a rectangular duct that has equal resistance to flow for equal flow and length. • Perimeter of round ducts is less than rectangular ducts. For rectangular duct aspect ratios from 2 to 4, the increase is approximately 30 to 55%. This increase results in increased insulation, including possible thickness to offset the additional heat transfer. For a comprehensive study of round and rectangular ducts as they affect system performance, consult McGill (1988). • Round ducts have an excellent resistance to low-frequency breakout noise (Schaffer 2011). • Duct rumble can occur in rectangular duct systems (Paulauskis 2016).
Note: Negative-pressure flat oval duct systems can be designed by using +2500 Pa sheet gages with the negative-pressure rectangular reinforcement welded to the duct. • Low-frequency breakout noise for flat oval is good, fair for rectangular (Schaffer 2011). • Duct rumble can occur in rectangular duct systems (Paulauskis 2016). • Perimeter. Using rectangular duct instead of flat oval when sized for equivalent diameter increases the perimeter roughly 17 to 7% for aspect ratios ranging from 1 to 4, respectively. This increase
Space Constraints. Space constraints and obstructions, particularly ceiling height, are frequent problems. In these cases, the choice is round, rectangular, or flat oval ducts, depending on the air quantity that needs to be conveyed by the duct. Table 9 and Figure 21 cover three design cases: 0.65, 2, and 5 Pa/m friction rates. Friction rate 2 Pa/m is at the middle range. In Table 9, six ceiling (plenum) heights ranging from 400 to 1100 mm, in approximately 200 mm increments, are covered. Space allocated for insulation and reinforcement is 50 mm all around. The aspect ratio of the rectangular and flat oval ducts is 2:1. Figure 21 shows the maximum airflow as a function of the design friction rates and plenum spaces, as noted. For example, the 900 mm plenum as a design friction rate of 2 Pa/m is 6360 L/s maximum for an 800 mm round duct; 16 300 L/s for a 1600 × 800 mm rectangular duct; and 14 500 L/s for a 1600 800 mm flat oval duct. The airflow capacity of two round 800 mm ducts is 6360 L/s each. When selecting a rectangular or flat oval duct, consider the following: • Rectangular duct has the advantage in mass because construction standards for flat oval exist only for +2500 Pa, whereas rectangular has seven pressure classes, starting at 125 Pa. All rectangular pressure classes are ±.
Fig. 20
Relative Weight of Rectangular Duct to Round Spiral Duct
Fig. 21 Maximum Airflow of Round, Flat Oval, and Rectangular Ducts as Function of Available Ceiling Space
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Duct Design
21.19
Table 9 Maximum Airflow of Round, Flat Oval and Rectangular Ducts as Function of Available Ceiling Space* A. Design Criterion: 0.65 Pa/m or 12.7 m/s Maximum Minimum Clearance for Duct, mm Single Round Duct Duct diameter, mm Airflow, L/s Velocity, m/s Rectangular Duct with Aspect Ratio = 2 Rectangular W × H, mm Airflow, L/s Velocity, m/s Equivalent diameter De, mm Flat Oval Duct with Aspect Ratio = 2 Flat Oval A × a, mm Airflow, L/s Velocity, m/s Equivalent diameter De, mm Two Round Ducts in Parallel Duct diameter, mm Airflow, L/s Velocity, m/s
500
600
730
900
1100
400 680 5.4
500 1230 6.3
630 2270 7.3
800 4290 8.5
1000 7730 9.8
800 × 400 2100 6.6 609
1000 × 500 3780 7.6 762
1200 × 600 6150 8.5 914
1600 × 800 13 100 10.2 1219
2000 × 1000 23 500 11.8 1523
800 × 400 2000 7.0 591
1000 × 500 3600 8.1 739
1200 × 600 5850 9.1 887
1600 × 800 12 500 10.9 1183
2000 × 1000 22 500 12.6 1479
Two 400 680 each 5.4
Two 500 1230 each 6.3
Two 630 2270 each 7.3
Two 800 4290 each 8.5
Two 1000 7730 each 9.8
500
600
730
900
1100
400 1250 9.9
500 2250 11.5
630 3950 12.7
800 6360 12.7
1000 10 000 12.7
800 × 400 3820 11.9 609
1000 × 500 6350 12.7 762
1200 × 600 9150 12.7 914
1600 × 800 16 300 12.7 1219
2000 × 1000 25 400 12.7 1523
800 × 400 3630 12.7 591
1000 × 500 5650 12.7 739
1200 × 600 8150 12.7 887
1600 × 800 14 500 12.7 1183
2000 × 1000 22 600 12.7 1479
Two 400 1250 each 9.9
Two 500 2250 each 11.5
Two 630 3950 each 12.7
Two 800 6360 each 12.7
Two 1000 10 000 each 12.7
500
600
730
900
1100
400 1910 15.2
500 2990 15.2
630 4750 15.2
800 7620 15.2
1000 11 900 15.2
800 × 400 4850 15.2 609
1000 × 500 7600 15.2 762
1200 × 600 10 970 15.2 914
1600 × 800 19 500 15.2 1219
2000 × 1000 30 400 15.2 1523
800 × 400 4350 15.2 591
1000 × 500 6800 15.2 739
1200 × 600 9800 15.2 887
1600 × 800 17 400 15.2 1183
2000 × 1000 27 200 15.2 1479
Two 400 1910 each 15.2
Two 500 2990 each 15.2
Two 630 4750 each 15.2
Two 800 7620 each 15.2
Two 1000 11 900 each 15.2
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B. Design Criterion: 2 Pa/m or 12.7 m/s Maximum Minimum Clearance for Duct, mm Single Round Duct Duct diameter, mm Airflow, L/s Velocity, m/s Rectangular Duct with Aspect Ratio = 2 Rectangular W × H, mm Airflow, L/s Velocity, m/s Equivalent Diameter De, mm Flat Oval Duct with Aspect Ratio = 2 Flat oval A × a, mm Airflow, L/s, Velocity, m/s Equivalent diameter De, mm Two Round Ducts in Parallel Duct diameter, mm Airflow, L/s Velocity, m/s C. Design Criterion: 5 Pa/m or 15.2 m/s Maximum Minimum Clearance for Duct, mm Single Round Duct Duct diameter, mm Airflow, L/s Velocity, m/s Rectangular Duct with Aspect Ratio = 2 Rectangular W × H, mm Airflow, L/s Velocity, m/s Equivalent Diameter De, mm Flat Oval Duct with Aspect Ratio = 2 Flat oval A × a, mm Airflow, L/s Velocity, mm Equivalent diameter De, mm Two Round Ducts in Parallel Duct diameter, mm Airflow, L/s Velocity, m/s
*European standard metric dimensions used to develop Table 7 (CEN 1998, 2007). Round duct diameters (mm): 80, 100, 125, 160, 200, 250, 315, 400, 500, 630, 800, 1000, 1250, and 1600. Rectangular/flat oval dimension (mm): 100, 150, 200, 250, 300, 400, 500, 600, 800, 1000, 1200, 1400, 1600, 1800, and 2000.
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21.20
2017 ASHRAE Handbook—Fundamentals (SI)
results in an increased surface area. Flat oval requires less insulation. For a comprehensive study of flat oval and rectangular ducts as they affect system performance, consult McGill (1995). • Duct Lengths. Rectangular duct is available in 1200, 1500, or 1800 mm lengths. Spiral round and flat oval can be provided in longer lengths.
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Testing and Balancing Each air duct system should be tested, adjusted, and balanced. Guidance and procedures are given in Chapter 38 of the 2015 ASHRAE Handbook—HVAC Applications and in ASHRAE Standard 111. To determine fan total (or static) pressure from field measurements taking into account fan system effect, consult AMCA (2011b) Publication 203, which provides numerous fan/system configurations encountered in the field. Many VAV noise complaints have been traced to control problems. Although most problems are associated with improper installation, many are caused by poor design. The designer should specify high-quality fans or air handlers within their optimum ranges, not at the edge of their operation ranges, where low system tolerances can lead to inaccurate fan flow capacity control. Also, in-duct static pressure sensors should be placed in duct sections having the lowest possible air turbulence (i.e., at least three equivalent duct diameters from any elbow, takeoff, transition, offset, or damper).
6.2
DESIGN RECOMMENDATIONS
• Engage the architect and structural engineer early to coordinate shafts for systems. Commentary. Consult Noise and Vibration Control for HVAC Systems (Schaffer 2011) for the following guidelines related to duct systems: • • • • •
Fig. 24 Guidelines for Centrifugal Fan Installations (Schaffer 2011)
significantly increase fan-generated noise. To account for fan inlet system effects, use the ASHRAE Duct Fitting Database (ASHRAE 2016) (ED7 and ER7 series). • For all except very-noise-sensitive applications, select VAV reheat boxes for a total pressure loss from 125 to 150 Pa; for a fanpowered VAV box, from 150 to 175 Pa. For details, see Taylor and Stein (2004). • VAV terminal unit inlet duct should be the same size as the inlet to the box, unless the box is in the critical path or the length exceeds about 4.5 m from the takeoff. Duct upstream of box inlets should be rigid sheet metal duct, 1.2 m minimum. Do not use flexible duct immediately upstream of VAV boxes. • See Figure 24 for Schaffer’s (2011) guidelines for the installation of single-duct, dual-duct, and induction terminal units, as well as parallel and series flow fan-powered units.
Selection of mechanical room walls Noise control for mechanical rooms Upward noise control for mechanical rooms Duct penetrations through walls Structural support of rooftop equipment for vibration control
• Route ducts as straight as possible to reduce pressure loss, noise, and first costs. • Use round spiral ducts whenever round ducts can fit within space constraints. • Avoid consecutive fittings and close-coupled fittings because they can significantly increase pressure losses. • Use return air plenums when possible because they reduce both energy and first costs. Plenum return requires fire-rated construction. • Design air distribution systems to minimize flow resistance and turbulence. High flow resistance increases fan pressure, which results in higher noise being generated by the fan, especially at low frequencies. Turbulence also increases flow noise generated by duct fittings and dampers, especially at low frequencies. • Efficient fittings create the least turbulence and noise. Figure 22 provides generalized guidelines for minimizing regenerated noise from takeoffs. Tables 10 and 11 provide specific guidance for tees and wyes. • Duct transitions should not exceed an included angle of 15°. • To avoid fan system effects, fans should discharge into duct sections that remain straight for as long as possible, up to 10 duct diameters from the fan discharge to allow flow to fully develop (Figure 23). Use the ASHRAE Duct Fitting Database (ASHRAE 2016) to account for fan outlet system effects (SD7 and SR7series). • Design duct connections at the fan inlet for uniform and straight airflow. Both turbulence and flow separation at the fan blades can
Fig. 22 Guidelines For Minimizing Regenerated Noise in Takeoff (Schaffer 2011)
Fig. 23 Guidelines for Centrifugal Fan Installations (Schaffer 2011)
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Duct Design
21.21 Table 10
Options for Selecting 90° Takeoff
Licensed for single user. © 2017 ASHRAE, Inc.
Loss Coefficient Efficiency
Maina
Branchb
Highest
0.15
0.64
Wye, 45°, Straight body branch with 45° elbow, 90° to main
—
0.15
0.74
SD5-11
Tee, Conical branch
—
0.15
0.87
SD5-10
Tee, Conical branch tapered into body
—
0.15
1.10
SD5-9
Tee
Lowest
0.15
1.80
Code
Description
SD5-12
Tee, 45° entry branch
SD5-4
aQ
s /Qc = 0.8; As /Ac = 0.69 bQ /Q = 0.2; A /A = 0.25 b c b c
• Place fan-powered mixing boxes away from noise-sensitive areas. • Use demand-based static pressure set-point reset to reduce fan energy and noise. • For constant-volume systems, select the fan to operate as near as possible to its rated peak efficiency. Also, select a fan that generates the lowest possible noise at required design conditions. Using an oversized or undersized fan that does not operate at or near rated peak efficiency can substantially increase noise levels. For VAV applications, see Schaffer (2011).
6.3 DESIGN METHODS Equal Friction Method. The equal friction method for sizing duct systems uses a constant friction rate. The target velocity determines the size of the first duct section both downstream and upstream of the fan. From the size determined by the target velocity, the design friction rate is determined to size all remaining duct sections except for connections to VAV and constant-volume (CV) terminal units and diffusers in the critical path, or for sections whose length exceeds around 4.5 m. In these cases, the inlet to terminal units should have at least three diameters of rigid duct and 2 m maximum of rigid or flexible duct (see Figure 9) to diffusers. The section upstream of the rigid duct to terminal units and the section of rigid/ flexible duct to diffusers should be sized by the design friction rate,
and an appropriate transition placed between sections. Refer to sections 8 and 9 in Figure 26 for an example. Static Regain Method. The static regain method uses the conservation of momentum principle, which results in an increase in static pressure when the velocity is reduced in an airstream. This method sizes the main and branch ducts after a junction so that the recovery in static pressure caused by the reduced velocity is approximately equal to the total pressure drop caused by the duct and fittings in the subject duct section according to Equation (41). This is done iteratively by selecting a size for a section and checking to see if the section’s static pressure loss is close to zero. The basic steps are to design the first section at the velocity recommended in Table 12. Then each downstream section is subsequently sized using the same duct size as the upstream section as a starting point. If there is no static regain, meaning there is a positive static pressure loss, then the downstream section will be the same size as the upstream section. But, if there is static regain, meaning the change in static pressure is negative, then an opportunity exists to use a smaller duct section. Typically, the next smaller size available that is smaller than the upstream section is selected and the calculations are repeated. Refer to column 12 of Table 15 for examples of static regain.
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21.22
2017 ASHRAE Handbook—Fundamentals (SI) Table 11
Options for Selecting 45° Takeoff (Wye)
Licensed for single user. © 2017 ASHRAE, Inc.
Loss Coefficient Code
Description
SD5-2
Wye, 45°, Conical branch tapered into body
SD5-3
Wye, 45°, Straight body branch with reduction transition
SD5-1
Wye, 45°
Efficiency
Main, Csa
Branch, Cbb
Highest
0.15
0.45
—
0.15
0.50
Lowest
0.15
0.70
aQ /Q = 0.8; A /A = 0.69 s c s c bQ /Q = 0.2; A /A = 0.25 b c b c
Table 12
Recommended Maximum Airflow Velocities to Achieve Specified Acoustic Design Criteria*
NC or RC Rating in Adjoining Occupancy
Duct Location
Maximum Airflow Velocity, m/s Rectangular Duct
Round Duct
1
2
3
4
In shaft or above solid drywall ceiling
45 35 25 or less
17.8 12.7 7.6
25.4 17.8 12.7
Above suspended acoustical ceiling
45 35 25 or less
12.7 8.9 5.1
22.9 15.2 10.2
Duct within occupied space
45 35 25 or less
10.2 7.4 4.8
19.8 13.2 8.6
For connections to VAV and CV terminal units and diffusers in the critical path or the section length with the terminal unit or diffuser exceeds around 4.5 m, the inlet to terminal units should have at least three diameters of rigid duct and 2 m maximum of rigid or flexible duct (see Figure 6) to diffusers. The section upstream of the rigid duct to terminal units and the section of rigid/flexible duct to diffusers should be sized by the static regain method and an appropriate transition placed between sections. Refer to sections 8 and 9 in Figure 26 for an example. pv1 = pt,1–2 + pv2 + ps,1–2 (40) Rearranging, ps,1–2 = (pv1 – pv2) – pt,1–2 pv1 = velocity pressure at section 1, Pa pv2 = velocity pressure at section 2, Pa
Balancing. When system unbalance is greater than about 30%, it is recommended that balance be improved by changing duct size, and/or fittings from a select few that have comparable efficiency without introducing additional significant turbulence (noise). Tables 10 and 11 are examples.
Noise Control
*Adapted from Table 9 in Chapter 48 of the 2015 ASHRAE Handbook—HVAC Applications and Schaffer 2011.
where
pt,1–2 = total pressure loss from section 1 to 2, Pa ps,1–2 = static pressure regain (+) or loss (–), Pa
(41)
Understanding what noise is and how to control it is fundamental to duct design. The underlying principles of sound and vibration are covered in Chapter 8. Chapter 48 of the 2015 ASHRAE Handbook— HVAC Applications contains technical discussions and design examples helpful to the design engineer. AHRI Standard 885 has procedures for estimating sound pressure levels in the occupied zone for the portion of the system downstream of terminal units. For guidance in designing HVAC systems to avoid noise and vibration problems, consult Schaffer (2011). Specifying quiet equipment and designing systems to avoid noise and vibration problems are necessary parts of the design process. Noise in system comes from fans and generated noise resulting from air turbulence in ducts, individual fittings, close-coupled fittings, dampers, air modulating equipment and diffusers/grilles. Objectionable self-generated duct noise can be controlled by limiting the duct velocity to the values listed in Table 12.
Goals The goal of duct design is an air distribution system without objectionable noise and minimum life-cycle cost (LCC). Noise can be controlled by limiting the duct velocity to the values listed in Table 12. For example, for ductwork above a suspended acoustical ceiling and with a maximum allowable NC or RC of 35 in the adjoining space, the maximum round and rectangular duct airflow velocities are 15.2 and 8.9 m/s, respectively. Use this velocity to size the first duct section either upstream or downstream from the fan for all duct design methods. Lowering duct velocity reduces system
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Duct Design
21.23
operating cost and minimizes noise from sources other than ducts (fittings, close-coupled fittings, and dampers). Commentary: In 1988, Dr. Robert J. Tsal developed the Tmethod (Tsal and Adler 1987; Tsal et al. 1988, 1990) to design systems with a minimum LCC. The T-method was technically sound, but the system cost was not sufficiently accurate. As a result, the T-method was removed from the Handbook in 2013.
Design Method to Use
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Supply Duct Sizing. • Size fan supply ducts by either the equal friction (EF) or static regain (SR) method. The duct velocity anywhere in the system should not exceed the velocity listed in Table 12. • Size ducts downstream of terminal boxes, toilet exhaust ducts, and other low-pressure systems (e.g., in Figure 25C) using the equal friction method with a friction rate in the range from 0.4 to 1.6 Pa/m such that the duct velocity in the duct anywhere in the system does not exceed the values in Table 12. Use Table 13 as a guide to select the design friction rate. • Terminal unit (VAV box) runouts should be full size with the exception that runouts in the “critical” path or a runout length greater than about 4.5 m should be a minimum 1.5 m of rigid duct full size and the remainder’s size determined by the design method (e.g., Example 8 and Figure 26). • Diffuser runouts should be full size, except for runouts in the critical path or a runout length greater than about 4.5 m: these runouts should be at least 2 m of flexible-duct full size, and the remainder’s size determined by the design method. Return Duct Systems. • Sizing ducted returns depend on economizer relief system. For systems with return fans (Figure 25A), the return air ducts are sized using the EF method. The SR method does not work for negative-pressure duct systems. If building relief is by relief fans or gravity/motorized dampers, pressure drop should be kept low as shown by the total pressure grade line associated with Figures 25B and 25C. Control dampers in the economizer system should be selected (parallel or opposed blade) and sized in compliance with ASHRAE Guideline 16. Louvers should be sized in accordance with the section on Louvers, under Duct Design. • Unducted return air shafts are typically sized for low pressure loss using either a fixed friction rate, velocity, or both. Typically,
shafts are simply sized based on velocity. Maximum velocities are generally in the 4.1 to 6.1 m/s range through the free area at the top of the shaft (highest airflow rate). • Diffuser/grille runouts should be designed by the design method of choice and a transition located at the outlet. For outlets in the critical path, the outlet duct size should be increased. Registers should not be used because of the damper. Example 8. For the VAV system shown in Figure 26, design the duct system by both the equal friction (EF) and static regain (SR) methods, and compare the section duct sizes, total pressure required for each path, and the unbalance between paths. The system is located in Denver (1655 m elevation) and the duct is spiral round, galvanized steel (absolute roughness = 0.12 mm). The duct system is located above a suspended acoustical ceiling, and the allowable background sound in the occupied spaces is NC-35. Terminals T1, T2, T3, and T4 (VAV boxes with a one-row hot-water coil) are 380 L/s. VAV box loss coefficients are 1.67. Solution: Use the ASHRAE Duct Fitting Database (ASHRAE 2016) to determine the air density, which is used to identify the velocity pressure and fitting loss coefficients for fittings. Density is 0.978 kg/m3 (Figure 27). The maximum design duct velocity from Table 12 is 15.2 m/s. The EF and SR calculations are shown by Tables 14 and 15 spread-sheets. For both the EF and SR designs, the duct size for Section 2 is 375 mm determined using CD11-3 (Figure 28). Sizing the remaining section depends on the design method. For the equal friction design method, the system equal friction (also from CD11-3) is 4.2 Pa/m. Knowing the system design friction rate, the duct sizes are determined by CD11-4. Figure 29 is an example for Section 4. For the static regain method, all sections other than the terminal unit runouts are sized by iterating using Equation (41) (column 12 on the spreadsheet). For example, the results of the iteration for Section 4 are summarized by Table 16. The solution is the diameter that gives the result from Equation (41) that is closest to zero: in this case, Tables 17 and 18 summarize the results. The SR design is slightly better balanced, 19% compared to 23%, and the SR design requires 9% less energy. In addition, the SR sizes are slightly larger, and thus potentially less noisy. Both methods have static regain, but the SR design uses the regain more efficiently.
6.4
INDUSTRIAL EXHAUST SYSTEMS
Chapter 32 of the 2015 ASHRAE Handbook—HVAC Applications discusses design criteria, including hood design, for industrial exhaust systems. Exhaust systems conveying vapors, gases, and smoke are designed by the equal-friction method. Systems conveying particulates are designed by the constant velocity method at duct velocities adequate to convey particles to the system air cleaner. For contaminant transport velocities, see Table 2 in Chapter 32 of the 2015 ASHRAE Handbook—HVAC Applications.
Table 13 Guide for Selecting Low-Pressure System Friction Rate* Airflow 0.4 Pa/m Q, D, V, L/s mm m/s 240 470 940 1 420 1 890 2 360 4 720 7 080 9 440 11 800 14 160
312 402 520 608 678 738 958 1118 1248 1358 1456
3.1 3.7 4.4 4.9 5.2 5.5 6.5 7.2 7,7 8.1 8.5
0.8 Pa/m
1.0 Pa/m
1.2 Pa/m
1.6 Pa/m
D, V, mm m/s
D, V, mm m/s
D, V, mm m/s
D, V, mm m/s
270 4.2 258 4.6 248 5.0 234 348 4.9 332 5.4 320 5.8 302 452 5.9 432 6.4 416 6.9 394 530 6.4 506 7.1 488 7.6 460 590 6.9 564 7.6 544 8.1 512 642 7.3 614 8.0 592 8.6 558 834 8.6 796 9.5 768 10.2 726 972 9.5 930 10.4 896 11.2 846 1086 10.2 1038 11.2 1002 12.0 946 1182 10.8 1130 11.5 1090 12.6 1030 1268 11.2 1212 12.3 1168 13.2 1104
5.6 6.6 7.7 8.5 9.2 9.7 11.4 12.6 13.4 14.2 14.8
*Table developed using ASHRAE Duct Fitting Database (ASHRAE 2016): CD11-4 ( = 0.12 mm; = 1.204 kg/m3).
Fig. 27
Air Density for Example 8
ASHRAE Duct Fitting Database (ASHRAE 2016)
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21.24
2017 ASHRAE Handbook—Fundamentals (SI) abrasive materials. Where potentially explosive or radioactive materials are conveyed, the prebalanced system is mandatory because contaminants could accumulate at the balancing devices. To balance systems by increasing airflow, use Equation (42), which assumes that all ductwork has the same diameter and that fitting loss coefficients, including main and branch tee coefficients, are constant.
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Two pressure-balancing methods can be considered when designing industrial exhaust systems. One method uses balancing devices (e.g., dampers, blast gates) to obtain design airflow through each hood. The other approach balances systems by adding resistance to ductwork sections (i.e., changing duct size, selecting different fittings, increasing airflow). This self-balancing method is preferred, especially for systems conveying
Fig. 25 Economizer Duct System Shown (ASHRAE Guideline 16)
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Duct Design
21.25 Table 14
2
3
Section
Upstream Section
1
Example 8, Equal Friction Design
Fitting
5
6
7
8 Duct Length, m
9
10
Source
Source
Source
Source
Source
Source
Source
Source
Drawings*
DFDB
Drawing
DFDB
DFDB
Drawing
DFDB
DFDB
CR11-1 1520
1
1
1.5 450 500
6.8 22
0.00
Section Total Duct Transition (H1 = 450 mm, W1 = 500 mm, Do = 375 mm, L = 600 in., Theta1 = 7o, Theta2 = 12o Sized at Maximum Velocity of 15.2 m/s using CD11-3 Section Total 1
2
2
4
Duct Tee, 45° Entry, Main (Dc = 375, Ds = 350, Db = 200)
CD11-3/ CD11-1
375
13.8 0.06
SD4-2 93
CD11-1 SD5-12
1140
350
11.8
Licensed for single user. © 2017 ASHRAE, Inc.
0.06
CD11-1 SD5-12
6
0.14 0.14
6 760
300
21
10.8 57
6
CD11-1 SD5-12
8
0.13
4.5 380
225
3
Duct Tee,45° Entry, Branch (Dc = 375, Ds = 350, Db = 200) VAV Box
CD11-1 SD5-12
18
9.6 45
0.14
72
0.58 1.67 2.25
200
72
0.38 1.67 2.05
72
0.29 1.67 1.96
12.1
Section Total
4
5
Duct Tee,45° Entry, Branch (Dc = 350, Ds = 300, Db = 200) VAV Box
CD11-1 SD5-12
200
12.1
Section Total
6
7
Duct Tee,45° Entry, Branch (Dc = 300, Ds = 225, Db = 200) VAV Box
CD11-1 SD5-12
200
12.1
Section Total
8
9
Duct Elbow, Die Stamped, r/D = 1.5 Transition (D1 = 225 mm, Do = 200 mm, L = 300 mm, Theta = 5o ) VAV Box
CD11-1 CD3-1 SD4-1
148 159 11
1.5 380
162 173 11
1.5 380
6 24 11
1.5 380
7 28
0.14
Section Total
2
10 31
0.13
Section Total Duct Tee,45° Entry, Main (Dc = 300, Ds = 225, Db = 200)
6 32 21
6
69
Duct Tee, 45° Entry, Main (Dc = 350, Ds = 300, Db = 200)
0 1 26
6 1520
Section Total
4
11
Total Velocity Loss Pressure Pressure pv, Pa Coefficient C Loss, Pa
Source Duct, Rectangular (Fan Outlet)
Fan
4
ASHRAE Air Fitting Quantity, Duct Size, Velocity, Code L/s mm m/s
141 152 22
3 0.11 380
200
12.1
0.04 72
Section Total *Symbols in column 3 (Dc, Ds, etc.) are input associated with subject fitting. Symbols are shown on drawings below DFDB input/output for each fitting.
1.67 1.82
131 153
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21.26
2017 ASHRAE Handbook—Fundamentals (SI) Table 15 Example 8, Static Regain Design
2
3 Fitting
Section
Upstream Section
1
5
7
8
Duct Size, mm
6
Velocity, m/s
Duct Length, m
9
10
Source
Source
Source
Source
Source
Source
Source
Source
DFDB
Drawing
DFDB
DFDB
Drawing
DFDB
DFDB
Static Regain Calculation
1520
450 500
6.8
CR11-1
1
1.5
22
0.00
Duct CD11-1 Transition (H1 = 450 mm, 1520 W1 = 500 mm, Do = 375 mm, L = SD4-2 600 in., Theta1 = 7°, Theta2 = 12° Sized at Maximum Velocity of 15.2 m/s using CD11-3
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Duct Tee, 45° Entry, Main (Dc = 375, 4a Ds = 375, Db = 200)
6 375
26
13.8
0.06 93
0.06
CD11-1 SD5-12
6 1140
375
15 0.13
10.3 52
2
4bb
0.13
CD11-1 SD5-12
350
0.14
11.8
0.14
CD11-1 SD5-12
6 760
350
6bb
SD5-12
0.14
325
41
0.15
4
SD5-12
6 760
300
57
0.13
6
4.5 380
325
aSymbols
bIndicates
8
(69 – 57) – 28 = –16 –16
3 0.20
4.6 10
Section Total
7 28
SD5-12
(69 – 41) – 20 = 8
0.13
10.8
CD11-1
24
21
Section Total Duct Tee, 45° Entry, Main (Dc = 325, 8a Ds = 325, Db = 200)
6 20
CD11-1
(69 – 31) – 14 = 24
0.15
9.2
Section Total Duct Tee, 45° Entry, Main (Dc = 350, 6c Ds = 300, Db = 200)
4
14
6 760
–7
10
14 CD11-1
(93 – 69) – 31 = –7
0.14
7.9 31
4
10 31
Section Total Duct Tee, 45° Entry, Main (Dc = 350, Ds = 325, Db = 200)
(93 – 52) – 22 = 19 19
21
6 1140
69
4
7 22
Section Total Duct Tee, 45° Entry, Main (Dc = 350, 6a Ds = 350, Db = 200)
6 32
Section Total Duct Tee, 45° Entry, Main (Dc = 375, Ds = 350, Db = 200)
0 1
Section Total
2
Regain, Pa, [pv1 – pv3] – pt
Source
1
2
12
Source
Section Total
1
11
Velocity Loss Total Pressure Coeffi- Pressure cient pv , Loss, Pa C Pa
Drawingsa Duct, Rectangular (Fan Outlet) Fan
4
ASHRAE Air Fitting Quantity, Code L/s
0.20
2 5
in column 3 (Dc, Ds, etc.) are input associated with subject fitting. Symbols are shown on drawings below DFDB input/output for each fitting. size selected.
(41 – 10) – 5 = 26 26
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Duct Design
21.27 Table 15
2
3 Fitting
Section
Upstream Section
1
5
ASHRAE Air Fitting Quantity, Code L/s
7
8
Duct Size, mm
6
Velocity, m/s
Duct Length, m
9
10
Source
Source
Source
Source
Source
Source
Source
Source
DFDB
Drawing
DFDB
DFDB
Drawing
DFDB
DFDB
Static Regain Calculation
380
300
5.4
Tee, 45° Entry, Main (Dc = 325, 8b Ds = 300, Db = 200)
Duct Tee, 45° Entry, Main (Dc = 325, 8c Ds = 275, Db = 200)
CD11-1 SD5-12
4.5
4 0.17
0.17
CD11-1 SD5-12
4.5 380
275
6.4
Licensed for single user. © 2017 ASHRAE, Inc.
Duct
CD11-1 SD5-12
0.16
4.5 380
250
7.7
0.14
CD11-1 SD5-12
11
1.5 380
200
0.58
12.1
VAV Box
1.67 72
2.25
Section Total
4
5
Duct
CD11-1
Tee, 45° Entry, Branch (Dc = 350, Ds = 325, Db = 200)
SD5-12
1.5 380
200
11
12.1
0.38 1.67 72
2.05
Section Total
7
Duct
CD11-1
Tee, 45° Entry, Branch (Dc = 325, Ds = 250, Db = 200)
SD5-12
11
1.5 380
200
0.23
12.1
1.67 72
1.9
Section Total
9
Duct
CD11-1 CD3-1
Transition (D1 = 250 mm, Do = 200mm, L = 300 mm, Theta = 10°)
SD4-1
3
22 0.11
380
200
12.1
0.05
VAV Box
1.67 72
aSymbols
137 148
Elbow, Die Stamped, r/D = 1.5
Section Total
148 159
VAV Box
8
162 173
VAV Box
6
4 15
Duct
(41 – 20) – 10 = 11 11
0.14
29
3
3
11
Section Total
Tee, 45° Entry, Main (Dc = 375, Ds = 350, Db = 200)
21
0.16
10
Tee, 45° Entry, Main (Dc = 325, Ds = 250, Db = 200)
(41 – 14) –6 = 21
7
20
2
2 6
Section Total
8db
Regain, Pa, [pv1 – pv3] – pt
Source
14
6
12
Source
Section Total
6
11
Velocity Loss Total Pressure Coeffi- Pressure cient pv , Loss, Pa C Pa
Drawingsa Duct 6
4
Example 8, Static Regain Design (Continued)
1.83
132 154
in column 3 (Dc, Ds, etc.) are input associated with subject fitting. Symbols are shown on drawings below DFDB input/output for each fitting. bIndicates size selected.
(41 – 29) – 15 = –3 –3
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21.28
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 26
System Layout for Example 8
Table 16 Static Regain Iteration Process for Section 4 Iteration Calculation
Licensed for single user. © 2017 ASHRAE, Inc.
1 2
Equal Friction
Duct Size, mm
( pv1 – pv2) – pi
Remarks
375 350
19 –7
Solution
4a 4b
Table 18 System Unbalance
Table 17 Summary of System Duct Sizes
Static Regain
Path
Pt , Pa
Unbalance
Pt , Pa
Unbalance
A B C D
206 223 244 269
23% 17% 9% 0%
206 223 232 252
19% 12% 8% 0%
Size, mm Section 1 2 4 6 8 3 5 7 9
EF
SR
450 500 450 500 375 375 350 350 300 325 225 250 200 200 200 200 200 200 200 200
Remarks Fan outlet Initial section Sections affected by design method Terminal runout Terminal runout Terminal runout Terminal runout
Fig. 28 Sizing Section 2 for EF and SR Design Examples Qc = Qd (Ph /Pl )0.5
(42)
where Qc = airflow rate required to increase Pl to Ph, L/s Qd = total airflow rate through low-resistance duct run, L/s Ph = absolute value of pressure loss in high-resistance ductwork section(s), Pa Pl = absolute value of pressure loss in low-resistance ductwork section(s), Pa
For systems conveying particulates, use elbows with a large centerline radius-to-diameter ratio (r/D), greater than 1.5 whenever possible. If r/D is 1.5 or less, abrasion in dust-handling systems can reduce the life of elbows. Elbows are often made of seven or more gores, especially in large diameters. For converging flow fittings, a 30° entry angle is recommended to minimize energy losses and abrasion in dust-handling systems. For the entry loss coefficients of hoods and equipment for specific operations, see Chapter 32 of the 2015 ASHRAE Handbook—HVAC Applications and ACGIH (2016). Commentary: Fitting CD11-5 of the ASHRAE Duct Fitting Database (ASHRAE 2016) is available for sizing minimumvelocity ducts conveying particulates. Example 9. For the metalworking exhaust system in Figures 30 and 31, size the ductwork and calculate fan static pressure requirement for an industrial exhaust designed to convey granular materials. Pressure-balance the system by changing duct sizes and adjusting airflow rates. Minimum particulate transport velocity for the chipping and grinding table ducts
Note: Diameter rounded up to next 25 mm increment.
Fig. 29 EF Design: Sizing Sections 4, 6, and 8 Knowing Design Friction Rate (Section 4 Shown) (sections 1 and 5, Figure 31) is 20 m/s. For ducts associated with the grinder wheels (sections 2, 3, 4, and 5), minimum duct velocity is 23 m/ s. Ductwork is galvanized steel, with absolute roughness of 0.09 mm. Use sizes for which spiral duct is available (80, 100, 125, 140, 150, 160, 180, 200, 224, 250, 280, 300, 315, 355, 400, 450, 500, 560, 600, 630, 710, 800, 900, 1000, 1120, 1250, 1300, 1400, 1500, 1600, 1800, 2000, 2100, 2200, 2300, 2400, and 2500 mm). The building is one story, and the design wind velocity is 9 m/s. For the stack, use design J shown in Figure 2 in Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications for complete rain protection; stack height, determined by calculations from Chapter 45, is 4.9 m above the roof. This height is based on minimized stack downwash; therefore, the stack discharge velocity must exceed 1.5 times the design wind velocity. Solution: The following table summarizes initial duct sizes and transport velocities for contaminated ducts upstream of the collector.
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Duct Design
21.29
Fig. 30 Metalworking Exhaust System for Example 9 The 22.8 m/s velocity in sections 2 and 3 is acceptable because the transport velocity is not significantly lower than 23 m/s. For the next available duct size (160 mm diameter), the duct velocity is 28.8 m/s, significantly higher than 23 m/s. Duct Design Airflow, Transport Section L/s Velocity, m/s 1 2, 3 4 5
850 290 each 580 1430
20 23 23 23
Duct Diameter, Duct Velocity, mm m/s 224 125 180 180
21.6 23.6 22.8 23.2
Design calculations up through the junction after sections 1 and 4 are summarized as follows: Design No. D1, mm 1 2 3
224 200 180
∆ p1, Pa
∆ p2+4, Pa
Imbalance, ∆ p1– ∆ p2+4
411 762 1320
794 850 712
–383 –88 +609
Q1 = 850 L/s Q2 = 290 L/s; D2 = 125 mm dia.
Q3 = 290 L/s; D3 = 125 mm dia. Q4 = 850 L/s; D4 = 180 mm dia.
For (initial) design 1, the imbalance between section 1 and section 2 (or 3) is 383 Pa, with section 1 requiring additional resistance. Decreasing section 1 duct diameter by manufactured spiral duct sizes results in the least imbalance, 88 Pa, when the duct diameter is 200 mm (design 2). Because section 1 requires additional resistance, estimate the new airflow rate using Equation (42): Qc,1 = 850(850/762)0.5 = 900 L/s At 900 L/s flow in section 1, 130 Pa imbalance remains at the junction of sections 1 and 4. By trial-and-error solution, balance is attained when the flow in section 1 is 860 L/s. The duct between the collector and fan inlet is 355 mm round to match the fan inlet (340 mm diameter).
Fig. 31
System Schematic with Section Numbers for Example 9
To minimize downwash, the stack discharge velocity must exceed 13.5 m/s, 1.5 times the design wind velocity (9 m/s) as stated in the problem definition. Therefore, the stack is 355 mm round, and the stack discharge velocity is 14.5 m/s. Table 19 summarizes the system losses by sections. The straight duct friction factor and pressure loss were calculated by Equations (18) and (19). Table 20 lists fitting loss coefficients and input parameters necessary to determine the loss coefficients. The fitting loss coefficients are from the ASHRAE Duct Fitting Database. Figure 32 shows a pressure grade line of the system. Fan total pressure, calculated by Equation (15), is 1992 Pa. To calculate the fan static pressure, use Equation (17): Ps = 1992 – 192 = 1800 Pa where 192 Pa is the fan outlet velocity pressure. The fan airflow rate is 1440 L/s, and its outlet area is 0.081 m3 (260 by 310 mm). Therefore, the fan outlet velocity is 17.9 m/s.
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21.30
2017 ASHRAE Handbook—Fundamentals (SI) Table 19 Total Pressure Loss Calculations by Sections for Example 9 Duct Summary of Airflow, Duct Size, Velocity, Velocity Length,b Fitting Loss L/s mm m/s Pressure, Pa m Coefficientsc
Duct Sectiona Duct Element 1 2, 3 4 5 — 6 7
Duct Fittings Duct Fittings Duct Fittings Duct Fittings Collector,e fabric Duct Fittings Duct Fittings
aSee
Figure 29. bDuct lengths are to fitting centerlines.
860 860 290 290 580 580 1440 1440 1440 1440 1440 1440 1440
200 — 125 — 180 — 280 — — 355 — 355 —
Table 20. dDuct pressure based on a 0.09 mm absolute roughness factor.
Licensed for single user. © 2017 ASHRAE, Inc.
2,3
5 6 7
aFrom
bFrom
7.32 — 2.7 — 3.84 — 2.7 — — 3.0 — 14.0 —
— 1.09 — 1.06 — 0.51 — 0.23 — — 0.03 — 1.77
40 — 54 — 32 — 18.5 — — 6 — 6 —
293 492 146 356 123 160 50 76 750 18 4 84 225
785 502 283 126 750 22 309
eCollector manufacturers set fabric bag cleaning mechanism to actuate at a pressure dif-
ference of 750 Pa between inlet and outlet plenums. Pressure difference across clean media is approximately 400 Pa.
Loss Coefficient Summary by Sections for Example 9
Type of Fitting
ASHRAE Fitting No.a
Hoodb
1
4
— 451 — 336 — 313 — 329 — — 127 — 127
cSee
Table 20 Duct Fitting Section Number
27.4 27.4 23.6 23.6 22.8 22.8 23.4 23.4 — 14.5 14.5 14.5 14.5
Duct Pressure Total Pressure Section Pressure Loss, Pa/md Loss, Pa Loss, Pa
Parameters
Loss Coefficient
— Hood face area: 0.9 by 1.2 m 1 2 Elbow CD3-10 90°, 7 gore, r/D = 2.5 4 Capped wye (45°), with 45° elbow ED5-6 Ab /Ac = 1 5 Wye (30°), main ED5-1 Qs /Qc = 0.60, As /Ac = 0.510, Ab /Ac = 0.413 Summation of Section 1 loss coefficients .................................................................................................................................................. — Type hood: For double wheels, dia. = 560 mm each, 6 Hoodc wheel width = 100 mm each; type takeoff: tapered 7 Elbow CD3-12 90°, 3 gore, r /D = 1.5 8 Symmetrical wye (60°) ED5-9 Qb /Qc = 0.5, Ab1 /Ac = 0.482, Ab2 /Ac = 0.482 Summation of Sections 2 and 3 loss coefficients ....................................................................................................................................... 9 Elbow CD3-10 90°, 7 gore, r /D = 2.5 10 Elbow CD3-13 60°, 3 gore, r /D = 1.5 5 Wye (30°), branch ED5-1 Qb /Qc = 0.40, As /Ac = 0.510, Ab /Ac = 0.413 Summation of Section 4 loss coefficients .................................................................................................................................................. 11 Exit, conical diffuser to collector ED2-1 L = 600 mm, L/Do = 2.14, A1/Ao 21 Summation of Section 5 loss coefficients .................................................................................................................................................. 12 Entry, bellmouth from collector ER2-1 r/D1 = 0.21, r = 75 mm, Co = 4.69 Summation of Section 6 loss coefficients .................................................................................................................................................. SD4-2 Fan outlet size: 260 by 310 mm; L = 460 mm 13 Diffuser, fan outletd 14 Capped wye (45°), with 45° elbow ED5-6 Ab /Ac = 1 15 Stackhead SD2-6 De /D = 1 Summation of Section 7 loss coefficients ..................................................................................................................................................
ASHRAE Duct Fitting Database. Industrial Ventilation (ACGIH 2016, Figure VS-80-19).
0.25 0.11 0.61 (Cb) 0.12 (Cs) 1.09 0.40 0.34 0.32 1.06 0.11 0.19 0.21 0.51 0.23 0.23 0.03 0.03 0.16 0.61 1.00 1.77
(Cb)
(Cb)
(C1) (Cb)
cFrom
Industrial Ventilation (ACGIH 2016, Figure VS-80-11). specified: Industrial exhauster for granular materials: 530 mm wheel diameter, 340 mm inlet diameter, 260 by 310 mm outlet, 6 kW motor.
dFan
Hood suction for the chipping and grinding table hood is 560 Pa, calculated by Equation (5) from Chapter 32 of the 2015 ASHRAE Handbook—HVAC Applications [Ps,h = (1 + 0.25)(451) = 560 Pa, where 0.25 is hood entry loss coefficient Co, and 451 is duct velocity pressure Pv a few diameters downstream from the hood]. Similarly, hood suction for each grinder wheel is 170 Pa P2,3 = (1 + 0.4)(336) = 470 Pa where 0.4 is the hood entry loss coefficient, and 336 Pa is the duct velocity pressure.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore.
Abushakra, B., I.S. Walker, and M.H. Sherman. 2004. Compression effects on pressure loss in flexible HVAC ducts. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 10(3):275-289. ACCA. 2014. Manual D—Residential duct systems. Air Conditioning Contractors of America, Washington, DC. ACGIH. 2016. Industrial ventilation: A manual of recommended practice for design, 29th ed. American Conference of Governmental Industrial Hygienists, Lansing, MI. AHRI. 2008. Procedure for estimating occupied space sound levels in the application of air terminals and air outlets. Standard 885-2008 with Addendum 1. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AMCA. 2011a. Fans and systems. AMCA Publication 201-02 (R2011). Air Movement and Control Association International, Arlington Heights, IL. AMCA. 2011b. Field performance measurement of fan systems. AMCA Publication 203-90 (R2011). Air Movement and Control Association International, Arlington Heights, IL.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Duct Design
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 32 Total Pressure Grade Line for Example 9 AMCA. 2016. Laboratory methods for testing fans for certified aerodynamic performance rating. ANSI/AMCA Standard 210-16. Also ANSI/ ASHRAE/AMCA Standard 51-16. AMCA. 2012. Laboratory methods of testing dampers for rating. ANSI/ AMCA Standard 500-D-12. Air Movement and Control Association International, Arlington Heights, IL. AMCA. 2015. Laboratory method of testing louvers for rating. ANSI/ AMCA Standard 500-L-12 (R2015). Air Movement and Control Association International, Arlington Heights, IL. AMCA. 2013. Certified ratings program—Product rating manual for air control. AMCA Publication 511-10 (Rev. 2013). Air Movement and Control Association International, Arlington Heights, IL. ASHRAE. 2016. ASHRAE duct fitting database, v. 6.00.05.10. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2016. ASHRAE. 2007. Energy-efficient design of low-rise residential buildings. ANSI/ASHRAE Standard 90.2-2007. ASHRAE. 2015. Energy conservation in existing buildings. ANSI/ ASHRAE/IES Standard 100-2015. ASHRAE. 2008. Measurement, testing, adjusting and balancing of building heating, ventilation and air-conditioning systems. ANSI/ASHRAE Standard 111-2008. ASHRAE. 2016. Method of testing HVAC air ducts and fittings. ANSI/ ASHRAE/SMACNA Standard 126-2016. ASHRAE. 2016. Methods of testing air terminal units. ANSI/ASHRAE Standard 130-2016. ASHRAE. In development. Method of test to determine leakage of operating HVAC air distribution systems. ANSI/ASHRAE Proposed Standard 215. ASHRAE. 2014. Selecting outdoor, return, and relief dampers for air-side economizer systems. ASHRAE Guideline 16-2014. Behls, H.F. 1971. Computerized calculation of duct friction. Building Science Series 39, p. 363. National Institute of Standards and Technology, Gaithersburg, MD. Brown, R.B. 1973. Experimental determinations of fan system effect factors. In Fans and systems, ASHRAE Symposium Bulletin LO-73-1, Louisville, KY (June). CEN. 1998. Ventilation for buildings—Sheet metal air ducts and fittings with rectangular cross-section—Dimensions. European Standard EN 1505-1998. European Committee for Standardization, Brussels. CEN. 2007. Ventilation for buildings—Sheet metal air ducts and fittings with circular cross-section—Dimensions. European Standard EN 15062007. European Committee for Standardization, Brussels. CEN. 2004. Ventilation for buildings—Ductwork—Measurement of ductwork surface area. European Standard EN 14239-2004. European Committee for Standardization, Brussels.
21.31 Clarke, M.S., J.T. Barnhart, F.J. Bubsey, and E. Neitzel. 1978. The effects of system connections on fan performance. ASHRAE Transactions 84(2): 227-263. Colebrook, C.F. 1938-1939. Turbulent flow in pipes, with particular reference to the transition region between the smooth and rough pipe laws. Journal of the Institution of Civil Engineers 11:133. Culp, C.H. 2011. HVAC flexible duct pressure loss measurements. ASHRAE Research Project RP-1333, Final Report. Farquhar, H.F. 1973. System effect values for fans. In Fans and systems, ASHRAE Symposium Bulletin LO-73-1, Louisville, KY (June). Griggs, E.I., and F. Khodabakhsh-Sharifabad. 1992. Flow characteristics in rectangular ducts. ASHRAE Transactions 98(1):116-127. Griggs, E.I., W.B. Swim, and G.H. Henderson. 1987. Resistance to flow of round galvanized ducts. ASHRAE Transactions 93(1):3-16. Heyt, J.W., and M.J. Diaz. 1975. Pressure drop in flat-oval spiral air duct. ASHRAE Transactions 81(2):221-232. Huebscher, R.G. 1948. Friction equivalents for round, square and rectangular ducts. ASHVE Transactions 54:101-118. Hutchinson, F.W. 1953. Friction losses in round aluminum ducts. ASHVE Transactions 59:127-138. Idelchik, I.E., M.O. Steinberg, G.R. Malyavskaya, and O.G. Martynenko. 1994. Handbook of hydraulic resistance, 3rd ed. CRC Press/Begell House, Boca Raton. Jones, C.D. 1979. Friction factor and roughness of United Sheet Metal Company spiral duct. United Sheet Metal, Division of United McGill Corp., Westerville, OH (August). Based on data in Friction loss tests, United Sheet Metal Company Spiral Duct, Ohio State University Engineering Experiment Station, File T-1011, September 1958. Klote, J.H., J.A. Milke, P.G. Tumbull, A. Kashef, and M.J. Ferreira. 2012. Handbook of smoke control engineering. ASHRAE. Kulkarni, D., S. Khaire, and S. Idem. 2009. Pressure loss of corrugated spiral duct. ASHRAE Transactions 115(1). Madison, R.D., and W.R. Elliot. 1946. Friction charts for gases including correction for temperature, viscosity and pipe roughness. ASHVE Journal (October). McGill. 1988. Round vs. rectangular duct. Engineering Report 147, United McGill Corp. (contact McGill Airflow Technical Service Department), Westerville, OH. McGill. 1995. Flat oval vs. rectangular duct. Engineering Report 150, United McGill Corp. (contact McGill Airflow Technical Service Department), Westerville, OH. Meyer, M.L. 1973. A new concept: The fan system effect factor. In Fans and Systems, ASHRAE Symposium Bulletin LO-73-1, Louisville, KY (June). Moody, L.F. 1944. Friction factors for pipe flow. ASME Transactions 66: 671. NFPA. 2008. Fire protection handbook, 20th ed. National Fire Protection Association. Quincy, MA. NFPA. 2015. Installation of air-conditioning and ventilating systems. ANSI/ NFPA Standard 90A. National Fire Protection Association, Quincy, MA. Osborne, W.C. 1966. Fans. Pergamon, London. Paulauskis, J.A. 2016. Understanding duct rumble. ASHRAE Journal 58(12):40-45. Schaffer, M.E. 2011. A practical guide to noise and vibration control for HVAC systems (SI), 2nd ed. ASHRAE. Swim, W.B. 1978. Flow losses in rectangular ducts lined with fiberglass. ASHRAE Transactions 84(2):216. Swim, W.B. 1982. Friction factor and roughness for airflow in plastic pipe. ASHRAE Transactions 88(1):269. Taylor. S.J., and J. Stein. 2004. Sizing VAV boxes. ASHRAE Journal 46(3): 30-35. Tsal, R.J., and M.S. Adler. 1987. Evaluation of numerical methods for ductwork and pipeline optimization. ASHRAE Transactions 93(1):17-34. Tsal, R.J., H.F. Behls, and R. Mangel. 1988. T-method duct design, part I: Optimization theory; Part II: Calculation procedure and economic analysis. ASHRAE Transactions 94(2):90-111. Tsal, R.J., H.F. Behls, and R. Mangel. 1990. T-method duct design, part III: Simulation. ASHRAE Transactions 96(2). UL. Published annually. Building materials directory. Underwriters Laboratories, Northbrook, IL. UL. Published annually. Fire resistance directory. Underwriters Laboratories, Northbrook, IL. UL. 2013. Fire dampers. ANSI/UL Standard 555, 7th ed. Underwriters Laboratories, Northbrook, IL.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
21.32
2017 ASHRAE Handbook—Fundamentals (SI)
UL. 2014. Smoke dampers. ANSI/UL Standard 555S, 5th ed. Underwriters Laboratories, Northbrook, IL. Wright, D.K., Jr. 1945. A new friction chart for round ducts. ASHVE Transactions 51:303-316.
BIBLIOGRAPHY
Licensed for single user. © 2017 ASHRAE, Inc.
Abushakra, B., I.S. Walker, and M.H. Sherman. 2002. A study of pressure losses in residential air distribution systems. Proceedings of the
ACEEE Summer Study 2002, American Council for an Energy Efficient Economy, Washington, D.C. LBNL Report 49700. Lawrence Berkeley National Laboratory, CA. Carrié, F.R., J. Andersson, and P. Wouters. 1999. Improving ductwork—A time for tighter air distribution systems. Air Infiltration and Ventilation Centre, Coventry, U.K. Hyderman, M., S. Taylor, and J. Stein. 2009. Advanced variable air volume VAV systems design guide, 3rd edition. Pacific Gas and Electric Company.
Related Commercial Resources
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Related Commercial Resources CHAPTER 22
PIPE DESIGN FUNDAMENTALS................................................................... 22.1 Codes and Standards ............................................................... 22.1 Design Considerations............................................................. 22.1 General Pipe Systems .............................................................. 22.1 Design Equations ..................................................................... 22.5 Sizing Procedure.................................................................... 22.10 Pipe-Supporting Elements ..................................................... 22.10 Pipe Expansion and Flexibility.............................................. 22.11 Pipe Bends and Loops............................................................ 22.12 PIPE AND FITTING MATERIALS........................................ 22.14 Pipe ........................................................................................ 22.14
Fittings ................................................................................... Joining Methods ..................................................................... Expansion Joints and Expansion Compensating Devices...... APPLICATIONS..................................................................... Water Piping .......................................................................... Service Water Piping.............................................................. Steam Piping .......................................................................... Low-Pressure Steam Piping................................................... Steam Condensate Systems .................................................... Gas Piping.............................................................................. Fuel Oil Piping.......................................................................
HIS CHAPTER discusses pipe systems, materials, design, installation, supports, stress calculations, pipe expansion and flexibility, bends and loops, and application of pipe systems commonly used for heating, air conditioning, refrigeration, and service water. When selecting and applying components; applicable local codes, state or provincial codes, and voluntary industry standards (some of which have been adopted by code jurisdictions) must be followed. Further details on specific piping systems can be found in application-specific chapters of the ASHRAE Handbook.
• Pressure and temperature of the fluid. • External environment of the pipe: outdoor installations deal with temperature extremes, environmental contaminants, and ultraviolet radiation. Other environments could contain caustic chemicals. Soil can contain elements that can be corrosive to underground pipe systems. • Installation cost. • Pipe’s resistance to chemical attack from the fluid.
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T
1. 1.1
FUNDAMENTALS CODES AND STANDARDS
The following organizations in the United States issue codes and standards for piping systems and components: ASME ASTM NFPA ICC MSS
American Society of Mechanical Engineers American Society for Testing and Materials National Fire Protection Association International Code Council Manufacturers Standardization Society of the Valve and Fittings Industry, Inc. AWWA American Water Works Association Parallel federal specifications also have been developed by government agencies and are used for many public works projects. Chapter IV of ASME Standard B31.9 lists applicable U.S. codes and standards for HVAC piping. In addition, it gives requirements for safe design and construction of piping systems for building heating and air conditioning. ASME Standard B31.5 gives similar requirements for refrigerant piping.
1.2
DESIGN CONSIDERATIONS
Pipes are conduits in which fluids [compressible (e.g., air, steam) and noncompressible (e.g., water)] flow in a system, in response to a pressure differential. Piping system designers should assess the following aspects: • Code requirements. • Load: the amount of energy or fluid to be moved through the pipe to where it is needed; determination of load is not covered in this chapter (see Chapters 16 to 18 for information on load calculations). • Working fluid and fluid properties in the pipe. The preparation of this chapter is assigned to TC 6.1, Hydronic and Steam Equipment and Systems.
22.1 Copyright © 2017, ASHRAE
22.18 22.18 22.20 22.22 22.22 22.23 22.29 22.30 22.32 22.35 22.36
When designing a fluid flow system, two related but distinct concerns emerge: sizing the pipe and determining the flow/pressure relationship. The two are often confused because they can use the same equations and design tools. Nevertheless, they should be determined separately. This chapter focuses on sizing the pipe during the design phase, and to this end presents design charts and tables for specific fluids in addition to the equations that describe fluid flow in pipes. Once a system has been sized, it should be analyzed with more detailed methods of calculation to determine the pump pressure, if applicable, required to achieve the desired flow. Computerized methods are well suited to handling the details of calculating losses around an extensive system. Not discussed in detail in this chapter, but of potentially great importance, are physical and chemical considerations such as pipe and fitting design; materials; and joining methods appropriate for working pressures and temperatures encountered, as well as resistance to chemical attack by the fluid. For more information, see Eshbach (2009), Heald (2002), and Nayyar (1999). For fluids not included in this chapter or for piping materials of different dimensions, manufacturers’ literature frequently supplies pressure drop charts. The Darcy-Weisbach equation, with the Moody chart or Colebrook equation, can be used as an alternative to pressure drop charts or tables.
1.3
GENERAL PIPE SYSTEMS
Metallic Pipe Systems Each HVAC system and, under some conditions, portions of a system require a study of the conditions of operation to determine suitable materials. For example, because the static pressure of water in a high-rise building is higher in the lower levels than in the upper levels, a heavier pipe or different materials may be required for different vertical zones. Table 1 lists some typical systems and materials used for heating and air-conditioning metallic piping. The list is not all inclusive, because piping systems are constantly being developed. The pressure and temperature rating of each component selected must be
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22.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Common Applications of Pipe, Fittings, and Valves for Heating and Air Conditioning
Application Size, mm Chilled water
Material
Type
51 Steel Type F (CW) Schedule 40 62.5 to 305 Steel A or B, Type E Schedule 40 (ERW)
Joint Type
Fitting Material
Thread Weld Flange
Cast iron Wrought steel Wrought steel Cast iron Cast iron Wrought or cast Cu
Copper, hard or soft Type K or L
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Copper, hard
10 to 25
PEX (barrier)
13 to 152
PE
Heating and 51 and recirculating smaller 6 to 305
Steel Type F (CW) Steel B Type E (ERW)
Solder Flared (soft) Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Weld Type M Solder Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Weld SDR-9 Crimp Clamp Expansion Compression Push fit Proprietary Schedule 40,f 80, Thermal fusion, SDR compression
Schedule 40 Schedule 40
Copper, hard or soft Type K or L
6 to 305
Steam and condensate
Copper, hard
Type M
10 to 25
PEX (barrier)
SDR-9
51 and smaller
Steel Type F (CW) or S
Schedule 40d
Steel B Type E (ERW) or S
Schedule 40d
Steel B Type E (ERW) or S
Schedule 80
51 to 305
Steel B Type E (ERW) or S
Schedule 40
Thread Weld Flange
Solder Braze Flared (soft) Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Weld Solder Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Weld Crimp Clamp Expansion Compression Push fit Proprietary Thread Thread Socket Thread Thread Socket Thread Socket Thread Socket Weld Flange
Systemg Class (When Temperature, Maximum Pressure at Applicable) °C Temperature,a,b kPa 121 121 121 121 121 38
862 2758 1724 1207 2758 2586 Type K soft 4378 Type K hard 1724 Type L soft 2999 Type L hard
Wrought or cast Cu
38
Wrought or cast Cu
38
1724 Type L soft 2586 Type K soft 2724 Type M hard
Wrought or cast Cu
38
1586 Type M soft
Bronze Brass Copper Engineered plastic
23
1000
PE
49 (60 limit for Varies with pipe wall thickness, some applica- grade, schedule, size. Check manufacturer’s documentation for design tions) ratings 207 to 758 at 54°C
Cast iron Wrought steel Wrought steel Cast iron Cast iron Wrought or cast Cu
125 Standard 150 125 250
125 Standard 150 125 250
121 121 121 121 121 93
862 2758 1724 862 2758 2069 Type K soft 4378 Type K hard 1413 Type L soft 2999 Type L hard
Wrought or cast Cu
93
Wrought or cast Cu
93
2069 Type K soft 1413 Type L soft 2724 Type M hard
Wrought or cast Cu
93
1379 Type M soft
Bronze Brass Copper Engineered plastic
93
545
Cast iron Malleable iron Forged steel Cast iron Malleable iron Forged steel Cast iron
125 150 3000 125 150 3000 250
621 621 621 690 862 2758 1379
Malleable iron Forged steel Wrought steel Wrought steel
300 3000 Standard 150
1724 2758 1724 1379
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Pipe Design Table 1
Application Size, mm
Common Applications of Pipe, Fittings, and Valves for Heating and Air Conditioning (Continued)
Material
Type
Joint Type
Fitting Material
Systemg Class (When Temperature, Maximum Pressure at Applicable) °C Temperature,a,b kPa 125 XS 300 250
Steel B Type E (ERW) or S
Schedule 80
Weld Flange
Cast iron Wrought steel Wrought steel Cast iron
Ground6 to 51 source heat pump 10 to 25
Copper, hard or soft PEX (barrier)
Type L or ACR
Flared or brazed
Wrought or cast Cu
93
SDR-9
Crimp Clamp Expansion Compression Push fit Proprietary
Bronze Brass Copper Engineered plastic
82
Refrigerant
Schedule 40
Weld
10 to 105
Steel B Type E (ERW) Copper, hard
Type L or ACR
Braze
Wrought or Forged Cu
93
2999 Type L hard, 4655 ACR soft
6 to 305
Copper, hard or soft Type K or L
Solder Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanically formed Braze Weld Solder Braze Crimp Clamp Expansion Compression Push fit Proprietary Thermal fusion, compression
Wrought or cast Cu
38
2551 Type K soft 4378 Type K hard 1724 Type L soft 2999 Type L hard
Wrought or cast Cu
38
Wrought or cast Cu Wrought or cast Cu Bronze Brass Copper Engineered plastic
38 38 23
2551 Type K soft 1724 Type L soft 3448 Type ACR hard 2000 Type ACR Soft 1000
PE
49 (60limit for Depends on pipe, grade, schedule, some applica- size. Generally 207 to 758 at 54°C tions) 49 Depends on pipe, grade, schedule, size. Generally 441 for SDR 11 at 49°C
Natural gas and LP
Licensed for single user. © 2017 ASHRAE, Inc.
22.3
690 4826 3448 1379 1413 Type L soft, 2999 Type L hard, 4655 ACR soft, 3448 ACR hard 690
Wrought steel
10 to 105
Copper, hard
ACR
10 to 25
PEX
SDR-9
13 to 152
PE
Schedule 40, 80, SDR
13 to 152
HDPE
SDR
Thermal fusion, compression
HDPE
Black Steel, B Type E (ERW) or S (seamless)
Schedule 40
Thread or weld
Copper, hard or soft
Type K or L
Black malleable iron 150 Wrought steel weld Forged steel flanges 150 Wrought or cast Cu
38
2069 Type K soft 4378 Type K hard 1724 Type L soft 2999 Type L hard
Wrought or cast Cu Wrought or cast Cu
38 38
2069 Type K soft, 1724 Type L soft 2724 Type M hard
ABS
71 limit
HDPE
49
Depends on pipe class: approximately 345 at 71°C Depends on pipe, grade, schedule, size. Generally 441 for SDR 11 at 49°C
Fuel oil, 51 to 305 aboveground 6 to 305
Copper, hard
6 to 305
ABS
13 to 152
HDPE
Compressed 62.5 and air smaller 62.5 10 to 105
13 to 102
Solder Flared (soft) Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze or weld Type M Solder Braze Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Schedule 40,f 80, Solvent weld, thread, SDR flange SDR-9 Thermal fusion, compression
Black steel
Schedule 40
Thread
Black malleable iron 150
177
Black steel Copper, hard
Schedule 40 ACR
Black malleable iron 150 Wrought or cast Cu
177 93
ABS HDPE
Schedule 40 Schedule 40, 80, SDR
Flange or weld Solder Flared (soft) Mechanical formed Braze Solvent weld
ABS HDPE
93 23
4655 ACR soft 3448 ACR hard 4655 ACR hard 1276
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22.4
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Common Applications of Pipe, Fittings, and Valves for Heating and Air Conditioning (Continued)
Application Size, mm 10 to 25 Potable water, 6 to 305 inside building
Licensed for single user. © 2017 ASHRAE, Inc.
6 to 305
Material
Type
PEX
SDR-9
Steel, galvanized Copper, hard or soft
Copper, hard
Fitting Material
Schedule 40
Thread
Type K or L
Galv. cast iron 150 Galv. cast iron 150 Wrought or cast Cu
Solderc Flared (soft) Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Wrought or cast Cu Weld Wrought or cast Cu Solderc Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Wrought or cast Cu Weld CPVC
Type M
CPVC
Schedule 40,f 80
10 to 25
PEX
SDR-9
13 to 152
PE
13 to 152
PP
6 to 305
Copper, hard
38 38
2551 Type K soft 1724 Type L soft 2724 Type M hard
38
1586 Type M soft
99 Limit, 93 operating 38
1000
Schedule 40,f 80, Thermal fusion, flange, SDR Threade
PP
49 (60 limit for Depends on pipe, grade, schedule, some applica- size generally 207 to 758 at 54°C tions) 82 345
Class 50 Type K or L
Cast iron Wrought or cast Cu
24 38
1724 2551 Type K soft 4378 Type K hard 1724 Type L soft 2999 Type L hard
Wrought or cast Cu
38
2551 Type K soft 1724 Type L soft
Bronze Wrought or cast Cu
38 38
2724 Type K hard
Wrought or cast Cu
38
1586 Type M soft
Bronze Brass Copper Engineered plastic
23
1000
PVC
66 limit, 60 operating
545 to 724, depending on schedule and size
Wrought or cast Cu ABS
38 71 limit
PVC
66 limit, 60 operating
1724 DWV hard Depends on pipe class: approximately 345 at 71°C 545 to 724, depending on schedule and size
Type M
Mechanical joint Solderc Flared (soft) Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Weld Flange Solderc Rolled groove (51 to 203) Press-connect (13 to 102) Push connect (13 to 51) Mechanical formed Braze Weld Crimp Clamp Expansion Compression Push fit Proprietary Solvent weld, thread,f thermal weld
SDR-9
6 to 508
PVC
Schedule 40, 80, 120, SDR
Copper, hard ABS
DWV Solder Schedule DWV, Solvent weld, thread, flange 40,f 80, SDR Schedule 40,f 80, Solvent weld, thread, 120, SDR thermal weld
PV
862 1034 2551 Type K soft 4378 Type K hard 1724 Type L soft 2999 Type L hard
Bronze Brass Copper Engineered plastic
PEX
32 to 508
38 38 38
Crimp Clamp Expansion Compression Push fit Proprietary Schedule 40,f 80, Thermal fusion, SDR compression
10 to 25
32 to 203 Drainage, waste, and 32 to 305 vent (DWV)
aMaximum
Joint Type
13 to 203
Through 152 Ductile iron Water services, under- 6 to 305 Copper, hard or ground soft
Systemg Class (When Temperature, Maximum Pressure at Applicable) °C Temperature,a,b kPa
allowable working pressures have been derated in this table. Higher system pressures can be used for lower temperatures and smaller pipe sizes. Pipe, fittings, joints, and valves must all be considered. bTemperature and pressure relationships can vary based on pipe material composition, size, class, and schedule. cLead- and antimony-based solders are prohibited for potable water systems. Brazing should be used.
PE
dPiping
codes typically require thicker-walled pipe for threaded joints to maintain corrosion allowance and pressure ratings. plumbing codes require both hot and cold water piping to have a 689 kPa at 82°C rating. fThreads are not recommended on Schedule 40 plastic pipe. gDesigner should confirm that all materials are suitably rated for intended operation. eAll
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Pipe Design
22.5 Table 2
Manufacturers’ Recommendationsa,b for Plastic Materials
Cold-water service Hot (60°C) water Potable-water service Drain, waste, and vent (DWV)R Demineralized water Deionized water Salt water Heating (93°C) hot water Natural gas Compressed air Sunlight and weather resistance Underground service Food handling R = Recommended N = Not recommended — = Insufficient information
PVC
CPVC
HDPE
PEX
PP
ABS
PVDF
RTRP
R N R R R R R N N N N R R
R R R
R R R — — — R N R R R R —
R R R — — — R N R R R — —
R R R R R R R N N N — R R
R R R R R R R N N R R R R
R R R — R R — — — — R — R
R R R — — R R R — — R R R
R R R N N N N R R
aBefore
selecting material, check availability of suitable range of sizes and fittings and of satisfactory joining method. Also have manufacturer verify the best material for purpose intended. local building codes for compliance of materials listed.
bConsult
L V 2 p = f ---- --------- D 2
Licensed for single user. © 2017 ASHRAE, Inc.
considered; the lowest rating establishes the operating limits of the system.
Nonmetallic (Plastic) Pipe Systems Nonmetallic pipe is used in HVAC and plumbing. Plastic is light, generally inexpensive, and corrosion resistant. Plastic also has a low “C” factor (i.e., its surface is very smooth), resulting in lower pumping power requirements and smaller pipe sizes. Plastic pipe’s disadvantages include rapid loss of strength at temperatures above ambient and a high coefficient of linear expansion. The modulus of elasticity of plastics is low, resulting in a short support span. Some jurisdictions do not allow certain plastics in buildings because of toxic products emitted during fires. Plenum-rated plastic and insulation may be used to achieve a plenum rating; check with the authority having jurisdiction (AHJ). Table 2 lists nonmetallic materials used for service water and heating and air-conditioning piping. The pressure and temperature rating of each component selected must be considered; the lowest rating establishes the operating limits of the system.
Special Systems Some piping systems are governed by separate codes or standards. Generally, any failure of the piping in these systems is dangerous to the public, so some local areas have adopted laws enforcing the codes, such as the following: • Boiler piping: ASME Standard B31.1 and the ASME Boiler and Pressure Vessel Code (Section I) specify piping inside code-required stop valves on boilers that operate above 100 kPa (gage) with steam, or above 1.1 MPa or 120°C with water. These codes require fabricators and contractors to be certified for such work. The field or shop work must also be inspected while it is in progress, by inspectors commissioned by the National Board of Boiler and Pressure Vessel Inspectors. • Refrigeration piping: ASHRAE Standard 15 and ASME Standard B31.5. • Plumbing systems: Local codes. • Sprinkler systems: NFPA Standard 13. • Fuel gas: NFPA Standard 54/ANSI Standard Z223.1.
1.4 DESIGN EQUATIONS Darcy-Weisbach Equation Pressure drop caused by fluid friction in fully developed flows of all well-behaved (Newtonian) fluids is described by the DarcyWeisbach equation: dominates; at high /D and Re (fully rough limit), the 2/D term
(1)
where p = pressure drop, Pa f = friction factor, dimensionless (from Moody chart, Figure 13 in Chapter 3) L = length of pipe, m D = internal diameter of pipe, m = fluid density at mean temperature, kg/m3 V = average velocity, m/s
This equation is often presented in specific energy form as L V 2 p h = ------- = f ---- ------ g D 2g
(2)
where h = energy loss, m
g = acceleration of gravity, m/s2
In this form, the fluid’s density does not appear explicitly (although it is in the Reynolds number that influences f ). The friction factor f is a function of pipe roughness , inside diameter D, and parameter Re, the Reynolds number: Re = DV/
(3)
where Re = Reynolds number, dimensionless = absolute roughness of pipe wall, m = dynamic viscosity of fluid, Pa·s
The friction factor is frequently presented on a Moody chart (Figure 13 in Chapter 3) giving f as a function of Re with /D as a parameter. A useful fit of smooth and rough pipe data for the usual turbulent flow regime is the Colebrook equation: 2 1 18.7 --------- = 1.74 – 2 log ----- + ------------------ D f Re f
(4)
Another form of Equation (4) appears in Chapter 21, but the two are equivalent. Equation (4) is useful in showing behavior at limiting cases: as /D approaches 0 (smooth limit), the 18.7/Re f term dominates.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
22.6
2017 ASHRAE Handbook—Fundamentals (SI) Table 3
K Factors: Threaded Steel Pipe Fittings
Nominal Pipe Dia., mm
90° Ell Reg.
90° Ell Long
45° Ell
Return Bend
TeeLine
TeeBranch
Globe Valve
Gate Valve
Angle Valve
Swing Check Valve
Bell Mouth Inlet
10 15 20 25 32 40 50 65 80 100
2.5 2.1 1.7 1.5 1.3 1.2 1.0 0.85 0.80 0.70
— — 0.92 0.78 0.65 0.54 0.42 0.35 0.31 0.24
0.38 0.37 0.35 0.34 0.33 0.32 0.31 0.30 0.29 0.28
2.5 2.1 1.7 1.5 1.3 1.2 1.0 0.85 0.80 0.70
0.90 0.90 0.90 0.90 0.90 0.90 0.90 0.90 0.90 0.90
2.7 2.4 2.1 1.8 1.7 1.6 1.4 1.3 1.2 1.1
20 14 10 9 8.5 8 7 6.5 6 5.7
0.40 0.33 0.28 0.24 0.22 0.19 0.17 0.16 0.14 0.12
— — 6.1 4.6 3.6 2.9 2.1 1.6 1.3 1.0
8.0 5.5 3.7 3.0 2.7 2.5 2.3 2.2 2.1 2.0
0.05 0.05 0.05 0.05 0.05 0.05 0.05 0.05 0.05 0.05
Square Projected Inlet Inlet 0.5 0.5 0.5 0.5 0.5 0.5 0.5 0.5 0.5 0.5
1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0 1.0
Source: Engineering Data Book (Hydraulic Institute 1990).
Licensed for single user. © 2017 ASHRAE, Inc.
Table 4 K Factors: Flanged Welded Steel Pipe Fittings Nominal Pipe Dia., mm
90° Ell Reg.
90° Ell Long
45° Ell Long
25 32 40 50 65 80 100 150 200 250 300
0.43 0.41 0.40 0.38 0.35 0.34 0.31 0.29 0.27 0.25 0.24
0.41 0.37 0.35 0.30 0.28 0.25 0.22 0.18 0.16 0.14 0.13
0.22 0.22 0.21 0.20 0.19 0.18 0.18 0.17 0.17 0.16 0.16
Return Return Bend Bend LongStandard Radius 0.43 0.41 0.40 0.38 0.35 0.34 0.31 0.29 0.27 0.25 0.24
0.43 0.38 0.35 0.30 0.27 0.25 0.22 0.18 0.15 0.14 0.13
TeeLine
TeeBranch
Glove Valve
Gate Valve
Angle Valve
Swing Check Valve
0.26 0.25 0.23 0.20 0.18 0.17 0.15 0.12 0.10 0.09 0.08
1.0 0.95 0.90 0.84 0.79 0.76 0.70 0.62 0.58 0.53 0.50
13 12 10 9 8 7 6.5 6 5.7 5.7 5.7
— — — 0.34 0.27 0.22 0.16 0.10 0.08 0.06 0.05
4.8 3.7 3.0 2.5 2.3 2.2 2.1 2.1 2.1 2.1 2.1
2.0 2.0 2.0 2.0 2.0 2.0 2.0 2.0 2.0 2.0 2.0
Source: Engineering Data Book (Hydraulic Institute 1990).
Equation (4) is implicit in f; that is, f appears on both sides, so a value for f is usually obtained iteratively.
Solution: Use Equation (7). From Table 3, the K for a 25 mm, 90° threaded elbow is 1.5.
Hazen-Williams Equation A less widely used alternative to the Darcy-Weisbach formulation for calculating pressure drop is the Hazen-Williams equation, which is expressed as V p = 6.819L ---- C
or
Example 1. Determine the pressure drop for 15°C water flowing at 1 m/s through a nominal 25 mm, 90° threaded elbow.
1.852
V h = 6.819L ---- C
1 ---- D
1.852
1.167
1 ---- D
(g)
(5)
p = 1.5 1000 12/2 = 750 Pa
The loss coefficient for valves appears in another form as Av , a dimensional coefficient expressing the flow through a valve at a specified pressure drop. Q = A v p
(8)
where 1.167
(6)
where C = roughness factor. Typical values of C are 150 for plastic pipe and copper tubing, 140 for new steel pipe, down to 100 and below for badly corroded or very rough pipe.
Valve and Fitting Losses Valves and fittings cause pressure losses greater than those caused by the pipe alone. One formulation expresses losses as V 2 V 2 p = K ------ or h = K ------ 2 2g
(7)
where K = geometry- and size-dependent loss coefficient (Tables 3 to 6)and = density of fluid 1000 kg/m3 for water at temperatures below 120°C.
Q = volumetric flow, m3/s Av = valve coefficient, m3/s at p = 1 Pa p = pressure drop, Pa
See the section on Control Valve Sizing in Chapter 47 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment for more information on valve coefficients. Example 2. Determine the volumetric flow through a valve with Av = 0.00024 for an allowable pressure drop of 35 kPa. Solution: Use Equation (8). Q = 0.00024 35 000 1000 = 0.0014 m3/s = 1.4 L/s
Alternative formulations express fitting losses in terms of equivalent lengths of straight pipe (see Tables 8 and 27). Pressure loss data for fittings are also presented in Idelchik (1986). Equation (7) and data in Tables 3 and 4 are based on the assumption that separated flow in the fitting causes the K factors to be independent of Reynolds number. In reality, the K factor for most pipe fittings
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Pipe Design
22.7 Table 5 Approximate Range of Variation for K Factors of Steel Fittings
90° Elbow
Regular threaded
±20% above 50 mm
Tee
Threaded, line or branch
±40% below 50 mm Long-radius threaded
45° Elbow Return bend (180°)
±25%
Regular flanged
±35%
Long-radius flanged
±30%
Regular threaded
±10%
Long-radius flanged
±10%
Regular threaded
±25%
Regular flanged
±35%
Long-radius flanged
±30%
±25%
Flanged, line or branch
±35%
Globe valve
Threaded
±25%
Flanged
±25%
Gate valve
Threaded
±25%
Flanged
±50%
Angle valve
Threaded
±20%
Flanged
±50%
Check valve
Threaded
±50%
Flanged
+200% –80%
Source: Engineering Data Book (Hydraulic Institute 1990).
Table 6 Summary of K Values for Steel Ells, Reducers, and Expansions ASHRAE Researchb,c Past a S.R.e
Licensed for single user. © 2017 ASHRAE, Inc.
ell (R/D = 1) thread 50 mm 100 mm S.R. ell (R/D = 1) weld 25 mm L.R. ell (R/D = 1.5) weld 50 mm L.R. ell (R/D = 1.5) weld 100 mm L.R. ell (R/D = 1.5) weld 150 mm L.R. ell (R/D = 1.5) weld 200 mm L.R. ell (R/D = 1.5) weld 250 mm L.R. ell (R/D = 1.5) weld 300 mm L.R. ell (R/D = 1.5) weld 400 mm L.R. ell (R/D = 1.5) weld 500 mm L.R. ell (R/D = 1.5) weld 600 mm L.R. ell (R/D = 1.5) weld Reducer (50 by 40 mm) thread (100 by 80 mm) weld (150 by 100 mm) weld (200 by 150 mm) weld (250 by 200 mm) weld (300 by 250 mm) weld (400 by 300 mm) weld (500 by 400 mm) weld (600 by 500 mm) weld Expansion (40 by 50 mm) thread (80 by 100 mm) weld (100 by 150 mm) weld (150 by 200 mm) weld (200 by 250 mm) weld (250 by 300 mm) weld (300 by 400 mm) weld (400 by 500 mm) weld (500 by 600 mm) weld
1.2 m/s
(1.0)d
0.60 to 1.0 0.30 to 0.34
to 1.0 0.50 to 0.7 0.22 to 0.33 (0.22)d 0.25 0.20 to 0.26 0.17 0.16 0.12 0.09 0.07 — 0.22
Source: Rahmeyer (2003a). aPublished data by Crane Co. (1988), Freeman (1941), and Hydraulic Institute (1990). bRahmeyer (1999a, 2002a).
varies with Reynolds number. Tests by Rahmeyer (1999a, 1999b, 2002a, 2002b) (ASHRAE research projects RP-968 and RP-1034) on 50 mm threaded and 100, 300, 400, 500, and 600 mm welded steel fittings demonstrate the variation and are shown in Tables 6 and 7. The studies also present K factors of diverting and mixing flows in tees, ranging from full through flow to full branch flow. They also examined the variation in K factors caused by variations in geometry among manufacturers and by surface defects in individual fittings. Hegberg (1995) and Rahmeyer (1999a, 1999b) discuss the origins of some of the data shown in Tables 6 and 7. The Hydraulic Institute (1990) data appear to have come from Freeman (1941),
— — — — — — — — — — — — —
2.4 m/s
3.6 m/s
0.60 0.37
0.68 0.34
0.736 0.33
— — 0.26 0.26 0.22 0.21 0.17 0.12 0.12 0.098 0.53 0.23 0.62 0.31 0.16 0.14 0.17 0.16 0.053 0.16 0.11 0.28 0.15 0.11 0.11 0.073 0.024 0.020
— — 0.24 0.24 0.20 0.17 0.17 0.12 0.10 0.089 0.28 0.14 0.54 0.28 0.14 0.14 0.16 0.13 0.053 0.13 0.11 0.28 0.12 0.09 0.11 0.076 0.021 0.023
— — 0.23 0.24 0.19 0.16 0.17 0.11 0.10 0.089 0.20 0.10 0.53 0.26 0.14 0.14 0.17 0.13 0.055 0.02 0.11 0.29 0.11 0.08 0.11 0.073 0.022 0.020
cDing
et al. (2005) ) Data published in 1993 ASHRAE Handbook—Fundamentals. eS.R.—short radius or regular ell; L.R.—long-radius ell.
d(
work that was actually performed in 1895. The work of Giesecke (1926) and Giesecke and Badgett (1931, 1932a, 1932b) may not be representative of present-day fittings. Further extending the work on determination of fitting K factors to PVC piping systems, Rahmeyer (2003a, 2003b) (ASHRAE research project RP-1193) found the data in Tables 8 and 9 giving K factors for Schedule 80 PVC 50, 100, 150, and 200 mm ells, reducers, expansions, and tees. The results of these tests are also presented in the cited papers in terms of equivalent lengths. In general, PVC fitting geometry varied much more from one manufacturer to another than steel fittings did.
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22.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 7
Summary of Test Data for Loss Coefficients K for Steel Pipe Tees
Licensed for single user. © 2017 ASHRAE, Inc.
ASHRAE Researchb,c Pasta
1.2 m/s
2.4 m/s
3.6 m/s
50 mm thread tee, 100% branch 100% line (flow-through) 100% mix
1.20 to 1.80 (1.4)d 0.50 to 0.90 (0.90)d —
0.93 0.19 1.19
— — —
— —
100 mm weld tee, 100% branch 100% line (flow-through) 100% mix
0.70 to 1.02 (0.70)d 0.15 to 0.34 (0.15)d —
— — —
0.57 0.06 0.49
— — —
150 mm weld tee, 100% branch 100% line (flow-through) 100% mix
— — —
— — —
0.56 0.12 0.88
— — —
200 mm weld tee, 100% branch 100% line (flow-through) 100% mix
— — —
— — —
0.53 0.08 0.70
— — —
250 mm weld tee, 100% branch 100% line (flow-through) 100% mix
— — —
— — —
0.52 0.06 0.77
— — —
300 mm weld tee, 100% branch 100% line (flow-through) 100% mix
0.52 0.09 —
0.70 0.062 0.88
0.63 0.091 0.72
0.62 0.096 0.72
400 mm weld tee, 100% branch 100% line (flow-through) 100% mix
0.47 0.07 —
0.54 0.032 0.74
0.55 0.028 0.74
0.54 0.028 0.76
aPublished
bRahmeyer
data by Crane Co. (1988), Freeman (1941), and Hydraulic Institute (1990). (1999b, 2002b).
cDing dData
et al. (2005). published in 1993 ASHRAE Handbook—Fundamentals.
Table 8 Test Summary for Loss Coefficients K and Equivalent Loss Lengths Schedule 80 PVC Fitting Injected molded elbow,
50 mm 100 mm 150 mm 200 mm
K
L, m
0.91 to 1.00 0.86 to 0.91 0.76 to 0.91 0.68 to 0.87
2.6 to 2.8 5.6 to 5.9 8.0 to 9.5 10.0 to 12.8
200 mm fabricated elbow, Type I, components Type II, mitered
0.40 to 0.42
5.9 to 6.2
0.073 to 0.76
10.8 to 11.2
150 by 100 mm injected molded reducer Bushing type
0.12 to 0.59 0.49 to 0.59
1.2 to 6.2 5.2 to 6.2
200 by 150 mm injected molded reducer Bushing type Gradual reducer type
0.13 to 0.63 0.48 to 0.68 0.21
1.9 to 9.3 7.1 to 10.0 3.1
100 by 150 mm injected molded expansion 0.069 to 1.19 Bushing type 0.069 to 1.14
0.46 to 7.7 0.46 to 7.4
150 by 200 mm injected molded expansion 0.95 to 0.96 Bushing type 0.94 to 0.95 Gradual reducer type 0.99
10.0 to 10.1 9.9 to 10.0 10.4
Losses in Multiple Fittings Typical fitting loss calculations are done as if each fitting is isolated and has no interaction with any other. Rahmeyer (2002c) (ASHRAE research project RP-1035) tested 50 mm threaded ells and 100 mm ells in two and three fitting assemblies of several geometries, at varying spacing. Figure 1 shows the geometries, and Figures 2 and 3 show the ratio of coupled K values to uncoupled K values (i.e., fitting losses for the assembly compared with the sum of losses from the same number of isolated fittings). The most important conclusion is that the interaction between fittings always reduces the loss. Also, although geometry of the assembly has a definite effect, the effects are not the same for 50 mm threaded and 100 mm welded ells. Thus, the traditional
Fig. 1
Close-Coupled Test Configurations
practice of adding together losses from individual fittings gives a conservative (high-limit) estimate.
Calculating Pressure Losses The most common engineering design flow loss calculation selects a pipe size for the desired total flow rate and available or allowable pressure drop. Because either formulation of fitting losses requires a known diameter, pipe size must be selected before calculating the detailed influence of fittings. A frequently used rule of thumb assumes that the design length of pipe is 50 to 100% longer than actual to account for fitting losses. After a pipe diameter has been selected on this basis, the influence of each fitting can be evaluated.
Stress Calculations Metallic Pipe. Although stress calculations are seldom required, the factors involved should be understood. The main areas of concern are (1) internal pressure stress, (2) longitudinal stress caused by pressure and mass, and (3) stress from expansion and contraction. ASME Standard B31 standards establish a basic allowable stress S equal to one-fourth of the minimum tensile strength of the material. This value is adjusted, as discussed in this section, because of the nature of certain stresses and manufacturing processes.
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Pipe Design
22.9 Table 9 Test Summary for Loss Coefficients K of PVC Tees Branching Schedule 80 PVC Fitting
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Fig. 2 Summary Plot of Effect of Close-Coupled Configurations for 50 mm Ells
K1-3
K1-2
50 mm injection molded branching tee, 100% 0.13 to 0.26 line flow 50/50 flow 0 to 0.12 100% branch flow — 100 mm injection molded branching tee, 100% 0.07 to 0.22 line flow 50/50 flow 0.03 to 0.13 100% branch flow — 150 mm injection molded branching tee, 100% 0.01 to 0.14 line flow 50/50 flow 0.06 to 0.11 100% branch flow — 150 mm fabricated branching tee, 100% line flow 0.21 to 0.22 50/50 flow 0.04 to 0.09 100% branch flow — 200 mm injection molded branching tee, 100% 0.04 to 0.09 line flow 50/50 flow 0.04 to 0.07 100% branch flow — 200 mm fabricated branching tee, 100% line flow 0.09 to 0.16 50/50 flow 0.08 to 0.13 100% branch flow —
— 0.74 to 1.02 0.98 to 1.39 — 0.74 to 0.82 0.97 to 1.12 — 0.70 to 0.84 0.95 to 1.15 — 1.29 to 1.40 1.74 to 1.88 — 0.64 to 0.75 0.85 to 0.96 — 1.07 to 1.16 1.40 to 1.62
Mixing PVC Fitting
Fig. 3 Summary Plot of Effect of Close-Coupled Configurations for 100 mm Ells Hoop stress caused by internal pressure is the major stress on pipes. Because some forming methods form a seam that may be weaker than the base material, ASME Standard B31.9 specifies a joint efficiency factor E, multiplied by the basic allowable stress to establish a maximum allowable stress value in tension SE. (Table A-1 in ASME Standard B31.9 lists values of SE for commonly used pipe materials.) The joint efficiency factor can be significant; for example, seamless pipe has a joint efficiency factor of 1, so it can be used to the full allowable stress (one-quarter of the tensile strength). In contrast, butt-welded pipe has a joint efficiency factor of 0.60, so its maximum allowable stress must be derated (SE = 0.6S). Equation (9) determines the minimum wall thickness for a given pressure. Equation (10) determines the maximum pressure allowed for a given wall thickness. pD t m = --------- + A 2S E
(9)
2S E t m – A p = ----------------------------D
(10)
50 mm injection molded mixing tee, 100% line flow 50/50 flow 100% mix flow 100 mm injection molded mixing tee, 100% line flow 50/50 flow 100% mix flow 150 mm injection molded mixing tee, 100% line flow 50/50 flow 100% mix flow 150 mm fabricated mixing tee, 100% line flow 50/50 flow 100% mix flow 200 mm injection molded mixing tee, 100% line flow 50/50 flow 100% mix flow 200 mm fabricated mixing tee, 100% line flow 50/50 flow 100% mix flow
K1-2
K3-2
0.12 to 0.25
—
1.22 to 1.19 0.89 to 1.88 — 0.89 to 1.54 0.07 to 0.18 — 1.19 to 1.88 0.98 to 1.88 — 0.88 to 1.02 0.06 to 0.14 — 1.26 to 1.80 — 0.19 to 0.21 2.94 to 3.32 — 0.04 to 0.09
1.02 to 1.60 0.90 to 1.07 — 2.57 to 3.17 1.72 to 1.98 —
1.10 to 1.60 0.96 to 1.32 — 0.81 to 0.93 0.13 to 0.70 — 2.36 to 10.62 2.02 to 2.67 — 1.34 to 1.53
Coefficients based on average velocity of 2.43 m/s. Range of values varies with fitting manufacturers. Line or straight flow is Q2/Q1 = 100%. Branch flow is Q2/Q1 = 0%.
where SA = allowable stress range, kPa Sc = allowable cold stress at coolest temperature system will experience, kPa Sh = allowable hot stress at hottest temperature system will experience, kPa
Both equations incorporate an allowance factor A to compensate for manufacturing tolerances, material removed in threading or grooving, and corrosion. For the seamless, butt-welded, and electric resistance welded (ERW) pipe most commonly used in HVAC work, the standards apply a manufacturing tolerance of 12.5%. Working pressure for steel pipe (see Table 16) has been calculated using a manufacturing tolerance of 12.5%, standard allowance for depth of thread (where applicable), and a corrosion allowance of 1.65 mm for pipes 65 mm and larger and 0.64 mm for pipes 50 mm
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22.10
2017 ASHRAE Handbook—Fundamentals (SI)
and smaller. Where corrosion is known to be greater or smaller, pressure rating can be recalculated using Equation (10). Higher pressure ratings than shown in Table 16 can be obtained (1) by using ERW or seamless pipe in lieu of continuous-weld (CW) pipe 100 mm and less, and seamless pipe in lieu of ERW pipe 125 mm and greater (because of higher joint efficiency factors); or (2) by using heavier-wall pipe. Longitudinal stresses caused by pressure, mass, and other sustained forces are additive, and the sum of all such stresses must not exceed the basic allowable stress S at the highest temperature at which the system will operate. Longitudinal stress caused by pressure equals approximately one-half the hoop stress caused by internal pressure; thus, at least one-half the basic allowable stress is available for mass and other sustained forces. This factor is taken into account in Table 11. Stresses caused by expansion and contraction are cyclical, and, because creep allows some stress relaxation, the ASME Standard B31 series allows designing to an allowable stress range SA as calculated by Equation (11). Table 15 lists allowable stress ranges for commonly used piping materials. SA = 1.25Sc + 0.25Sh
(11)
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where SA = allowable stress range, kPa Sc = allowable cold stress at coolest temperature system will experience, kPa Sh = allowable hot stress at hottest temperature system will experience, kPa
Nonmetallic. Both thermoplastics and thermosets have an allowable stress derived from a hydrostatic design basis stress (HDBS). The HDBS is determined by a statistical analysis of both static and cyclic stress rupture test data as set forth in ASTM Standard D2837 for thermoplastics and ASTM Standard D2992 for glass-fiber-reinforced thermosetting resins. The allowable stress, called the hydrostatic design stress (HDS), is obtained by multiplying the HDBS by a service factor. HDS values recommended by some manufacturers and those allowed by ASME Standard B31 are listed in Table 18. The pressure design thickness for plastic pipe can be calculated using the code stress values and the formula in Equation (12): t = pD/(2S + p) = = = =
SIZING PROCEDURE
A procedure for sizing piping systems is as follows: 1. Determine system type (open, closed, compressible, incompressible, pumped, gravity feed, domestic, etc.). 2. Determine type and properties of fluid to be conveyed in the pipe. 3. Determine temperatures used (high, low) and temperature differentials. 4. Identify system pressures encountered in the system (working, maximum, low, fill, and relief pressures). 5. Determine load at each device (e.g., heating or cooling requirements, fixture units for plumbing) to find flow. 6. Sketch main, risers, and branches, and indicate equipment to be served and each device’s flow rate. 7. Determine flow of supply pipe for each pipe segment run by summing the loads at the furthest device and running back to the source. 8. Determine flow of each return pipe by starting at the first device returning water and summing the loads back to the source (when applicable). 9. Determine equivalent length of pipe in the main lines, risers, branches, and returns. Because pipe sizes are not known, the exact equivalent length of various fittings cannot be determined. Add the equivalent lengths, starting at the beginning and proceeding along the mains, risers, branches, and returns (when applicable). 10. In domestic or gravity feed: calculate the approximate design value of the average pressure drop per metre length of pipe in equivalent length determined in step 9. In pumped system: calculate pressure drop H using the flow rate and pressure drop for pipe from Equations (2) or (6), the valves and fittings using head drop from Equation (7), and head from the devices from the manufacturer’s data. p = (ps – 9.8H – pf – pm)/L
(14)
where
pressure design thickness, mm internal design pressure, kPa (gage) pipe outside diameter, mm hydrostatic design stress (HDS), kPa
The minimum required wall thickness can be found by adding an allowance for mechanical strength, threading, grooving, erosion, and corrosion to the calculated pressure design thickness. Another method of rating pressure rating of piping used by manufacturers is the standard dimension ratio (SDR), which is the ratio of the pipe diameter to the wall thickness: SDR= D/s
1.5
(12)
where t p D S
There are many formulations of the polymers used for piping materials, and different joining methods for each, so manufacturers’ recommendations should be followed. Most catalogs give pressure ratings for pipe and fittings at various temperatures up to the maximum the material will withstand.
(13)
where D = pipe outside diameter, mm s = pipe wall thickness, mm
An SDR of 11 means that the outside diameter D of the pipe is 11 times the thickness of the wall s. A high SDR means that the pipe’s wall is thin compared to its diameter, and a low SDR means that the pipe’s wall is thick relative to pipe diameter. SDR is inversely correlated with pressure rating: high SDR indicates a low pressure rating, whereas low-SDR pipes have higher pressure ratings.
p = average pressure loss per metre of equivalent length of pipe, kPa ps = pressure at the source, kPa pf = minimum pressure required to operate device, kPa pm = pressure drop through any meters, kPa H = height of highest fixture above source (if open system), m L = equivalent length determined in step 4, m
11. In domestic or gravity system: from the expected rate of flow (step 5) and p (step 10), select pipe sizes. In pumped system: select the pump using the flow rate and calculated H.
1.6
PIPE-SUPPORTING ELEMENTS
Pipe-supporting elements consist of (1) hangers, which support from above; (2) supports, which bear load from below; and (3) restraints, such as anchors and guides, that limit or direct movement as well as support loads. Pipe-supporting elements must withstand all static and dynamic conditions, including the following: • Mass of pipe, valves, fittings, insulation, and fluid contents, including test fluid if using heavier-than-normal media • Occasional loads such as ice, wind, and seismic forces or testing loads (e.g., hydrostatic loads on a steam pipe)
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Pipe Design
22.11
Table 10 Capacities of ASTM A36 Steel Threaded Rods Rod Diameter, mm 6.4 10 13 16 19 22 25 32
Root Area of Coarse Thread, mm2
Maximum Load,* N
17.4 43.9 81.3 130.3 194.8 270.3 356.1 573.5
1 070 2 720 5 030 8 060 12 100 16 800 22 100 35 600
*Based on allowable stress of 83 MPa reduced by 25% using root area in accordance with ASME Standard B31.1 and MSS Standard SP-58.
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• Forces imposed by thermal expansion and contraction of pipe bends and loops • Frictional, spring, and pressure thrust forces imposed by expansion joints in the system • Frictional forces of guides and supports • Other loads (e.g., water hammer, vibration, reactive force of relief valves) • Test load and force In addition, pipe-supporting elements must be evaluated in terms of stress at the points of connection to the pipe and the building. Stress at the point of connection to the pipe is especially important for base elbow and trunnion supports, because this stress is usually the limiting parameter, not the strength of the structural member. Loads on anchors, cast-in-place inserts, and other attachments to concrete should not be more than one-fifth the ultimate strength of the attachment, as determined by manufacturers’ tests. All loads on the structure should be communicated to and coordinated with the structural engineer. The ASME B31 standards establish criteria for the design of pipe-supporting elements, and the Manufacturers Standardization Society of the Valve and Fittings Industry (MSS) has established standards for the design, fabrication, selection, and installation of pipe hangers and supports based on these codes. MSS Standard SP-69 and the catalogs of many manufacturers illustrate the various hangers and components and provide information on the types to use with different pipe systems. Table 10 lists maximum safe loads for threaded steel rods, and Tables 11 and 12 show suggested pipe support spacing for metal and PVC pipes, respectively. Loads on most pipe-supporting elements are moderate and can be selected safely in accordance with manufacturers’ catalog data and the information presented in this section; however, some loads and forces can be very high, especially in multistory buildings and for large-diameter pipe, particularly where expansion joints are used at a high operating pressure. Consequently, a qualified engineer should design or review all anchors and pipe-supporting elements, especially for the following: • • • •
Steam systems operating above 100 kPa (gage) Hydronic systems operating above 1.1 MPa or 120°C Risers over 10 stories or 30 m Systems with expansion joints, especially for pipe diameters 80 mm and greater • Pipe sizes over 300 mm diameter • Anchor loads greater than 44 kN • Moments on pipe or structure in excess of 1.4 kN·m
Table 11 Suggested Hanger Spacing and Rod Size for Straight Horizontal Runs Nominal O.D., mm 15 20 25 40 50 65 80 100 150 200 250 300 350 400 450 500
Hanger Spacing, m Standard Steel Pipe*
Copper Tube
Water
Steam
Water
2.1 2.1 2.1 2.7 3.0 3.4 3.7 4.3 5.2 5.8 6.1 7.0 7.6 8.2 8.5 9.1
2.4 2.7 2.7 3.7 4.0 4.3 4.6 5.2 6.4 7.3 7.9 9.1 9.8 10.7 11.3 11.9
1.5 1.5 1.8 2.4 2.4 2.7 3.0 3.7 4.3 4.9 5.5 5.8
Rod Size, mm 6.4 6.4 6.4 10 10 10 10 13 13 16 19 22 25 25 32 32
Source: Adapted from MSS Standard SP-69 *Spacing does not apply where span calculations are made or where concentrated loads are placed between supports such as flanges, valves, specialties, etc.
are hoop stress caused by internal pressure, and longitudinal stresses caused by pressure, mass, and other sustained loads. Detailed stress calculations are seldom required for HVAC applications because standard pipe has ample thickness to sustain the pressure and longitudinal stress caused by mass (assuming hangers are spaced in accordance with Table 11). Support spacings for PVC and CPVC pipe systems are influenced by operating temperatures. Table 12 recommends horizontal spacing based on pipe size, schedule, material (PVC or industrialgrade CPVC), and operating temperature. Hangers and supports should not be clamped tightly because the axial movement of the pipe would be restricted. The charts are based on continuous spans and uninsulated lines carrying liquids. They are not applicable where loads between supports are concentrated (e.g., for valves, flanges) or where there is a change in direction. Hangers/supports should be located adjacent to joints, branch connections, and changes in direction. Risers should be in installed independently of adjacent horizontal hangers/supports. For cast iron pipe, maximum spacing should be 3.7 m, with at least one hanger/support for each pipe section.
1.7
PIPE EXPANSION AND FLEXIBILITY
Temperature changes cause dimensional changes in all materials. Table 13 shows the coefficients of expansion for metallic piping materials commonly used in HVAC. For systems operating at high temperatures, such as steam and hot water, the rate of expansion is high, and significant movements can occur in short runs of piping. Even though rates of expansion may be low for systems operating in the range of 5 to 40°C, such as chilled and condenser water, they can cause large movements in long runs of piping, which are common in distribution systems and high-rise buildings. Therefore, in addition to design requirements for pressure, mass, and other loads, piping systems must accommodate thermal and other movements to prevent the following:
Hanger Spacing and Pipe Wall Thickness
• Failure of pipe and supports from overstress and fatigue • Leakage of joints • Detrimental forces and stresses in connected equipment
Table 11 suggests minimum pipe hanger spacing for use unless exceeded by the local authority having jurisdiction or engineering calculations. The primary factors determining pipe wall thickness
An unrestrained pipe operates at the lowest overall stress level. Anchors and restraints are needed to support pipe mass and to protect equipment connections. Anchor forces and bowing of pipe
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22.12
2017 ASHRAE Handbook—Fundamentals (SI)
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Table 12 Suggested Maximum Spacing Between Hangers/Support for PVC and CPVC Pipe
anchored at both ends are generally too large to be acceptable, so general practice is to never anchor a straight run of steel pipe at both ends. Piping must be allowed to expand or contract through thermal changes. Ample flexibility can be attained by designing pipe bends and loops or by including supplemental devices, such as expansion joints. End reactions transmitted to rotating equipment, such as pumps or turbines, may deform the equipment case and cause bearing misalignment that may ultimately cause the component to fail. Consequently, manufacturers’ recommendations on allowable forces and movements that may be placed on their equipment should be followed.
1.8
Fig. 4 Guided Cantilever Beam
PIPE BENDS AND LOOPS
Detailed stress analysis requires involved mathematical analysis and is generally performed by computer programs. However, such involved analysis is not typically required for most HVAC systems because the piping arrangements and temperature ranges at which they operate are usually simple to analyze. Expansion stresses discussed in this section relate only to aboveground pipe located in open air, or preinsulated pipe.
L Bends The guided cantilever beam method of evaluating L bends can be used to design L bends, Z bends, pipe loops, branch take-off connections, and some more complicated piping configurations. The guided cantilever equation [see Equation (17)] is generally conservative because it assumes that the pipe arrangement does not rotate. The anchor force results will be higher because of the lack of rotation, and rigorous analysis is recommended for complicated or expensive systems. Equation (15) may be used to calculate the length of leg BC needed to accommodate thermal expansion or contraction of leg AB for a guided cantilever beam (Figure 4).
L=
3 DE --------------SA
(15)
where L = length of leg BC required to accommodate thermal expansion of long leg AB, mm = thermal expansion or contraction of leg AB, mm D = actual pipe outside diameter, mm E = modulus of elasticity, kPa SA = allowable stress range, kPa
For the commonly used A53 Grade B seamless or ERW pipe, an allowable stress SA of 155 MPa (see Table 15) can be used without overstressing the pipe. However, this can result in very high end reactions and anchor forces, especially with large-diameter pipe. Designing to a stress range SA of 103 MPa and assuming E = 193 GPa, Equation (15) reduces to Equation (16), which provides reasonably low end reactions without requiring too much extra pipe. In addition, Equation (16) may be used with A53 continuous (butt-) welded, seamless, and ERW pipe, and B88 drawn copper tubing. L = 75 D
(16)
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Pipe Design Table 13
22.13 Thermal Expansion of Metal Pipe
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Vacuum
Linear Thermal Expansion, mm/m Saturated Steam Pressure, Temperature, Carbon Type 304 kPa (gage) °C Steel Stainless Steel Copper –34 –29 –23 –18 –12 –7
–0.16 –0.10 –0.05 0.00 0.07 0.13
–0.25 –0.17 –0.08 0.00 0.09 0.18
–0.27 –0.18 –0.09 0.00 0.10 0.20
–100.7 –100.7 –100.0 –99.3 –98.6 –97.9 –96.5 –94.5 –89.6 –81.4 –69.0 –49.6 –22.1
0 4 10 16 21 27 32 38 49 60 71 82 93
0.20 0.25 0.32 0.38 0.44 0.51 0.57 0.63 0.76 0.88 1.02 1.14 1.27
0.30 0.38 0.47 0.56 0.65 0.75 0.84 0.93 1.13 1.31 1.49 1.68 1.87
0.31 0.38 0.47 0.57 0.66 0.75 0.85 0.94 1.14 1.33 1.50 1.71 1.92
0 17.2 71.0 142.7 238.6 360.6 517.1 712.3 953.6 1249
100 104 116 127 138 149 160 171 182 193
1.35 1.41 1.54 1.68 1.82 1.96 2.11 2.25 2.40 2.54
1.98 2.07 2.26 2.45 2.64 2.83 3.03 3.23 3.43 3.63
2.03 2.10 2.30 2.49 2.68 2.88 3.08 3.28 3.48 3.68
1604 9039
204 304 404 504
2.69 4.11 5.67 7.31
3.83 5.65 7.56 9.54
4.06 5.77 7.72 9.76
The guided cantilever method of designing L bends assumes no restraints; therefore, care must be taken in supporting the pipe. For horizontal L bends, it is usually necessary to place a support near point B (see Figure 4), and any supports between points A and C must provide minimal resistance to piping movement; this is done by using slide plates or hanger rods of ample length, with hanger components selected to allow for swing no greater than 4°. For L bends containing both vertical and horizontal legs, any supports on the horizontal leg must be spring hangers designed to support the full mass of pipe at normal operating temperature with a maximum load variation of 25%. The force developed in an L bend that must be sustained by anchors or connected equipment is determined by the following equation: 12E c I F = ----------------6 10 L 3 where F Ec I L
= = = = =
force, kN modulus of elasticity, kPa moment of inertia, mm4 length of offset leg, mm deflection of offset leg, mm
(17)
Fig. 5 Z Bend in Pipe
Z Bends Z bends, as shown in Figure 5, are very effective for accommodating pipe movements. A simple and conservative method of sizing Z bends is to design the offset leg to be 65% of the values used for an L bend in Equation (15): L = 48.7 D
(18)
where L = length of offset leg, mm = anchor-to-anchor expansion, mm D = pipe outside diameter, mm
The force developed in a Z bend can be calculated with acceptable accuracy as follows: F = C1(D/L)2
(19)
where C1 F D L
= = = = =
101 kN/mm force, kN pipe outside diameter, mm length of offset leg, mm anchor-to-anchor expansion, mm
U Bends and Pipe Loops Pipe loops or U bends are commonly used in long runs of piping. A simple method of designing pipe loops is to calculate the anchorto-anchor expansion and, using Equation (15), determine the length L necessary to accommodate this movement. The pipe loop dimensions can then be determined using W = L/5 and H = 2W. Note that guides must be spaced no closer than twice the height of the loop, and piping between guides must be supported, as described in the section on L Bends, when the length of pipe between guides exceeds the maximum allowable hanger spacing for the size pipe. Table 14 lists pipe loop dimensions for pipe sizes 25 to 600 mm and anchor-to-anchor expansion (contraction) of 50 to 300 mm. No simple method has been developed to calculate pipe loop force; however, it is generally low. A conservative estimate is 35 N per millimetre diameter (e.g., a 50 mm pipe will develop 1.75 kN of force and a 300 mm pipe will develop 10.5 kN of force). Additional analysis should be done for pipes greater than 300 mm in diameter, because other simplified methodologies predict higher anchor forces.
Expansion and Contraction Control of Other Materials To design expansion and contraction loops and bends for other materials, consult the Copper Development Association (CDA 2010) for copper pipes, and Plastic Pipe and Fitting Association (PPFA 2009) for plastic pipes.
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22.14
2017 ASHRAE Handbook—Fundamentals (SI)
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Table 14
Pipe Loop Design for A53 Grade B Carbon Steel Pipe Through 200°C
Pipe Nom., O.D., mm
W
H
W
H
W
H
W
H
W
H
W
H
25 50 80 100 150 200 250 300 350 400 450 500 600
0.6 0.9 1.1 1.2 1.5 1.7 1.8 2.0 2.1 2.3 2.4 2.6 2.7
1.2 1.8 2.1 2.4 3.0 3.4 3.7 4.0 4.3 4.6 4.9 5.2 5.5
0.9 1.2 1.5 1.7 2.0 2.3 2.6 2.7 2.9 3.0 3.4 3.5 3.8
1.8 2.4 3.0 3.4 4.0 4.6 5.2 5.5 5.8 6.1 6.7 7.0 7.6
1.1 1.5 1.8 2.0 2.4 2.7 3.0 3.4 3.5 3.8 4.0 4.3 4.4
2.1 3.0 3.7 4.0 4.9 5.5 6.1 6.7 7.0 7.6 7.9 8.5 8.8
1.2 1.7 2.0 2.3 2.7 3.2 3.5 3.8 4.0 4.3 4.6 4.9 5.3
2.4 3.4 4.0 4.6 5.5 6.4 7.0 7.6 7.9 8.5 9.1 9.8 10.7
1.4 1.8 2.3 2.6 3.0 3.7 4.0 4.3 4.6 4.9 5.2 5.5 5.9
2.7 3.7 4.6 5.2 6.1 7.3 7.9 8.5 9.1 9.8 10.4 11.0 11.9
1.5 2.1 2.4 2.7 3.4 4.0 4.3 4.7 4.9 5.3 5.6 5.9 6.4
3.0 4.3 4.9 5.5 6.7 7.9 8.5 9.4 9.8 10.7 11.3 11.9 12.8
Anchor-to-Anchor Expansion, mm 50
100
Notes: 1. W and H dimensions are metres. 2. L is determined from Equation (15). W = L/5
150
H = 2W
200
2H + W = L
250
300
3. Approximate force to deflect loop = 35 N/mm pipe diameter. For example, 200 mm pipe creates 7600 N of force.
Cold Springing of Pipe Cold springing or cold positioning of pipe consists of offsetting or springing the pipe in a direction opposite the expected movement. Cold springing is not recommended for most HVAC piping. Furthermore, cold springing does not allow designing a pipe bend or loop for twice the calculated movement. For example, if a particular L bend can accommodate 75 mm of movement from a neutral position, cold springing does not allow the L bend to accommodate 150 mm of movement.
Analyzing Existing Piping Configurations Piping is best analyzed using a computer stress analysis program, which can provide all pertinent data, including stress, movements, and loads. Services can perform such analysis if programs are not available in house. However, many situations do not require such detailed analysis. A simple, satisfactory method for single and multiplane systems is to divide the system with real or imaginary anchors into a number of single-plane units, as shown in Figure 6, that can be evaluated as L and Z bends.
2.
PIPE AND FITTING MATERIALS 2.1
PIPE
Steel Pipe Steel pipe is manufactured by several processes. Seamless pipe (Type S), made by piercing or extruding, has no longitudinal seam. Other manufacturing methods roll a strip or sheet of steel (skelp) into a cylinder and weld a longitudinal seam. A continuous-weld (Type F CW) furnace butt-welding (BW; i.e., welding pipe in a single plane) process forces and joins the edges together at high temperature. An electric current welds the seam in electric-resistance-welded (Type E ERW) pipe. ASTM standards such as A53 and A106 specify steel
Fig. 6
Multiplane Pipe System
pipe A and B grades. The A grade has a lower tensile strength and is not widely used. The ASME pressure piping codes require that a longitudinal joint efficiency factor E (Table 15) be applied to each type of seam when calculating the allowable stress. ASME Standard B36.10M specifies the dimensional standard for wrought steel pipe. Steel pipe is manufactured with wall thicknesses identified by schedule or weight class. Although schedule numbers and weight class designations are related, they are not constant for all pipe sizes. United States standard (STD) and Schedule 40 pipe have the same wall thickness through 250 mm NPS. For 300 mm and larger standard weight pipe, the wall thickness remains constant at 10 mm, whereas Schedule 40 wall thickness increases with each size. A similar equality exists between Extra Strong (XS) and Schedule 80 pipe through 200 mm; above 200 mm, XS pipe has a 12.7 mm wall, whereas Schedule 80 increases in wall thickness. Table 16 lists properties of representative steel pipe. Joints in steel pipe are made by welding or by using threaded, flanged, or grooved fittings or socket welding. Unreinforced welded-in branch connections weaken a main pipeline, and added reinforcement is necessary, unless the excess wall thickness of both mains and branches is sufficient to sustain the pressure.
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Pipe Design
22.15 Table 15 Allowable Stressesa for Pipe and Tube
ASTM Specification A53 steel A53 steel A53 steel A106 steel B88 copper
Grade
Type
Manufacturing Process
— B B B —
F S E S —
Cont. weld Seamless ERW Seamless Hard drawn
Available Sizes, mm 15 to 100 15 to 660 50 to 500 15 to 660 8 to 300
Minimum Basic Tensile Allowable Strength, MPa Stress S, MPa 310 413 413 413 248
77.5 103 103 103 62
Joint Efficiency Factor E 0.6 1.0 0.85 1.0 1.0
Allowable Stressb SE, MPa 46.5 103 87.6 103 62
Allowable Stress Rangec SA, MPa 117 155 155 155 93.1
aListed
stresses are for temperatures to 340°C for steel pipe (to 205°C for Type F) and to 120°C for copper tubing. be used for internal pressure stress calculations in Equations (10) and (11). cTo be used only for piping flexibility calculations; see Equations (12) and (13).
bTo
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The ASME Standard B31 series gives formulas and guidelines for determining whether reinforcement is required. Such calculations are seldom needed in HVAC applications because (1) the fitting is designed in accordance with a standard listed in the applicable ASME B31 table and used within the pressure and temperature limits of that standard, and (2) fittings such as tees and reinforced outlet fittings provide integral reinforcement. Type F steel pipe is not allowed for ASME Standard B31.5 refrigerant piping.
Copper Tube Because of their inherent resistance to corrosion and ease of installation, copper and copper alloys are often used in heating, airconditioning, refrigeration, and water supply installations. The two main standards for copper tube are (1) ASTM Standard B88, which includes Types K, L, M, and DWV for water and drain service; and (2) ASTM Standard B280, which specifies air-conditioning and refrigeration (ACR) tube for refrigeration service. Types K, L, M, and DWV designate descending wall thicknesses for copper tube. All types have the same outside diameter (OD) for corresponding sizes. Table 17 lists properties of ASTM B88 copper tube. In the plumbing industry, tube of nominal size approximates the inside diameter. The heating and refrigeration trades specify copper tube by the outside diameter. ACR tubing has a different set of wall thicknesses. Types K, L, and M tube may be hard drawn or annealed (soft) temper. Copper tubing is joined with soldered or brazed, wrought or cast copper capillary socket-end fittings. See Table 20 for lists pressure/ temperature ratings of soldered and brazed joints. Small copper tube is also joined by flare or compression fittings. Hard-drawn tubing has a higher allowable stress than annealed tubing, but if hard tubing is joined by soldering or brazing, the annealed allowable stress should be used. Brass pipe and copper pipe are also made in steel pipe thicknesses for threading. High cost has eliminated these materials from the market, except for special applications. The heating and air-conditioning industry generally uses Types L and M tubing, which have higher internal working pressure ratings than the solder joints used at fittings. Type K may be used with brazed joints for higher pressure-temperature requirements or for direct burial. Type M should be used with care where exposed to potential external damage. Copper and brass should not be used in ammonia refrigerating systems, or in acidic drains from condensing boilers. The section on Special Systems covers other limitations on refrigerant piping.
Ductile Iron and Cast Iron Cast-iron soil pipe comes as Class 4000 series. It is not used under pressure because the pipe is not suitable and the joints are not restrained. Cast-iron pipe and fittings typically have bell and spigot ends for lead and oakum joints or elastomer push-on joints. Castiron pipe and fittings are also furnished with no-hub ends for joining
with no-hub clamps. Local plumbing codes specify permitted materials and joints. Ductile iron has now replaced cast iron for pressure pipe. Ductile iron is stronger, less brittle, and similar to cast iron in corrosion resistance. It is commonly used for buried pressure water mains or in other locations where internal or external corrosion is a problem. Joints are made with flanged fittings, mechanical joint (MJ) fittings, or elastomer gaskets for bell and spigot ends. Bell and spigot and MJ joints are not self-restrained, though restrained MJ systems are available. Ductile-iron pipe is made in seven thickness classes for different service conditions. AWWA Standard C150/A21.50 covers the proper selection of pipe classes.
Nonmetallic (Plastic) Selecting a plastic for a specific purpose requires attention to the temperatures, pressures, chemicals, and stresses the piping will be subjected to in the specific application. All are suitable for cold water. Plastic pipe should not be used for compressed gases or compressed air if the pipe’s material is subject to brittle failure. For other liquids and chemicals, refer to charts provided by plastic pipe manufacturers and distributors. Table 18 gives properties of the various plastics discussed in this section; the last column gives the relative cost of small pipe in each category. Table 2 lists some applications pertinent to HVAC. The following are brief descriptions of common uses for the various materials. Plastic piping materials fall into two main categories: thermoplastics and thermosets. Thermoplastics melt and are formed by extruding or molding. They are usually used without reinforcing filaments. Thermosets are cured and cannot be reformed. They are normally used with glass fiber reinforcing filaments. For the purposes of this chapter, thermoplastic piping is made of the following materials: PVC. Because polyvinyl chloride has the best overall range of properties at the lowest cost, it is the most widely used plastic. It is joined by solvent cementing, threading, or flanging. Gasketed pushon joints are also used for larger sizes. ASTM Standards D1784, D1785, and D2665 cover PVC pipe. CPVC. Chlorinated polyvinyl chloride has the same properties as PVC and can withstand a higher temperature before losing strength. It is joined by the same methods as PVC. ASTM Standards D1784 and 1785 discuss CPVC. PE. Low-density polyethylene (LDPE) is a flexible, low-mass tubing with good low-temperature properties. It is used in the food and beverage industry and for instrument tubing. Joins are mechanical, such as compression fittings or push-on connectors and clamps. See ASTM Standard D2239 for details. HDPE. High-density polyethylene is a tough, weather-resistant material used for large pipelines in the gas industry. Fabricated fittings are available. It is joined by heat fusion for large sizes; flare, compression, or insert fittings can be used on small sizes. ASTM Standard D3350 discusses HDPE.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
22.16
2017 ASHRAE Handbook—Fundamentals (SI) Table 16
Licensed for single user. © 2017 ASHRAE, Inc.
U.S. Surface Area Wall Inside Nominal Nominal Thickness Diameter Outside, Inside, Size, Size, m2/m d, mm m2/m in. mm Schedulea t, mm 1/4
8
3/8
10
1/2
15
3/4
20
1
25
1 1/4
32
1 1/2
40
2
50
2 1/2
65
3
80
4
100
6
150
8
200
10
250
12
300
14
350
16
400
18
450
aNumbers
500
Cross Section Metal Flow Area, Area, mm2 mm2
Working Pressurec ASTM A53 B to 200°C
Mass Pipe, kg/m
Water, kg/m
Mfr. Process
Joint Typeb
kPa (gage)
0.631 0.796 0.844 1.098 1.265 1.618 1.68 2.19
0.067 0.046 0.123 0.091 0.196 0.151 0.344 0.279
CW CW CW CW CW CW CW CW
T T T T T T T T
1296 6006 1400 5654 1476 5192 1496 4695
2.50 3.23 3.38 4.45 4.05 5.40 5.43 7.47 8.62 11.40
0.558 0.464 0.965 0.828 1.313 1.140 2.165 1.905 3.089 2.734
CW CW CW CW CW CW CW CW CW CW
T T T T T T T T W W
1558 4427 1579 4096 1593 3972 1586 3799 3675 5757
40 ST 80 XS 40 ST 80 XS 40 ST 80 XS 40 ST 80 XS
2.24 3.02 2.31 3.20 2.77 3.73 2.87 3.91
9.25 7.67 12.52 10.74 15.80 13.87 20.93 18.85
0.043 0.043 0.054 0.054 0.067 0.067 0.084 0.084
0.029 0.024 0.039 0.034 0.050 0.044 0.066 0.059
80.6 101.5 107.7 140.2 161.5 206.5 214.6 279.7
67.1 46.2 123.2 90.7 196.0 151.1 344.0 279.0
40 ST 80 XS 40 ST 80 XS 40 ST 80 XS 40 ST 80 XS 40 ST 80 XS
3.38 4.55 3.56 4.85 3.68 5.08 3.91 5.54 5.16 7.01
26.64 24.31 35.05 32.46 40.89 38.10 52.50 49.25 62.71 59.00
0.105 0.105 0.132 0.132 0.152 0.152 0.190 0.190 0.229 0.229
0.084 0.076 0.110 0.102 0.128 0.120 0.165 0.155 0.197 0.185
318.6 412.1 431.3 568.7 515.5 689.0 690.3 953 1 099 1 454
557.6 464.1 965.0 827.6 1 313 1 140 2 165 1 905 3 089 2 734
40 ST 80 XS 40 ST 80 XS 40 ST 80 XS 30 40 ST 80 XS
5.49 7.62 6.02 8.56 7.11 10.97 7.04 8.18 12.70
77.93 73.66 102.26 97.18 154.05 146.33 205.0 202.7 193.7
0.279 0.279 0.359 0.359 0.529 0.529 0.688 0.688 0.688
0.245 0.231 0.321 0.305 0.484 0.460 0.644 0.637 0.608
1 438 1 946 2 048 2 844 3 601 5 423 4 687 5 419 8 234
4 769 4 261 8 213 7 417 18 639 16 817 33 000 32 280 29 460
11.27 15.25 16.04 22.28 28.22 42.49 36.73 42.46 64.51
4.769 4.261 8.213 7.417 18.64 16.82 33.01 32.28 29.46
CW CW CW CW ERW ERW ERW ERW ERW
W W W W W W W W W
3323 5288 2965 4792 4799 8336 3627 4433 7626
30 40 ST XS 80 30 ST 40 XS 80
7.80 9.27 12.70 15.06 8.38 9.53 10.31 12.70 17.45
257.5 254.5 247.7 242.9 307.1 304.8 303.2 298.5 289.0
0.858 0.858 0.858 0.858 1.017 1.017 1.017 1.017 1.017
0.809 0.800 0.778 0.763 0.965 0.958 0.953 0.938 0.908
6 498 7 683 10 388 12 208 8 307 9 406 10 158 12 414 16 797
52 060 50 870 48 170 46 350 74 060 72 970 72 190 69 940 65 550
50.91 60.20 81.39 95.66 65.09 73.70 79.59 97.28 131.62
52.06 50.87 48.17 46.35 74.06 72.97 72.21 69.96 65.57
ERW ERW ERW ERW ERW ERW ERW ERW ERW
W W W W W W W W W
3344 4178 6116 7453 3096 3641 4020 5157 7419
30 ST 40 XS 80 30 ST 40 XS
9.53 11.10 12.70 19.05 9.53 12.70
336.6 333.4 330.2 317.5 387.4 381.0
1.117 1.117 1.117 1.117 1.277 1.277
1.057 1.047 1.037 0.997 1.217 1.197
10 356 12 013 13 681 20 142 11 876 15 708
88 970 87 290 85 610 79 160 117 800 114 000
81.15 94.13 107.21 157.82 93.06 123.09
88.96 87.30 85.63 79.17 117.8 114.0
ERW ERW ERW ERW ERW ERW
W W W W W W
3316 3999 4695 7453 2903 4109
ST
9.53 11.10 12.70 14.27 9.53 12.70 15.06
438.2 435.0 431.8 428.7 489.0 482.6 477.9
1.436 1.436 1.436 1.436 1.596 1.596 1.596
1.376 1.367 1.357 1.347 1.536 1.516 1.501
13 396 15 556 17 735 19 863 14 916 19 762 23 325
150 800 148 600 146 450 144 300 187 700 182 900 179 400
104.98 121.90 138.97 155.65 116.88 154.85 182.78
150.8 148.6 146.4 144.3 187.4 182.9 179.4
ERW ERW ERW ERW ERW ERW ERW
W W W W W W W
2579 3110 3654 4185 2324 3289 4006
30
20
Steel Pipe Data
XS 40 20 ST 30 XS 40
are schedule numbers per ASME Standard B36.10M; ST = Standard; XS = (2)An arbitrary corrosion allowance of 0.64 mm for pipe sizes through NPS 2 and 1.65 mm Extra Strong. from NPS 2 1/2 through 20, plus = Thread; W = Weld (3) A thread cutting allowance for sizes through NPS 2. cWorking pressures were calculated per ASME Standard B31.9 using furnace buttweld (continuous weld, CW) pipe through 100 mm and electric resistance weld Because the pipe wall thickness of threaded standard pipe is so small after deducting the (ERW) thereafter. The allowance A has been taken as allowance A, the mechanical strength of the pipe is impaired. It is good practice to limit (1) 12.5% of t for mill tolerance on pipe wall thickness, plus standard threaded pipe pressure to 620 kPa (gage) for steam and 860 kPa (gage) for water. bT
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Pipe Design
22.17 Table 17 Copper Tube Data
Wall ThickU.S. ness Nominal Size, in. Type t, mm 1/4 3/8 1/2
5/8 3/4 1
Licensed for single user. © 2017 ASHRAE, Inc.
1 1/4
1 1/2
2
2 1/2 3
3 1/2 4
5
6
8
10 12
aWhen
K 0.89 L 0.76 K 1.24 L 0.89 M 0.64 K 1.24 L 1.02 M 0.71 K 1.24 L 1.07 K 1.65 L 1.14 M 0.81 K 1.65 L 1.27 M 0.89 K 1.65 L 1.40 M 1.07 DWV 1.02 K 1.83 L 1.52 M 1.24 DWV 1.07 K 2.11 L 1.78 M 1.47 DWV 1.07 K 2.41 L 2.03 M 1.65 K 2.77 L 2.29 M 1.83 DWV 1.14 K 3.05 L 2.54 M 2.11 K 3.40 L 2.79 M 2.41 DWV 1.47 K 4.06 L 3.18 M 2.77 DWV 1.83 K 4.88 L 3.56 M 3.10 DWV 2.11 K 6.88 L 5.08 M 4.32 DWV 2.77 K 8.59 L 6.35 M 5.38 K 10.29 L 7.11 M 6.45
Diameter Outside D, mm 9.53 9.53 12.70 12.70 12.70 15.88 15.88 15.88 19.05 19.05 22.23 22.23 22.23 28.58 28.58 28.58 34.93 34.93 34.93 34.93 41.28 41.28 41.28 41.28 53.98 53.98 53.98 53.98 66.68 66.68 66.68 79.38 79.38 79.38 79.38 92.08 92.08 92.08 104.78 104.78 104.78 104.78 130.18 130.18 130.18 130.18 155.58 155.58 155.58 155.58 206.38 206.38 206.38 206.38 257.18 257.18 257.18 307.98 307.98 307.98
Inside d, mm 7.75 8.00 10.21 10.92 11.43 13.39 13.84 14.45 16.56 16.92 18.92 19.94 20.60 25.27 26.04 26.80 31.62 32.13 32.79 32.89 37.62 38.23 38.79 39.14 49.76 50.42 51.03 51.84 61.85 62.61 63.37 73.84 74.80 75.72 77.09 85.98 87.00 87.86 97.97 99.19 99.95 101.83 122.05 123.83 124.64 126.52 145.82 148.46 149.38 151.36 192.61 196.22 197.74 200.84 240.00 244.48 246.41 287.40 293.75 295.07
Surface Area Outside, m2/m
Inside, m2/m
0.030 0.030 0.040 0.040 0.040 0.050 0.050 0.050 0.060 0.060 0.070 0.070 0.070 0.090 0.090 0.090 0.110 0.110 0.110 0.110 0.130 0.130 0.130 0.130 0.170 0.170 0.170 0.170 0.209 0.209 0.209 0.249 0.249 0.249 0.249 0.289 0.289 0.289 0.329 0.329 0.329 0.329 0.409 0.409 0.409 0.409 0.489 0.489 0.489 0.489 0.648 0.648 0.648 0.648 0.808 0.808 0.808 0.968 0.968 0.968
0.0244 0.0250 0.0320 0.0344 0.0360 0.0421 0.0436 0.0454 0.0521 0.0530 0.0594 0.0628 0.0646 0.0792 0.0817 0.0841 0.0994 0.1009 0.1030 0.1033 0.1183 0.1201 0.1219 0.1228 0.1564 0.1585 0.1603 0.1628 0.1942 0.1966 0.1990 0.2320 0.2350 0.2378 0.2423 0.2701 0.2733 0.2761 0.3078 0.3115 0.3139 0.3200 0.3834 0.3889 0.3917 0.3975 0.4581 0.4663 0.4694 0.4755 0.6050 0.6163 0.6212 0.6309 0.7541 0.7681 0.7742 0.9028 0.9229 0.9269
using soldered or brazed fittings, the joint determines the limiting pressure. pressures were calculated using ASME Standard B31.9 allowable stresses. A 5% mill tolerance has been used on the wall thickness. Higher tube ratings can be calculated using the allowable stress for lower temperatures.
bWorking
Cross Section Metal Flow Area, Area, mm2 mm2 24 21 45 33 24 57 48 34 70 60 106 75 55 139 109 77 173 147 114 108 226 190 157 135 343 292 243 177 487 413 337 666 554 446 281 852 714 596 1084 895 776 478 1610 1266 1108 737 2309 1698 1484 1016 4314 3212 2741 1771 6705 5004 4259 9621 6722 6112 cIf
47 50 82 94 103 141 151 164 215 225 281 312 333 502 532 564 785 811 845 850 1 111 1 148 1 181 1 203 1 945 1 997 2 045 2 111 3 004 3 079 3 154 4 282 4 395 4 503 4 667 5 806 5 944 6 063 7 538 7 727 7 846 8 144 11 699 12 042 12 201 12 572 16 701 17 311 17 525 17 993 29 137 30 238 30 710 31 680 45 241 46 942 47 686 64 873 67 771 68 382
Mass Tube, kg/m 0.216 0.188 0.400 0.295 0.216 0.512 0.424 0.302 0.622 0.539 0.954 0.677 0.488 1.249 0.973 0.691 1.543 1.316 1.015 0.967 2.025 1.701 1.399 1.204 3.070 2.606 2.171 1.585 4.35 3.69 3.02 5.96 4.95 3.98 2.51 7.62 6.39 5.33 9.69 8.00 6.94 4.27 14.39 11.32 9.91 6.59 20.64 15.18 13.27 9.09 38.56 28.71 24.50 15.83 59.93 44.73 38.07 85.99 60.09 54.63
Water, kg/m 0.047 0.050 0.082 0.094 0.103 0.141 0.151 0.164 0.215 0.225 0.281 0.312 0.333 0.502 0.532 0.564 0.785 0.811 0.845 0.850 1.111 1.148 1.182 1.203 1.945 1.997 2.045 2.111 3.004 3.079 3.154 4.282 4.395 4.503 4.667 5.806 5.944 6.063 7.538 7.727 7.846 8.144 11.70 12.04 12.20 12.57 16.70 17.31 17.53 17.99 29.14 30.24 30.71 31.62 45.15 46.94 47.69 64.87 67.77 68.38
Working Pressurea,b,c ASTM B88 to 120°C MPa (gage) Annealed Drawn 5.868 5.033 6.164 4.399 3.144 4.930 4.027 2.820 4.109 3.523 4.668 3.234 2.303 3.634 2.792 1.958 2.972 2.517 1.924 1.827 2.786 2.324 1.896 1.627 2.455 2.069 1.717 1.241 2.275 1.917 1.558 2.193 1.813 1.448 0.903 2.082 1.738 1.441 2.041 1.675 1.448 0.883 1.965 1.531 1.338 0.883 1.972 1.434 1.255 0.855 2.096 1.544 1.317 0.841 2.096 1.551 1.317 2.103 1.455 1.317
11.004 9.432 11.556 8.253 5.895 9.246 7.543 5.282 7.702 6.605 8.757 6.061 4.309 6.812 5.240 3.668 5.571 4.716 3.599 3.427 5.226 4.351 3.558 3.048 4.606 3.951 3.220 2.331 4.268 3.592 2.917 4.109 3.392 2.717 1.696 3.903 3.254 2.703 3.827 3.144 2.717 1.655 3.682 2.875 2.510 1.655 3.696 2.696 2.351 1.600 3.930 2.903 2.468 1.579 3.937 2.910 2.468 3.937 2.724 2.468
soldered or brazed fittings are used on hard-drawn tubing, use the annealed ratings. Full-tube allowable pressures can be used with suitably rated flare or compression-type fittings.
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22.18
2017 ASHRAE Handbook—Fundamentals (SI) Table 18 Properties of Pipe Materialsa
Material Designation
Type and Grade
Metals Copper Steel Stainless steel
Type L A 53 B 304
Drawn ERW Drawn or Welded
248 413
Thermoplastics PVC 1120 PVC 1200 PVC 2120 CPVC 4120
T I,G1 T I,G2 T II,G1 T IV,G1
12454-B 12454-C 14333-D 23447-B
52
14
55
14
PE 2306 PE 3306 PE 3406 HDPE 3408 PP ABS
Licensed for single user. © 2017 ASHRAE, Inc.
Hydrostaticb Design Stress, Tensile MPa (at 23°C) Strength, ASME MPa (at B31 23°C) Mfr.
ABS 1210 ABS 1316 ABS 2112 PVDF
Gr. P23 Gr. P34 Gr. P33 Gr. P34 Acrylonitrile copolymer T I,G2 T I,G3 T II,G1
Cell No.
355434-C 6-3-3
ERW Drawn
HDSb Upper ASME Limit, Mfr. B31 MPa
62 88
11 4.9
5-2-2 3-5-5 4-4-5
Thermosetting Epoxy-glass RTRP-11AF PEX A,B,Cd Polyester-glass RTRP-12EF For Comparison Steel A 53 B Copper Type L
34 34 38
Upper Temperature Limit, °C
14 14 14 14 4.3 4.3 4.3 5.5
204 427 177
60 99
60 100 80
7 11 8.6 48
8.8
138
303 22 303
55 4.3 62
99 93 93
413 248
88 62
66 66 66 99 60 70 82 82 99
56 63
8900 7800 7900
1600
3.0
1400
43
2.90 2.83
2.2
1550
80
640 70 450
5.5
82 82 82 135
4.4 6.9 5.5 2.1
82
427 204
a Properties listed
are for the specific materials listed; each plastic has other formulations. Consult the manufacturer of the system chosen. These values are for comparative purposes. b Hydrostatic design stress (HDS) is equivalent to allowable design stress.
PP. Polypropylene is a low-mass plastic used for pressure applications and also for chemical waste lines, because it is inert to a wide range of chemicals. A broad variety of drainage fittings are available. For pressure uses, regular fittings are made. It is joined by heat fusion. See ASTM Standards F2830 and F2389 for details. ABS. Acrylonitrile butadiene styrene is a high-strength, impactand weather-resistant material. Some formulations can be used for beverage industry. A wide range of fittings is available. It is joined by solvent cement, threading, or flanging. ASTM Standards D2661 and D3965 cover ABS. PVDF. Polyvinylidene fluoride is widely used for ultrapure water systems and in the pharmaceutical industry and has a wide temperature range. This material is over 20 times more expensive than PVC. It is joined by heat fusion, and fittings are made for this purpose. For smaller sizes, mechanical joints can be used. See ASTM Standard D2122 for information on PVDF. Thermosetting piping used in HVAC is called (1) reinforced thermosetting resin (RTR) and (2) fiberglass-reinforced plastic (FRP). RTR and FRP are interchangeable and refer to pipe and fittings commonly made of (1) fiberglass-reinforced epoxy resin, (2) fiberglassreinforced vinyl ester, and (3) fiberglass-reinforced polyester. Pipe and fittings made from epoxy resin are generally stronger and operate at a higher temperature than those made from polyester or vinyl ester resins, so they are more likely to be used in HVAC. PEX. Cross-linked polyethylene is made from high-density polyethylene (HDPE) and contains cross-linked bonds in the polymer structure. This changes the thermoplastic to a thermoset. It can be used up to 150°C. PEX is used in building services pipework systems, hydronic radiant heating and cooling systems, and domestic water piping. PEX comes in two types: barrier and nonbarrier. The barrier, a thin sheet of aluminum between layers of PEX material or
Impact Modulus of Coefficient of Thermal Relative Density, Strength, N Elasticity, Expansion, Conductivity, Pipe 3 kg/m (at 23°C) GPa (at 23°C) m/(m·K) W/(m·K) Costc
960 910 1060
117 190 193
17.1 11.4 17.6
33.5 3.8 1.2
3.5 1.3
0.159
1.0
2.92
54 63 54 63
0.137
2.9
0.62 0.90 1.03 0.76 0.83 1.65
144 126 108 216 108 101
0.389 0.187 0.245
1.1 2.9 3.4
1.72 2.34
0.115
28.0
0.75
1780
200
0.86
99 72 72 142
48 0.54 34
940
200
6.90 0.52 6.90
16 to 23 162 16 to 20
0.418 0.462 0.187
63 56
7800 8900
1600
11.4 17.1
49.6
190 117
1.3 3.5
c
Based on cost of pipe only, without factoring in fittings, joints, hangers, and labor. d A, B, and C are the three manufacturing processes of PEX pipe. The classifications are not related to a ranking system.
a layer of polymer film, prevents oxygen dissolved in water from diffusing through the pipe and corroding metal components. Nonbarrier PEX is acceptable for plumbing systems. PEX can be ordered as A, B, or C (these designations refer to the manufacturing process and not the pipe’s structural or chemical properties). All PEX tubing (A, B, C) comply with the same standards: refer to ASTM Standards F876, F877, and F2023; CSA Standard B137.5; and NSF/ANSI Standards 14 and 61 for further information.
2.2
FITTINGS
Table 19 lists standards that give dimensions and pressure ratings for fittings, flanges, and flanged fittings. These data are also available from manufacturers’ catalogs.
2.3
JOINING METHODS
Threading Threading as per ASME Standard B1.202M is the most common method for joining small-diameter steel or brass pipe. Pipe with a wall thickness less than standard should not be threaded. ASME Standard B31.5 limits the threading for various refrigerants and pipe sizes.
Soldering and Brazing Copper tube is usually joined by soldering or brazing socket end fittings. Brazing materials melt above 540°C and produce a stronger joint than solder. Table 20 lists soldered and brazed joint strengths. ASME Standard B16.22-specified wrought copper solder joint fittings and ASME Standard B16.18-specified cast copper solder joint fittings are pressure rated the same way as annealed Type L copper
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Pipe Design
22.19 Table 19 Applicable Standards for Fittings
Steela Pipe flanges and flanged fittings Factory-made wrought steel butt-welding fittings Forged fittings, socket-welding and threaded Wrought steel butt-welding short radius elbows and returns Cast Iron, Malleable Iron, Ductile Ironb Cast iron pipe flanges and flanged fittings Malleable iron threaded fittings Gray iron threaded fittings Cast iron threaded drainage fittings
ASME Std. B16.5 B16.9 B16.11 B16.9 ASME Std. B16.1 B16.3 B16.4 B16.12
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Ductile iron pipe flanges and flanged fittings, Classes 150 and 300 B16.42 ASME Std. Copper and Bronzec Cast bronze threaded fittings, Classes 125 and 25 B16.15 Cast copper alloy solder joint pressure fittings B16.18 Wrought copper and copper alloy solder joint pressure fittings B16.22 Cast copper alloy solder joint drainage fittings, DWV B16.23 Cast copper alloy pipe flanges and flanged fittings, Classes 150, 300, 400, 600, 900, 1500, and 2500 B16.24
Copper and Bronzec(Continued) Cast copper alloy fittings for flared copper tubes
ASME Std. B16.26
Wrought copper and wrought copper alloy solder joint drainage fittings B16.29 ASTM Std. Nonmetallicd Threaded PVC plastic pipe fittings, Schedule 80 D2464 Threaded PVC plastic pipe fittings, Schedule 40 D2466 Socket-Type PVC plastic pipe fittings, Schedule 80 D2467 Reinforced epoxy resin gas pressure pipe and fittings D2517 Threaded CPVC plastic pipe fittings, Schedule 80 F437 Socket-Type CPVC plastic pipe fittings, Schedule 40 F438 Socket-Type CPVC plastic pipe fittings, Schedule 80 F439 Insert fittings for PEX tubing F877 Plastic brass, bronze, and copper insert fittings for PEX tubing F877 Solvent cements for PVC plastic piping systems D2564 Solvent cements for CPVC plastic pipe and fittings F493
aWrought steel butt-welding fittings are made to match steel pipe wall thicknesses and are rated at the same working pressure as seamless pipe. Flanges and flanged fittings are rated
by working steam pressure classes. Forged steel fittings are rated from 14 to 41 MPa in classes and are used for high-temperature and high-pressure service for small pipe sizes. numbers refer to maximum working saturated steam gage pressure (in pounds per square inch. Multiply these values by 6.9 to convert to kilopascals). For liquids at lower temperatures, higher pressures are allowed. Groove-end fittings of these materials are made by various manufacturers who publish their own ratings. cClasses refer to maximum working steam gage pressure (in pounds per square inch. Multiply these values by 6.9 to convert to kilopascals). At ambient temperatures, higher liquid pressures are allowed. Solder joint fittings are limited by the strength of the soldered or brazed joint (see Table 20). dRatings of plastic fittings match the pipe of corresponding schedule number.
bClass
Table 20 Internal Working Pressure for Copper Tube Joints Internal Working Pressure, kPa Sat. Steam and Condensate
Water and Noncorrosive Liquids and Gasesa
Service Temperature, °C
8 to 25
32 to 50
65 to 100
125 to 200a
solder 50-50 (ASTM B32 Gr 50A)
38 66 93 120
1380 1030 690 590
1210 860 620 520
1030 690 520 350
900 620 480 310
690 480 350 280
— — — 100
95-5 tin/antimonyc solder (ASTM B32 Gr 50TA)
38 66 93 120
3450 2760 2070 1380
2760 2410 1720 1200
2070 1900 1380 1030
1860 1720 1240 930
1030 1030 970 760
— — — 100
38 to 93 120 175
d
d
d
d
d
2070 1860
1450 1310
1170 1030
1030 1030
1030 1030
— — 830
Alloy Used for Joints tin/leadb
Brazing alloys melting at or above 540°C
Nominal Tube Size (Types K, L, M), mm
Source: Based on ASME Standard B31.9 aSolder joints are not to be used for (1) Flammable or toxic gases or liquids (2) Gas, vapor, or compressed air in tubing over 100 mm, unless maximum pressure is limited to 140 kPa (gage).
tube of the same size. Health concerns have caused many jurisdictions to ban solder containing lead or antimony for joining pipe in potable-water systems. Lead-based solder, in particular, must not be used for potable water.
Flared and Compression Joints Flared and compression fittings can be used to join copper, steel, stainless steel, and aluminum tubing. Properly rated fittings can keep the joints as strong as the tube.
250 to 300a
8 to 200
bLead
solders must not be used in potable-water systems. solder is allowed for potable-water supplies in some jurisdictions. dRated pressure for temperatures up to 93°C is that of the tube being joined.
c Tin/antimony
Flanges Flanges can be used for large pipe and all piping materials. They are commonly used to connect to equipment and valves, and wherever the joint must be opened to allow service or replacement of components. For steel pipe, flanges are available in pressure ratings to 17 MPa. High-tensile-strength bolts must be used for highpressure flanged joints. For welded pipe, weld neck, slip-on, or socket weld flanges are available. Thread-on flanges are available for threaded pipe.
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22.20 Flanges are generally flat faced or raised face. Flat-faced flanges with full-faced gaskets are most often used with cast iron and materials that cannot take high bending loads. Raised-face flanges with ring gaskets are preferred with steel pipe because they facilitate increasing the sealing pressure on the gasket to help prevent leaks. Other facings, such as O ring and ring joint, are available for special applications. All flat-faced, raised-face, and lap-joint flanges require a gasket between the mating flange surfaces. Gaskets are made from rubber, synthetic elastomers, cork, fiber, plastic, polytetrafluoroethylene (PTFE), metal, and combinations of these materials. The gasket must be compatible with the flowing media and the temperatures at which the system operates.
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Welding Welded-steel pipe joints offer the following advantages: • Do not age, dry out, or deteriorate as gasketed joints do • Can accommodate greater vibration and water hammer and higher temperatures and pressures than other joints • For critical service, can be tested by several nondestructive examination (NDE) methods, such as radiography or ultrasound • Provide maximum long-term reliability The applicable sections of the ASME Standard B31 series and the ASME Boiler and Pressure Vessel Code give rules for welding. ASTM Standard B31 requires that all welders and welding procedure specifications (WPS) be qualified. Separate WPS are needed for different welding methods and materials. The qualifying tests and the variables requiring separate procedure specifications are set forth in the ASME Boiler and Pressure Vessel Code, Section IX. The manufacturer, fabricator, or contractor is responsible for the welding procedure and welders. ASME Standard B31.9 requires visual examination of welds and outlines limits of acceptability. The following welding processes are often used in the HVAC industry: • Shielded metal arc welding (SMAW), also called stick welding): the molten weld metal is shielded by vaporization of the electrode coating. • Gas metal arc welding (GMAW), also called metal inert gas (MIG) welding: the electrode is a continuously fed wire shielded by argon or carbon dioxide gas from the welding gun nozzle. • Gas tungsten arc welding (GTAW), also called tungsten insert gas (TIG) welding: this process uses a nonconsumable tungsten electrode surrounded by a shielding gas. The weld material may be provided from a separate noncoated rod.
Integrally Reinforced Outlet Fittings Integrally reinforced outlet fittings are used to make branch and take-off connections and are designed to allow welding directly to pipe without supplemental reinforcing. Fittings are available with threaded, socket welded, or butt-weld outlets.
Solvent Cement Solvent cement welds nonmetallic pipe together by softening surface of the materials being joined. It is different from gluing, which hardens and holds the material together. Sometimes this join is called a solvent-welded joint.
Rolled-Groove Joints Grooved joints require special grooved fittings and a shallow groove cut or rolled into the pipe end. These joints can be used with steel, cast iron, ductile iron, copper, and plastic pipes. A segmented clamp engages the grooves and a special gasket uses internal pressure to tighten the seal. Some clamps are designed with clearance between tongue and groove to accommodate misalignment and thermal movements, and others are designed to limit movement and
2017 ASHRAE Handbook—Fundamentals (SI) provide a rigid system. Manufacturers’ data give temperature and pressure limitations.
Bell-and-Spigot Joints A bell-and-spigot joint is mechanical joint consists of a sleeve slightly larger than the outside diameter of the pipe. The pipe ends are inserted into the sleeve, and gaskets are packed into the annular space between the pipe and coupling and held in place by retainer rings. This type of joint can accept some axial misalignment, but it must be anchored or otherwise restrained to prevent axial pullout or lateral movement. Manufacturers provide pressure/temperature data.
Press-Connect (Press Fit) Joints These joints rely on an elastomeric gasket or seal and an approved pressing tool and jaws to seal the joint.
Push-Connect Joints Push-connect joining use and integral elastomeric seal or gasket and stainless steel ring to make a leak-free joint. There are two common types, both of which form strong, permanent joints: one type is removable for servicing, and the other type is not easily removed after installation.
Unions Unions allow disassembly of threaded pipe systems. Unions are three-part fittings with a mating machined seat on the two parts that thread onto the pipe ends. A threaded locking ring holds the two ends tightly together. A union also allows threaded pipe to be turned at the last joint connecting two pieces of equipment. Companion flanges (a pair) for small pipe serve the same purpose.
2.4
EXPANSION JOINTS AND EXPANSION COMPENSATING DEVICES
Although the inherent flexibility of the piping should be used to the maximum extent possible, expansion joints must be used where movements are too large to accommodate with pipe bends or loops or where insufficient room exists to construct a loop of adequate size. Typical situations are tunnel piping and risers in high-rise buildings, especially for steam and hot-water pipes where large thermal movements are involved. Packed and packless expansion joints and expansion compensating devices are used to accommodate movement, either axially or laterally. In the axial method of accommodating movement, the expansion joint is installed between anchors in a straight-line segment and accommodates axial motion only. This method has high anchor loads, primarily because of pressure thrust. It requires careful guiding, but expansion joints can be spaced conveniently to limit movement of branch connections. The axial method finds widest application for long runs without natural offsets, such as tunnel and underground piping and risers in tall buildings. The lateral or offset method requires the device to be installed in a leg perpendicular to the expected movement and accommodates lateral movement only. This method generally has low anchor forces and minimal guide requirements. It finds widest application in lines with natural offsets, especially where there are few or no branch connections. Packed expansion joints depend on slipping or sliding surfaces to accommodate the movement and require some type of seals or packing to seal the surfaces. Most such devices require some maintenance but are not subject to catastrophic failure. Further, with most packed expansion joint devices, any leaks that develop can be repacked under full line pressure without shutting down the system. Packless expansion joints depend on the flexing or distortion of the sealing element to accommodate movement. They generally do
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Pipe Design
22.21
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 7 Packed Slip Expansion Joint not require any maintenance, but maintenance or repair is not usually possible. If a leak occurs, the system must be shut off and drained, and the entire device must be replaced. Further, catastrophic failure of the sealing element can occur and, although likelihood of such failure is remote, it must be considered in certain design situations. Packed expansion joints are preferred where long-term system reliability is of prime importance (using types that can be repacked under full line pressure) and where major leaks can be life threatening or extremely costly. Typical applications are risers, tunnels, underground pipe, and distribution piping systems. Packless expansion joints are generally used where even small leaks cannot be tolerated (e.g., for gas and toxic chemicals), where temperature limitations preclude the use of packed expansion joints, and for very-largediameter pipe where packed expansion joints cannot be constructed or the cost would be excessive. In all cases, expansion joints should be installed, anchored, and guided in accordance with expansion joint manufacturers’ recommendations.
joint from the system. The packing technology of the packed slip expansion joint, explained previously, has been incorporated into the flexible ball joint design; now, packed flexible ball joints have self-lubricating semiplastic packing that can be injected under full line pressure without shutting off the system (Figure 8). Standard flexible ball joints are available in sizes 32 to 760 mm with threaded (32 to 50 mm), weld, and flange ends for pressures to 2.1 MPa and temperatures to 400°C. Flexible ball joints are available in larger sizes and for higher temperature and pressure ranges.
Packed Expansion Joints
Packless Expansion Joints
There are two types of packed expansion joints: packed slip expansion joints and flexible ball joints. Packed Slip Expansion Joints. These are telescoping devices designed to accommodate axial movement only. Packing seals the sliding surfaces. The original packed slip expansion joint used multiple layers of braided compression packing, similar to the stuffing box commonly used with valves and pumps; this arrangement requires shutting and draining the system for maintenance and repair. Advances in design and packing technology have eliminated these problems, and most current packed slip joints use selflubricating semiplastic packing, that can be injected under full line pressure without shutting off the system (Figure 7). (Many manufacturers use asbestos-based packings, unless requested otherwise. Asbestos-free packings, such as flake graphite, are available and, although more expensive, should be specified in lieu of products containing asbestos.) Standard packed slip expansion joints are constructed of carbon steel with weld or flange ends in sizes 40 to 910 mm for pressures up to 2.1 MPa and temperatures up to 425°C. Larger, highertemperature, and higher-pressure designs are available. Standard single joints are generally designed for 100, 200, or 300 mm axial traverse; double joints with an intermediate anchor base can accommodate twice these movements. Special designs for greater movements are available. Flexible Ball Joints. These joints are used in pairs to accommodate lateral or offset movement and must be installed in a leg perpendicular to the expected movement. The original flexible ball joint design incorporated only inner and outer containment seals that could not be serviced or replaced without removing the ball
Types include metal bellows expansion joints, rubber expansion joints, and flexible hose or pipe connectors. Metal Bellows Expansion Joints. These expansion joints have a thin-walled convoluted section that accommodates movement by bending or flexing. The bellows material is generally Type 304, 316, or 321 stainless steel, but other materials are commonly used to satisfy service conditions. Small-diameter expansion joints 20 to 80 mm are generally called expansion compensators and are available in all-bronze or steel construction. Metal bellows expansion joints can generally be designed for the pressures and temperatures commonly encountered in HVAC systems and can also be furnished in rectangular configurations for ducts and chimney connectors. Overpressurization, improper guiding, and other forces can distort the bellows element. For low-pressure applications, such distortion can be controlled by the geometry of the convolution or the thickness of the bellows material. For higher pressure, internally pressurized joints require reinforcing. Externally pressurized designs are not subject to such distortion and are not generally furnished without supplemental bellows reinforcing. Single- and double-bellows expansion joints primarily accommodate axial movement only, similar to packed slip expansion joints. Although bellows expansion joints can accommodate some lateral movement, the universal tied bellows expansion joint better accommodates large lateral movement. This device operates much like a pair of flexible ball joints, except that bellows elements are used instead of flexible ball elements. The tie rods on this joint contain the pressure thrust, so anchor loads are much lower than with axial-type expansion joints.
Fig. 8
Flexible Ball Joint
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22.22
2017 ASHRAE Handbook—Fundamentals (SI) Table 21 Piping System Design Maximum Flow Rate for Energy Conservationa,b 2000
Operating Hours/year Pipe Size, mm Nominal
Other
IPS Sched. 40 Std. ID
63 76 102 127 152 203 255 305 305
62.7 77.9 102.3 128.2 153.8 202.7 254.5 303.2
2000 and 4400
Variable Flow/ Variable Speed
4400
Variable Flow/ Variable Speed
Other
Variable Flow/ Variable Speed
Other
L/s
m/s
L/s
m/s
L/s
m/s
L/s
m/s
L/s
m/s
L/s
m/s
7.57 11.35 22.08 25.87 47.32 75.71 113.55 157.72 NA
2.5 2.5 2.7 2.0 2.5 2.3 2.2 2.2 2.5
11.35 17.03 33.45 39.12 64.00 113.55 170.35 239.75 NA
3.7 3.5 4.1 3.0 3.7 3.5 3.3 3.3 13.04.0
5.35 8.83 16.40 19.55 39.55 56.78 82.02 119.87 NA
1.7 1.8 2.0 1.5 1.9 1.8 1.5 1.7 2.0
8.20 13.25 25.25 29.65 54.25 88.33 126.18 182.95 NA
2.5 2.8 3.1 2.3 2.9 2.7 2.5 2.5 2.9
4.29 6.95 13.25 15.77 27.75 44.15 63.09 94.65 NA
1.5 1.5 1.5 1.2 1.5 1.5 1.2 1.3 1.5
6.95 10.72 20.19 23.35 42.90 69.40 100.95 145.11 NA
2.2 2.2 2.5 1.8 2.3 2.1 2.0 2.0 2.3
aSource:
bThis
Based on ASHRAE Standard 90.1-2013 Table 6.5.4.5 with the addition of IPS and calculation for velocity in metres per second. table does not apply to district energy systems, and velocities in larger-bore piping can exceed these values per an interpretation of the ASHRAE 90.1 committee.
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Table 22 Water Velocities Based on Type of Service Type of Service
Velocity, m/s
General service City water
1.2 to 3.0 0.9 to 2.1 0.6 to 1.5 1.8 to 4.6 1.2 to 2.1
Boiler feed Pump suction and drain lines a Crane
b Carrier
Co. (1976).
(1960).
c Grinnell
Reference a, b, c a, b c a, c a, b Company (1951).
Rubber Expansion Joints. Similar to single-metal bellows expansion joints, rubber expansion joints incorporate a nonmetallic elastomeric bellows sealing element and generally have more stringent temperature and pressure limitations. Although rubber expansion joints can be used to accommodate expansion and contraction of the piping, they are primarily used as flexible connectors at equipment to isolate sound and vibration and eliminate stress at equipment nozzles. Flexible Hose. This type of hose can be constructed of elastomeric material or corrugated metal with an outer braid for reinforcing and end restraint. Flexible hose is primarily used as a flexible connector at equipment to isolate sound and vibration and eliminate stress at equipment nozzles; however, flexible metal hose is well suited for use as an offset-type expansion joint, especially for copper tubing and branch connections off risers.
3.
APPLICATIONS 3.1
WATER PIPING
Flow Rate Limitations Stewart and Dona (1987) surveyed the literature relating to water flow rate limitations. Noise, erosion, and installation and operating costs all limit the maximum and minimum velocities in piping systems. If piping sizes are too small, noise levels, erosion levels, and pumping costs can be unfavorable. If piping sizes are too large, installation costs are excessive. Therefore, pipe sizes are chosen to minimize initial cost while avoiding the undesirable effects of high velocities. ASHRAE Standard 90.1 has been accepted by authorities having jurisdiction (AHJs) as a code and, as such, limits the flow for energy conservation. The table (Table 21) is reproduced with modification showing velocity limitations. Various upper limits of water velocity and/or pressure drop in piping and piping systems are used. One recommendation places a velocity limit of 1.2 m/s for 50 mm pipe and smaller, and a pressure drop limit of 400 Pa/m for piping over 50 mm. Other guidelines are based on the type of service (Table 22) or annual operating hours
Table 23
Maximum Water Velocity to Minimize Erosion
Normal Operation, h/yr
Water Velocity, m/s
1500 2000 3000 4000 6000
4.6 4.4 4.0 3.7 3.0
Source: Carrier (1960).
(Table 23). These limitations are imposed either to control the levels of pipe and valve noise, erosion, and water hammer pressure or for economic reasons. Carrier (1960) recommends that the velocity not exceed 4.6 m/s in any case.
Noise Generation Velocity-dependent noise in piping and piping systems results from any or all of four sources: turbulence, cavitation, release of entrained air, and water hammer. In investigations of flow-related noise, Ball and Webster (1976), Marseille (1965), and Rogers (1953, 1954, 1956) reported that velocities on the order of 3 to 5 m/s lie within the range of allowable noise levels for residential and commercial buildings. The experiments showed considerable variation in noise levels obtained for a specified velocity. Generally, systems with longer pipe and with more numerous fittings and valves were noisier. In addition, sound measurements were taken under widely differing conditions; for example, some tests used plastic-covered pipe, whereas others did not. Thus, no detailed correlations relating sound level to flow velocity in generalized systems are available. Noise generated by fluid flow in a pipe increases sharply if cavitation or release of entrained air occurs. Usually, the combination of high water velocity with a change in flow direction or a decrease in pipe cross section, causing a sudden pressure drop, is necessary to cause cavitation. Ball and Webster (1976) found that at their maximum velocity of 13 m/s, cavitation did not occur in straight 10 and 15 mm pipe; using the apparatus with two elbows, cold-water velocities up to 6.5 m/s caused no cavitation. Cavitation did occur in orifices of 1:8 area ratio (orifice flow area is one-eighth of pipe flow area) at 1.5 m/s and in 1:4 area ratio orifices at 3 m/s (Rogers 1954). Some data are available for predicting hydrodynamic (liquid) noise generated by control valves. The International Society of Automation compiled prediction correlations in an effort to develop control valves for reduced noise levels (ISA 2007). The correlation to predict hydrodynamic noise from control valves is SL = 10 log Av + 20 log p – 30 log t + 76.6
(20)
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Pipe Design where SL Av Q p t
Water Hammer = = = = =
sound level, dB valve coefficient, m3/(s·Pa)0.5 flow rate, m3/s pressure drop across valve, Pa downstream pipe wall thickness, mm
Air entrained in water usually has a higher partial pressure than the water. Even when flow rates are small enough to avoid cavitation, the release of entrained air may create noise. Every effort should be made to vent the piping system or otherwise remove entrained air.
Erosion Erosion in piping systems is caused by water bubbles, sand, or other solid matter impinging on the inner surface of the pipe. Generally, at velocities lower than 3 m/s, erosion is not significant as long as there is no cavitation. When solid matter is entrained in the fluid at high velocities, erosion occurs rapidly, especially in bends. Thus, high velocities should not be used in systems where sand or other solids are present or where slurries are transported.
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22.23
With age, the internal surfaces of pipes become increasingly rough. This reduces the available flow with a fixed pressure supply. However, designing with excessive age allowances may result in oversized piping. Age-related decreases in capacity depend on type of water, type of pipe material, temperature of water, and type of system (open or closed) and include • Sliming (biological growth or deposited soil on the pipe walls): occurs mainly in unchlorinated, raw water systems. • Caking of calcareous salts: occurs in hard water (i.e., water bearing calcium salts) and increases with water temperature. • Corrosion (incrustations of ferrous and ferric hydroxide on the pipe walls): occurs in metal pipe in soft water. Because oxygen is necessary for corrosion to take place, significantly more corrosion takes place in open systems. Allowances for expected decreases in capacity are sometimes treated as a specific amount (percentage). Dawson and Bowman (1933) added an allowance of 15% friction loss to new pipe (equivalent to an 8% decrease in capacity). The HDR Design Guide (1981) increased the friction loss by 15 to 20% for closed piping systems and 75 to 90% for open systems. Carrier (1960) indicates a factor of approximately 1.75 between friction factors for closed and open systems. Obrecht and Pourbaix (1967) differentiated between the corrosive potential of different metals in potable water systems and concluded that iron is the most severely attacked, then galvanized steel, lead, copper, and finally copper alloys (e.g., brass). Freeman (1941) and Hunter (1941) showed the same trend. After four years of coldand hot-water use, copper pipe had a capacity loss of 25 to 65%. Aged ferrous pipe has a capacity loss of 40 to 80%. Smith (1983) recommended increasing the design discharge by 1.55 for uncoated cast iron, 1.08 for iron and steel, and 1.06 for cement or concrete. The Plastic Pipe Institute (1971) found that corrosion is not a problem in plastic pipe; the capacity of plastic pipe in Europe and the United States remains essentially the same after 30 years in use. Extensive age-related flow data are available for use with the Hazen-Williams empirical equation. Difficulties arise in its application, however, because the original Hazen-Williams roughness coefficients are valid only for the specific pipe diameters, water velocities, and water viscosities used in the original experiments. Thus, when the Cs are extended to different diameters, velocities, and/or water viscosities, errors of up to about 50% in pipe capacity can occur (Sanks 1978; Williams and Hazen 1933).
When any moving fluid (not just water) is abruptly stopped, as when a valve closes suddenly, large pressures can develop. Although detailed analysis requires knowledge of the elastic properties of the pipe and the flow-time history, the limiting case of rigid pipe and instantaneous closure is simple to calculate. Under these conditions, ph = csV
(21)
where ph cs V
= = = =
pressure rise caused by water hammer, Pa fluid density, kg/m3 velocity of sound in fluid, m/s fluid flow velocity, m/s
The cs for water is 1439 m/s, although the pipe’s elasticity reduces the effective value. Example 3. What is the maximum pressure rise if water flowing at 3 m/s is stopped instantaneously? Solution: ph = 1000 1439 3 = 4.32 MPa
3.2
SERVICE WATER PIPING
Sizing service water piping differs from sizing process lines in that design flows in service water piping are determined by the probability of simultaneous operation of multiple individual loads such as water closets, urinals, lavatories, sinks, and showers. The full-flow characteristics of each load device are readily obtained from manufacturers; however, service water piping sized to handle all load devices simultaneously would be seriously oversized. Thus, a major issue in sizing service water piping is to determine the diversity of the loads. The procedure shown in this chapter uses the work of R.B. Hunter for estimating diversity (Hunter 1940, 1941). The present-day plumbing designer is usually constrained by building or plumbing codes, which specify the individual and collective loads to be used for pipe sizing. Frequently used codes (including the ICC International Plumbing Code and the PHCC National Standard Plumbing Code) contain procedures quite similar to those shown here. The designer must be aware of the applicable code for the location being considered. Federal mandates are forcing plumbing fixture manufacturers to reduce design flows to many types of fixtures, but these may not yet be included in locally adopted codes. Also, the designer must be aware of special considerations; for example, toilet usage at sports arenas will probably have much less diversity than codes allow and thus may require larger supply piping than the minimum specified by codes. Table 24 gives the rate of flow desirable for many common fixtures and the average pressure necessary to give this rate of flow. Pressure varies with fixture design. In estimating load, the rate of flow is frequently computed in fixture units that are relative indicators of flow. Table 25 gives the demand weights in terms of fixture units for different plumbing fixtures under several conditions of service, and Figure 9 gives the estimated demand corresponding to any total number of fixture units. Figures 10 and 11 provide more accurate estimates at the lower end of the scale. The estimated demand load for fixtures used intermittently on any supply pipe can be obtained by multiplying the number of each kind of fixture supplied through that pipe by its weight from Table 25, adding the products, and then referring to the appropriate curve of Figure 9, 10, or 11 to find the demand corresponding to the total fixture units. In using this method, note that the demand for fixture or supply outlets other than those listed in the table of fixture units is not yet included in the estimate. The demands for outlets (e.g., hose connections and air-conditioning apparatus) that are likely to impose
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22.24
2017 ASHRAE Handbook—Fundamentals (SI)
Table 24 Proper Flow and Pressure Required During Flow for Different Fixtures Flow Pressure, kPa (gage)a Flow, L/s
Fixture Ordinary basin faucet Self-closing basin faucet Sink faucet, 10 mm Sink faucet, 15 mm Dishwasher Bathtub faucet Laundry tube cock, 8 mm Shower Ball cock for closet Flush valve for closet Flush valve for urinal Garden hose, 15 m, and sill cock a Flow
55 85 70 35 105 to 175 35 35 85 105 70 to 140 105 210
0.2 0.2 0.3 0.3 —b 0.4 0.3 0.2 to 0.6 0.2 1.0 to 2.5c 1.0 0.3
pressure is that in pipe at entrance to fixture. see manufacturers’ data. range because of variation in design and type of flush valve closets.
b Varies; c Wide
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Table 25 Demand Weights of Fixtures in Fixture Unitsa
Fixture or Groupb
Occupancy
Water closet
Public
Type of Supply Control
Flush valve Flush tank Pedestal urinal Public Flush valve Stall or wall urinal Public Flush valve Flush tank Lavatory Public Faucet Bathtub Public Faucet Shower head Public Mixing valve Service sink Office, etc. Faucet Kitchen sink Hotel or restaurant Faucet Water closet Private Flush valve Flush tank Lavatory Private Faucet Bathtub Private Faucet Shower head Private Mixing valve Bathroom group Private Flush valve for closet Flush tank for closet Separate shower Private Mixing valve Kitchen sink Private Faucet Laundry trays (1 to 3) Private Faucet Combination fixture Private Faucet
Weight in Fixture Unitsc 10 5 10 5 3 2 4 4 3 4 6 3 1 2 2 8 6 2 2 3 3
Fig. 9 Demand Versus Fixture Units, Mixed System, High Part of Curve (Adapted from Hunter 1941)
Source: Hunter (1941). a For supply outlets likely to impose continuous demands, estimate continuous supply separately, and add to total demand for fixtures. b For fixtures not listed, weights may be assumed by comparing fixture to listed one using water in similar quantities and at similar rates. c Given weights are for total demand. For fixtures with both hot- and cold-water supplies, weights for maximum separate demands can be assumed to be 75% of listed demand for the supply.
continuous demand during heavy use of the weighted fixtures should be estimated separately and added to demand for fixtures used intermittently to estimate total demand. The Hunter curves in Figures 9, 10, and 11 are based on use patterns in residential buildings and can be erroneous for other usages such as sports arenas. Williams (1976) discusses the Hunter assumptions and presents an analysis using alternative assumptions. So far, the information presented shows the design rate of flow to be determined in any particular section of piping. The next step is to determine the size of piping. As water flows through a pipe, the pressure continually decreases along the pipe because of loss of energy from friction. The problem is then to ascertain the minimum pressure in the street main and the minimum pressure required to operate the topmost fixture. (A pressure of 100 kPa may be ample for most flush valves, but manufacturers’ requirements should be consulted. Some fixtures require a pressure up to 175 kPa. A minimum of 55 kPa should be allowed for other fixtures.) The pressure differential
Fig. 10
Estimate Curves for Demand Load (Adapted from Hunter 1941)
overcomes pressure losses in the distributing system and the difference in elevation between the water main and the highest fixture. The pressure loss (in kPa) resulting from the difference in elevation between the street main and the highest fixture can be obtained by multiplying the difference in elevation in metres by the conversion factor 9.8. Pressure losses in the distributing system consist of pressure losses in the piping itself, plus the pressure losses in the pipe fittings, valves, and the water meter, if any. Approximate design pressure losses and flow limits for disk-type meters for various rates of flow are given in Figure 12. Water authorities in many localities
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Fig. 11 Section of Figure 10 on Enlarged Scale
Fig. 13 Variation of Pressure Loss with Flow Rate for Various Faucets and Cocks
Fig. 12 Pressure Losses in Disk-Type Water Meters require compound meters for greater accuracy with varying flow; consult the local utility. Design data for compound meters differ from the data in Figure 12. Manufacturers give data on exact pressure losses and capacities. Figure 13 shows the variation of pressure loss with rate of flow for various faucets and cocks. The water demand for hose bibbs or other large-demand fixtures taken off the building main frequently results in inadequate water supply to the upper floor of a building. This condition can be prevented by sizing the distribution system so that pressure drops from the street main to all fixtures are the same. An ample building main (not less than 25 mm where possible) should be maintained until all branches to hose bibbs have been connected. Where street main pressure is excessive and a pressure-reducing valve is used to prevent water hammer or excessive pressure at fixtures, hose bibbs should be connected ahead of the reducing valve. The principles involved in sizing upfeed and downfeed systems are the same. In the downfeed system, however, the difference in elevation between the overhead supply mains and the fixtures provides the pressure required to overcome pipe friction. Because friction pressure loss and height pressure loss are not additive, as in an upfeed system, smaller pipes may be used with a downfeed system.
Plastic Pipe The maximum safe water velocity in a thermoplastic piping system under most operating conditions is typically 1.5 m/s; however, higher velocities can be used in cases where the operating characteristics of valves and pumps are known so that sudden changes in flow velocity can be controlled. The total pressure in the system at any time (operating pressure plus surge of water hammer) should not exceed 150% of the pressure rating of the system.
Procedure for Sizing Cold-Water Systems The recommended procedure for sizing piping systems is as follows: 1. Sketch the main lines, risers, and branches, and indicate the fixtures to be served. Indicate the rate of flow of each fixture. 2. Using Table 25, compute the demand weights of the fixtures in fixture units. 3. Determine the total demand in fixture units and, using Figure 9, 10, or 11, find the expected demand. 4. Determine the equivalent length of pipe in the main lines, risers, and branches. Because the sizes of the pipes are not known, the exact equivalent length of various fittings cannot be determined. Add the equivalent lengths, starting at the street main and proceeding along the service line, main line of the building, and up the riser to the top fixture of the group served. 5. Determine the average minimum pressure in the street main and the minimum pressure required for operation of the topmost fixture, which should be 50 to 175 kPa above atmospheric. 6. Calculate the approximate design value of the average pressure drop per unit length of pipe in equivalent length determined in step 4 and using Equation (1). p = ( ps – 9.8H – pf – pm)/L where p ps pf pm H L
= = = = = =
average pressure loss per metre of equivalent length of pipe, kPa pressure in street main, kPa minimum pressure required to operate topmost fixture, kPa pressure drop through water meter, kPa height of highest fixture above street main, m equivalent length determined in step 4, m
If the system is downfeed supply from a gravity tank, height of water in the tank, converted to kPa by multiplying by 9.8, replaces the street main pressure, and the term 9.8H is added
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22.26
2017 ASHRAE Handbook—Fundamentals (SI)
instead of subtracted in calculating p. In this case, H is the vertical distance of the fixture below the bottom of the tank. The pressure conversion factor 9.8 is determined by the mass of water occupying a 1 m3 volume, or 9800 N/m2 (9.8 kPa/m). 7. From the expected rate of flow determined in step 3 and the value of p calculated in step 6, choose the sizes of pipe from Figure 14, 15, or 16. Example 4. Assume a minimum street main pressure of 375 kPa; a height of topmost fixture (a urinal with flush valve) above street main of 15 m; an equivalent pipe length from water main to highest fixture of 30 m; a total load on the system of 50 fixture units; and that the water closets are flush valve operated. Find the required size of supply main. Solution: Use Equation (1): p = ( ps – 9.8H – pf – pm)/L ps = Street main pressure (given) = 375 kPa H = 15 m (given) Pf = 105 kPa from Table 24 Flow = 3.2 L/s from Figure 11
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For a trial run, use 40 mm; then Pm = 45 kPa from Figure 12 at 3.2 L/s. The pressure drop available for overcoming friction in pipes and fittings is 375 – 9.8 15 – 105 – 45 = 78 kPa. At this point, estimate the equivalent pipe length of the fittings on the direct line from the street main to the highest fixture. The exact equivalent length of the various fittings cannot be determined because the pipe sizes of the building main, riser, and branch leading to the highest fixture are not yet known, but a first approximation is necessary to tentatively select pipe sizes. If the computed pipe sizes differ from those used in determining the equivalent length of pipe fittings, a recalculation using the computed pipe sizes for the fittings will be necessary. It is common practice for the first trial to assume that the total equivalent length of the pipe fittings is 50% of the total length of pipe. In this example, 30 m 50% = 15 m. The permissible pressure loss per metre of equivalent pipe is 78/(30 + 15) = 1.7 kPa/m. A 40 mm building main is adequate. The sizing of the branches of the building main, the risers, and the fixture branches follows these principles. For example, assume that one of the branches of the building main carries the cold-water supply for three water closets, two bathtubs, and three lavatories. Using the permissible pressure loss of 1.7 kPa/m, the size of branch (determined from Table 25 and Figures 14 and 11) is found to be 40 mm. Items included in the computation of pipe size are as follows:
Table 26 is a guide to minimum pipe sizing where flush valves are used.
Fig. 14
Fixtures, No. and Type
Fixture Units (Table 25 and Note c)
3 flush valves 2 bathtubs 3 lavatories Total
36 = 18 0.75 2 2 = 3 0.75 3 1 = 2.25 = 23.25
Demand Pipe Size (Figure 11) (Figure 14)
2.4 L/s
40 mm
Velocities exceeding 3 m/s cause undesirable noise in the piping system. This usually governs the size of larger pipes in the system, whereas in small pipe sizes, the friction loss usually governs the selection because the velocity is low compared to friction loss. Velocity is the governing factor in downfeed systems, where friction loss is usually neglected. Velocity in branches leading to pump suctions should not exceed 1.5 m/s. If the street pressure is too low to adequately supply upper-floor fixtures, the pressure must be increased. Constant- or variable-speed booster pumps, alone or in conjunction with gravity supply tanks, or hydropneumatic systems may be used. Flow control valves for individual fixtures under varying pressure conditions automatically adjust flow at the fixture to a predetermined quantity. These valves allow the designer to (1) limit flow at the individual outlet to the minimum suitable for the purpose, (2) hold total demand for the system more closely to the required minimum, and (3) design the piping system as accurately as is practicable for the requirements.
Hydronic System Piping The Darcy-Weisbach equation with friction factors from the Moody chart or Colebrook equation (or, alternatively, the HazenWilliams equation) is fundamental to calculating pressure drop in hot- and chilled-water piping; however, charts calculated from these equations (such as Figures 14, 15, and 16) provide easy determination of pressure drops for specific fluids and pipe standards. In addition, tables of pressure drops can be found in Crane Co. (1976) and Hydraulic Institute (1990). The Reynolds numbers represented on the charts in Figures 14, 15, and 16 are all in the turbulent flow regime. For smaller pipes and/ or lower velocities, the Reynolds number may fall into the laminar regime, in which the Colebrook friction factors are no longer valid. Most tables and charts for water are calculated for properties at 15°C. Using these for hot water introduces some error, although the answers are conservative (i.e., cold-water calculations overstate the pressure drop for hot water). Using 15°C water charts for 90°C water should not result in errors in p exceeding 20%.
Friction Loss for Water in Commercial Steel Pipe (Schedule 40)
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Pipe Design
22.27
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Fig. 15
Friction Loss for Water in Copper Tubing (Types K, L, M)
Fig. 16 Friction Loss for Water in Plastic Pipe (Schedule 80) Table 26 Allowable Number of 25 mm Flush Valves Served by Various Sizes of Water Pipe* Pipe Size, mm 32 40 50 65 75 100
No. of 25 mm Flush Valves 1 2 to 4 5 to 12 13 to 25 26 to 40 41 to 100
*Two 20 mm flush valves are assumed equal to one 25 mm flush valve but can be served by a 25 mm pipe. Water pipe sizing must consider demand factor, available pressure, and length of run.
Range of Usage of Pressure Drop Charts General Design Range. The general range of pipe friction loss used for design of hydronic systems is between 100 and 400 Pa per metre of pipe. A value of 250 Pa/m represents the mean to which most systems are designed. Wider ranges may be used in specific designs if certain precautions are taken. Piping Noise. Closed-loop hydronic system piping is generally sized below certain arbitrary upper limits, such as a velocity limit of 1.2 m/s for 50 mm pipe and under, and a pressure drop limit of
400 Pa/m for piping over 50 mm in diameter. Velocities in excess of 1.2 m/s can be used in piping of larger size. This limitation is generally accepted, although it is based on relatively inconclusive experience with noise in piping. Water velocity noise is not caused by water but by free air, sharp pressure drops, turbulence, or a combination of these, that cause cavitation or flashing of water into steam. Therefore, higher velocities may be used if proper precautions are taken to eliminate air and turbulence.
Air Separation Air in hydronic systems is usually undesirable because it causes flow noise, allows oxygen to react with piping materials, and sometimes even prevents flow in parts of a system. Air may enter a system at an air/water interface in an open system or in an expansion tank in a closed system, or it may be brought in dissolved in makeup water. Most hydronic systems use air separation devices to remove air. The solubility of air in water increases with pressure and decreases with temperature; thus, separation of air from water is best achieved at the point of lowest pressure and/or highest temperature in a system. For more information, see Chapter 13 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment.
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22.28
2017 ASHRAE Handbook—Fundamentals (SI) Table 27 Equivalent Length in Metres of Pipe for 90° Elbows Pipe Size, mm
Velocity, m/s
15
20
25
32
40
50
65
90
100
125
150
200
250
300
0.33 0.67 1.00 1.33 1.67
0.4 0.4 0.5 0.5 0.5
0.5 0.6 0.6 0.6 0.7
0.7 0.8 0.8 0.8 0.9
0.9 1.0 1.1 1.1 1.2
1.1 1.2 1.3 1.3 1.4
1.4 1.5 1.6 1.7 1.8
1.6 1.8 1.9 2.0 2.1
2.0 2.3 2.5 2.5 2.6
2.6 2.9 3.1 3.2 3.4
3.2 3.6 3.8 4.0 4.1
3.7 4.2 4.5 4.6 4.8
4.7 5.3 5.6 5.8 6.0
5.7 6.3 6.8 7.1 7.4
6.8 7.6 8.0 8.4 8.8
2.00 2.35 2.67 3.00 3.33
0.5 0.5 0.5 0.5 0.5
0.7 0.7 0.7 0.7 0.8
0.9 0.9 0.9 0.9 0.9
1.2 1.2 1.3 1.3 1.3
1.4 1.5 1.5 1.5 1.5
1.8 1.9 1.9 1.9 1.9
2.2 2.2 2.3 2.3 2.4
2.7 2.8 2.8 2.9 3.0
3.5 3.6 3.6 3.7 3.8
4.3 4.4 4.5 4.5 4.6
5.0 5.1 5.2 5.3 5.4
6.2 6.4 6.5 6.7 6.8
7.6 7.8 8.0 8.1 8.2
9.0 9.2 9.4 9.6 9.8
Table 28
Iron and Copper Elbow Equivalents*
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Fitting
Iron Pipe
Copper Tubing
Elbow, 90° 45° 90° long-radius 90° welded
1.0 0.7 0.5 0.5
1.0 0.7 0.5 0.5
Reduced coupling Open return bend Angle radiator valve Radiator or convector
0.4 1.0 2.0 3.0
0.4 1.0 3.0 4.0
Boiler or heater Open gate valve Open globe valve
0.5 12.0
3.0
0.7 17.0
4.0
Sources: Giesecke (1926) and Giesecke and Badgett (1931, 1932a). *See Table 10 for equivalent length of one elbow.
In the absence of venting, air can be entrained in the water and carried to separation units at flow velocities of 0.5 to 0.6 m/s or more in pipe 50 mm and under. Minimum velocities of 0.6 m/s are therefore recommended. For pipe sizes 50 mm and over, minimum velocities corresponding to a pressure loss of 75 Pa are normally used. Maintaining minimum velocities is particularly important in the upper floors of high-rise buildings where the air tends to come out of solution because of reduced pressures. Higher velocities should be used in downcomer return mains feeding into air separation units located in the basement. Example 5. Determine the iron pipe size for a circuit requiring 1.25 L/s flow. Solution: Enter Figure 4 at 1.25 L/s, read up to pipe size within normal design range (100 to 400 Pa/m), and select 40 mm. Velocity is 1 m/s and pressure loss is 300 Pa/m.
Valve and Fitting Pressure Drop Valves and fittings can be listed in elbow equivalents, with an elbow being equivalent to a length of straight pipe. Table 27 lists equivalent lengths of 90° elbows; Table 28 lists elbow equivalents for valves and fittings for iron and copper. Example 6. Determine equivalent length of pipe for a 100 mm open gate valve at a flow velocity of approximately 1.33 m/s. Solution: From Table 27, at 1.33 m/s, each elbow is equivalent to 3.2 m of 100 mm pipe. From Table 28, the gate valve is equivalent to 0.5 elbows. The actual equivalent pipe length (added to measured circuit length for pressure drop determination) will be 3.2 0.5, or 1.6 m of 100 mm.
Tee Fitting Pressure Drop. Pressure drop through pipe tees varies with flow through the branch. Figure 17 shows pressure drops for nominal 25 mm tees of equal inlet and outlet sizes and for the flow patterns shown. Idelchik (1986) also presents data for threaded tees. Different investigators present tee loss data in different forms,
Fig. 17 Elbow Equivalents of Tees at Various Flow Conditions (Giesecke and Badgett 1931, 1932b)
sources. As an estimate of the upper limit to tee losses, a pressure or head loss coefficient of 1.0 may be assumed for entering and leaving 2 /2). flows (i.e., p = 1.0Vin2 /2 + 1.0V out Example 7. Determine the pressure or head losses for a 25 mm (all openings) threaded pipe tee flowing 25% to the side branch, 75% through. The entering flow is 1 L/s (1.79 m/s). Solution: From Figure 17, bottom curve, the number of equivalent elbows for the through-flow is 0.15 elbows; the through-flow is 0.75 L/s (1.34 m/s); and the pressure drop is based on the exit flow rate. Table 27 gives the equivalent length of a 25 mm elbow at 1.33 m/s as 0.8 m. Using Figure 14, the head loss is 900 Pa/m for 25 mm pipe and 0.75 L/s flow.
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Pipe Design
22.29 ity, the quieter the system. When condensate must flow against the steam, even in limited quantity, the steam’s velocity must not exceed limits above which the disturbance between the steam and the counterflowing water may (1) produce objectionable sound, such as water hammer, or (2) result in the retention of water in certain parts of the system until the steam flow is reduced sufficiently to allow water to pass. These limits are a function of (1) pipe size; (2) pitch of the pipe if it runs horizontally; (3) quantity of condensate flowing against the steam; and (4) freedom of the piping from water pockets that, under certain conditions, act as a restriction in pipe size. Table 30 lists maximum capacities for various size steam lines. Equivalent Length of Run. All tables for the flow of steam in pipes based on pressure drop must allow for pipe friction, as well as for the resistance of fittings and valves. These resistances are generally stated in terms of straight pipe; that is, a certain fitting produces a drop in pressure equivalent to the stated length of straight run of the same size of pipe. Table 31 gives the length of straight pipe usually allowed for the more common types of fittings and valves. In all pipe sizing tables in this chapter, length of run refers to the equivalent length of run as distinguished from the actual length of pipe. A common sizing method is to assume the length of run and to check this assumption after pipes are sized. For this purpose, length of run is usually assumed to be double the actual length of pipe.
= (0.15)(0.8)(900) = 108 Pa = 0.108 kPa pressure drop h = (0.15)(0.8)(900)/(1000)(9.8) = 0.0110 m head loss From Figure 17, top curve, the number of equivalent elbows for the branch flow of 25% is 13 elbows; the branch flow is 0.75 L/s (0.45 m/s); and the head loss or pressure drop is based on the exit flow rate. Table 27 gives the equivalent of a 25 mm elbow at 0.45 m/s as 0.75 m. Using Figure 14, the pressure drop is 130 Pa/m for 25 mm pipe and 0.25 L/s flow. p = (13)(0.75)(130) = 1268 Pa = 1.268 kPa pressure drop h = (13)(0.75)(130)/(1000)(9.8) = 0.129 m head loss p
3.3
STEAM PIPING
Pressure losses in steam piping for flows of dry or nearly dry steam are governed by Equations (2) to (8) in the section on Design Equations. This section incorporates these principles with other information specific to steam systems.
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Pipe Sizes Required pipe sizes for a given load in steam heating depend on the following factors: • The initial pressure and the total pressure drop that can be allowed between the source of supply and the end of the return system • The maximum velocity of steam allowable for quiet and dependable operation of the system, taking into consideration the direction of condensate flow • The equivalent length of the run from the boiler or source of steam supply to the farthest heating unit Initial Pressure and Pressure Drop. Table 29 lists pressure drops commonly used with corresponding initial steam pressures for sizing steam piping. Several factors, such as initial pressure and pressure required at the end of the line, should be considered, but it is most important that (1) the total pressure drop does not exceed the initial gage pressure of the system (in practice, it should never exceed one-half the initial gage pressure); (2) pressure drop is not great enough to cause excessive velocities; (3) a constant initial pressure is maintained, except on systems specially designed for varying initial pressures (e.g., subatmospheric pressure), that normally operate under controlled partial vacuums; and (4) for gravity return systems, pressure drop to heating units does not exceed the water column available for removing condensate (i.e., height above the boiler water line of the lowest point on the steam main, on the heating units, or on the dry return). Maximum Velocity. For quiet operation, steam velocity should be 40 to 60 m/s, with a maximum of 75 m/s. The lower the velocTable 30
Example 8. Using Table 31, determine the equivalent length of pipe for the run shown.
Measured length = 40 m 100 mm gate valve = 0.6 m Four 100 mm elbows = 10.8 m
Table 29 Pressure Drops Used for Sizing Steam Pipea Initial Steam Pressure, kPab
Pressure Drop, Pa/m
Total Pressure Drop in Steam Supply Piping, kPa
Vacuum return 101 108 115 135 170 205 310 445 790 1140
30 to 60 7 30 30 60 115 225 450 450 to 1100 450 to 1100 450 to 2300
7 to 14 0.4 0.4 to 1.7 3.5 10 20 30 35 to 70 70 to 105 105 to 170 170 to 210
a Equipment, control valves, and so forth must be selected based on delivered pressures.
b Subtract
101 to convert to pressure above atmospheric.
Comparative Capacity of Steam Lines at Various Pitches for Steam and Condensate Flowing in Opposite Directions Nominal Pipe Diameter, mm
Pitch of Pipe, mm/m 20 40 80 120 170 250 350 420
20 Capacity 0.4 0.5 0.7 0.8 0.9 1.0 1.2 1.3
Source: Laschober et al. (1966).
25
Maximum Velocity 2.4 3.4 4.0 4.3 4.9 5.2 6.7 6.7
Capacity 0.9 1.1 1.5 1.6 1.9 2.2 2.4 2.6
32
Maximum Velocity 2.7 3.7 4.6 5.2 5.8 6.7 7.3 7.6
Capacity 1.5 2 2.5 3.1 3.4 3.9 4.2 4.9
40
Maximum Velocity 3.4 4.3 5.2 6.1 6.7 7.6 7.9 9.4
Capacity 2.5 3.3 4.2 4.7 5.3 5.9 6.4 7.5
Capacity in g/s; velocity in m/s.
50
Maximum Velocity 3.7 4.9 5.8 6.7 7.3 7.9 8.5 10.1
Capacity 5.4 6.8 8.7 10.5 11.7 12.5 12.9 14.5
Maximum Velocity 4.6 5.5 7.3 8.2 9.1 9.8 9.8 10.1
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22.30
2017 ASHRAE Handbook—Fundamentals (SI) Two 100 mm tees Equivalent
On completion of the sizing, the drop could be checked by taking the longest line and actually calculating the equivalent length of run from the pipe sizes determined. If the calculated drop is less than that assumed, the pipe size is adequate; if it is more, an unusual number of fittings is probably involved, and either the lines must be straightened, or the next larger pipe size must be tried.
= 11 m = 62.4m
Sizing Charts Figure 18 is the basic chart for determining the flow rate and velocity of steam in Schedule 40 pipe for various values of pressure drop per unit length, based on saturated steam at standard pressure (101.325 kPa). Using the multiplier chart (Figure 19), Figure 18 can be used at all saturation pressures between 101 and 1500 kPa (see Example 10).
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3.4
High-Pressure Steam Piping Many heating systems for large industrial buildings use highpressure steam [100 to 1000 kPa (gage)]. These systems usually have unit heaters or large built-up fan units with blast heating coils.
LOW-PRESSURE STEAM PIPING
Values in Table 32 (taken from Figure 18) provide a more rapid means of selecting pipe sizes for the various pressure drops listed and for systems operated at 25 and 85 kPa (gage). The flow rates shown for 25 kPa can be used for saturated pressures from 7 to 41 kPa, and those shown for 85 kPa can be used for saturated pressures from 55 to 110 kPa with an error not exceeding 8%. Both Figure 18 and Table 32 can be used where the flow of condensate does not inhibit the flow of steam. Columns B and C of Table 33 are used in cases where steam and condensate flow in opposite directions, as in risers or runouts that are not dripped. Columns D, E, and F are for one-pipe systems and include risers, radiator valves and vertical connections, and radiator and riser runout sizes, all of which are based on the critical velocity of the steam to allow counterflow of condensate without noise. Return piping can be sized by Table 34, using pipe capacities for wet, dry, and vacuum return lines for several values of pressure drop per metres of equivalent length.
Table 31 Nominal Pipe Diameter, mm
Example 9. What pressure drop should be used for the steam piping of a system if the measured length of the longest run is 150 m, and the initial pressure must not exceed 14 kPa above atmospheric? Solution: It is assumed, if the measured length of the longest run is 150 m, that when the allowance for fittings is added, the equivalent length of run does not exceed 300 m. Then, with the pressure drop not over one-half of the initial pressure, the drop could be 7 kPa or less. With a pressure drop of 7 kPa and a length of run of 300 m, the drop would be 23 Pa/m; if the total drop were 3.5 kPa, the drop would be 12 Pa/m. In both cases, the pipe could be sized for a desired capacity according to Figure 18.
Table 32 Nominal 14 Pa/m Pipe Size, Sat. Press., kPa mm 25 85
Equivalent Length of Fittings to Be Added to Pipe Run Length to Be Added to Run, m Gate Valvea
Standard Side Elbow Outlet Teeb
Globe Valvea
Angle Valvea
15 20 25 32
0.4 0.5 0.7 0.9
0.9 1.2 1.5 1.8
0.1 0.1 0.1 0.2
4 5 7 9
2 3 4 5
40 50 65 80
1.1 1.3 1.5 1.9
2.1 2.4 3.4 4.0
0.2 0.3 0.3 0.4
10 14 16 20
6 7 8 10
100 125 150
2.7 3.3 4.0
5.5 6.7 8.2
0.6 0.7 0.9
28 34 41
14 17 20
200 250 300 350
5.2 6.4 8.2 9.1
11 14 16 19
1.1 1.4 1.7 1.9
55 70 82 94
28 34 40 46
a Valve
in full-open position. apply only to a tee used to divert the flow in the main to the last riser.
b Values
Flow Rate of Steam in Schedule 40 Pipe Pressure Drop, Pa/m
28 Pa/m
58 Pa/m
113 Pa/m
170 Pa/m
225 Pa/m
450 Pa/m
Sat. Press., kPa 25 85
Sat. Press., kPa 25 85
Sat. Press., kPa 25 85
Sat. Press., kPa 25 85
Sat. Press., kPa 25 85
Sat. Press., kPa 25 85
20 25 32 40
1.1 2.1 4.5 7.1
1.4 2.6 5.7 8.8
1.8 3.3 6.7 11
2.0 3.9 8.3 13
2.5 4.7 9.8 15
3.0 5.8 12 19
50 65 80 90
14 22 40 58
17 27 48 69
20 33 59 84
24 39 69 101
29 48 83 125
36 58 102 153
42 68 121 178
52 83 146 214
53 86 150 219
64 103 180 265
60 98 174 252
74 120 210 305
89 145 246 372
107 173 302 435
100 125 150 200
81 151 242 491
101 180 290 605
120 212 355 702
146 265 422 882
178 307 499 1 020
213 378 611 1 260
249 450 718 1 440
302 536 857 1 800
309 552 882 1 830
378 662 1 080 2 230
363 643 1 060 2 080
436 769 1 260 2 580
529 945 1 500 3 020
617 1 080 1 790 3 720
250 300
907 1 440
1 110 1 730
1 290 2 080
1 590 2 460
1 890 2 950
2 290 3 580
2 650 4 160
3 280 5 040
3 300 5 170
4 030 6 240
3 780 6 050
4 660 7 250
5 380 8 540
6 550 10 200
Notes: 1. Flow rate is in g/s at initial saturation pressures of 25 and 85 kPa (gage). Flow is based on Moody friction factor, where the flow of condensate does not inhibit the flow of steam.
3.7 6.8 14 22
4.4 8.3 17 26
4.5 8.6 18 27
5.4 10 21 33
5.3 10 20 31
6.3 12 25 38
7.6 14 29 45
9.2 17 35 54
2. The flow rates at 25 kPa cover saturated pressure from 7 to 41 kPa, and the rates at 85 kPa cover saturated pressure from 55 to 110 kPa with an error not exceeding 8%. 3. The steam velocities corresponding to the flow rates given in this table can be found from Figures 18 and 19.
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22.31
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Pipe Design
Notes: Based on Moody Friction Factor where flow of condensate does not inhibit flow of steam. See Figure 19 for obtaining flow rates and velocities of all saturation pressures between 101 to 1500 kPa; see also Examples 9 and 10.
Fig. 18
Flow Rate and Velocity of Steam in Schedule 40 Pipe at Saturation Pressure of 101 kPa
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22.32
2017 ASHRAE Handbook—Fundamentals (SI) Temperatures are controlled by a modulating or throttling thermostatic valve or by face or bypass dampers controlled by the room air temperature, fan inlet, or fan outlet.
Table 33 Steam Pipe Capacities for Low-Pressure Systems Capacity, g/s Two-Pipe System
One-Pipe Systems
Use of Basic and Velocity Multiplier Charts
Condensate Flowing Against Steam
Nominal Pipe Size, mm Vertical Horizontal A
Ba
Cb
Dc
E
Fb
1.0 1.8 3.9 6.0 12
0.9 1.8 3.4 5.3 11
0.8 1.4 2.5 4.8 9.1
— 0.9 2.0 2.9 5.3
0.9 0.9 2.0 2.0 2.9
65 80 90 100 125
20 36 49 64 132
17 25 36 54 99
14 25 36 48 —
— — — — —
5.3 8.2 15 23 35
150 200 250 300 400
227 472 882 1450 2770
176 378 718 1200 2390
— — — — —
— — — — —
69 — — — —
Example 10. Given a flow rate of 0.85 kg/s, an initial steam pressure of 800 kPa, and a pressure drop of 2.5 kPa/m, find the size of Schedule 40 pipe required and the velocity of steam in the pipe. Solution: The following steps are shown by the broken line on Figures 18 and 19. 1. Enter Figure 18 at a flow rate of 0.85 kg/s, and move vertically to the horizontal line at 800 kPa 2. Follow along inclined multiplier line (upward and to the left) to horizontal 101 kPa line. The equivalent mass flow at 101 kPa is about 0.30 kg/s. 3. Follow the 0.30 kg/s line vertically until it intersects the horizontal line at 2500 Pa/m pressure drop. Nominal pipe size is 65 mm. The equivalent steam velocity at 101 kPa is about 165 m/s. 4. To find the steam velocity at 800 kPa, locate the value of 165 m/s on the ordinate of the velocity multiplier chart (Figure 19) at 101 kPa. 5. Move along the inclined multiplier line (downward and to the right) until it intersects the vertical 800 kPa pressure line. The velocity is about 65 m/s. Note: Steps 1 through 5 would be rearranged or reversed if different data were given.
3.5
Notes: 1. For one- or two-pipe systems in which condensate flows against steam flow. 2. Steam at average pressure of 7 kPa (gage) used as basis of calculating capacities.
STEAM CONDENSATE SYSTEMS
The majority of steam systems used in heating applications are two-pipe systems (steam pipe and condensate pipe). This discussion is limited to sizing the condensate lines in two-pipe systems.
a Do not use column B for pressure drops of less than 13 Pa per metre of equivalent run.
Use Figure 18 or Table 31 instead. b Pitch of horizontal runouts to risers and radiators should be not less than 40 mm/m. Where this pitch cannot be obtained, runouts over 2.5 m in length should be one pipe size larger than that called for in this table. c Do not use column D for pressure drops of less than 9 Pa per metre of equivalent run, except on sizes 80 mm and over. Use Figure 18 or Table 31 instead.
Two-Pipe Systems When steam is used for heating a liquid to 102°C or less (e.g., in domestic water heat exchangers, domestic heating water converters,
Return Main
Table 34 Return Main and Riser Capacities for Low-Pressure Systems, g/s
Riser
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20 25 32 40 50
Supply Risers Upfeed
Radiator Valves and Radiator Vertical and Riser Connections Runouts
Pipe Size, mm
Wet
Wet
Dry
Vac.
Wet
Dry
Vac.
Wet
Dry
Vac.
Wet
Dry
Vac.
G
H
I
J
K
L
M
N
O
P
Q
R
S
T
U
V
20 25 32 40 50
— 16 27 43 88
— 8 16 26 59
— — — — —
— 18 31 50 102
— 9 19 30 67
5 18 31 49 103
— 22 38 60 126
— 10 21 33 72
13 22 38 60 126
— 32 54 85 176
— 13 27 43 93
18 31 54 85 179
— 44 76 120 252
— 14 30 48 104
25 44 76 120 252
65 80 90 100 125 150
149 237 347 489 — —
96 184 248 369 — —
— — — — — —
199 268 416 577 — —
109 197 277 422 — —
171 275 410 567 993 1590
212 338 504 693 — —
120 221 315 473 — —
212 338 504 693 1220 1950
296 473 693 977 — —
155 284 407 609 — —
300 479 716 984 1730 2770
422 674 1010 1390 — —
171 315 451 678 — —
20 25 32 40 50
— — — — —
6 14 31 47 95
— — — — —
— — — — —
6 14 31 47 95
18 31 49 103 171
— — — — —
6 14 31 47 95
22 38 60 126 212
— — — — —
6 14 31 47 95
31 54 85 179 300
— — — — —
65 80 90 100 125
— — — — —
— — — — —
— — — — —
— — — — —
— — — — —
275 410 564 993 1590
— — — — —
— — — — —
338 504 693 1220 1950
— — — — —
— — — — —
479 716 984 1730 2772
— — — — —
7 Pa/m
9 Pa/m
Dry Vac.
14 Pa/m
28 Pa/m
57 Pa/m
113 Pa/m Wet Dry W
Vac.
X
Y
— — — — —
— — — — —
36 62 107 169 357
422 674 1010 1390 2440 3910
— — — — — —
— — — — — —
596 953 1424 1953 3440 5519
6 14 31 47 95
44 76 120 252 422
— — — — —
— — — — —
62 107 169 357 596
— — — — —
674 1010 1390 2440 3910
— — — — —
— — — — —
953 1424 1953 3440 5519
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Pipe Design
Fig. 19
22.33
Velocity Multiplier Chart for Figure 18
or air-heating coils), the devices are usually provided with a steam control valve. As the control valve throttles, the absolute pressure in the load device decreases, removing all pressure motivation for flow in the condensate return system. To ensure the flow of steam condensate from the load device through the trap and into the return system, it is necessary to provide a vacuum breaker on the device ahead of the trap. This ensures a minimum pressure at the trap inlet of atmospheric pressure plus whatever liquid leg the designer has provided. Then, to ensure flow through the trap, it is necessary to design the condensate system so that it will never have a pressure above atmospheric in the condensate return line. Vented (Open) Return Systems. To achieve this pressure requirement, the condensate return line is usually vented to the atmosphere (1) near the point of entrance of the flow streams from the load traps, (2) in proximity to all connections from drip traps, and (3) at transfer pumps or feedwater receivers. The dry return lines in a vented return system have flowing liquid in the bottom of the line and gas or vapor in the top (Figure 20A). The liquid is the condensate, and the gas may be steam, air, or a mixture of the two. The flow phenomenon for these dry return systems is open channel flow, which is best described by the Manning equation: 23
12
1.00Ar S Q = -----------------------------------n
(22)
Fig. 20
Types of Condensate Return Systems
where Q A r n S
= = = = =
volumetric flow rate, m3/s cross-sectional area of conduit, m2 hydraulic radius of conduit, m coefficient of roughness (usually 0.012) slope of conduit, m/m
Table 35 is a solution to Equation (22) that shows pipe size capacities for steel pipes with various pitches. Recommended practice is to size vertical lines by the maximum pitch shown, although they would actually have a capacity far in excess of that shown. As pitch increases, hydraulic jump that could fill the pipe and other transient effects that could cause water hammer should be avoided. Flow values in Table 35 are calculated for Schedule 40 steel pipe, with a factor of safety of 3.0, and can be used for copper pipes of the same nominal pipe size. The flow characteristics of wet return lines (Figure 20B) are best described by the Darcy-Weisbach equation [Equation (1)]. The motivation for flow is the fluid pressure difference between the entering section of the flooded line and the leaving section. It is common practice, in addition to providing for the fluid pressure
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22.34
2017 ASHRAE Handbook—Fundamentals (SI) described for closed returns, and in accordance with Table 34 or Table 37, as applicable. Passage of fluid through the steam trap is a throttling or constantenthalpy process. The resulting fluid on the downstream side of the trap can be a mixture of saturated liquid and vapor. Thus, in nonvented returns, it is important to understand the fluid’s condition when it enters the return line from the trap. The condition of the condensate downstream of the trap can be expressed by the quality x, defined as
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differential, to slope the return in the direction of flow to a collection point such as a dirt leg to clear the line of sediment or solids. Table 36 is a solution to Equation (1) that shows pipe size capacity for steel pipes with various available fluid pressures. Table 36 can also be used for copper tubing of equal nominal pipe size. Nonvented (Closed) Return Systems. For systems with a continual steam pressure difference between the point where the condensate enters the line and the point where it leaves (Figure 20C), Table 34 or Table 35, as applicable, can be used for sizing the condensate lines. Although these tables express condensate capacity without slope, common practice is to slope the lines in the direction of flow to a collection point (similar to wet returns) to clear the lines of sediment or solids. When saturated condensate at pressures above the return system pressure enters the return (condensate) mains, some of the liquid flashes to steam. This occurs typically at drip traps into a vented return system or at load traps leaving process load devices that are not valve controlled and typically have no subcooling. If the return main is vented, the vent lines relieve any excessive pressure and prevent a backpressure phenomenon that could restrict flow through traps from valved loads; the pipe sizing would be as described for vented dry returns. If the return line is not vented, flash steam causes a pressure rise at that point and the piping could be sized as
mv x = -----------------ml + mv where mv = mass of saturated vapor in condensate ml = mass of saturated liquid in condensate
Likewise, the volume fraction Vc of the vapor in the condensate is expressed as Vv Vc = ----------------Vl + Vv
Condensate Flow,
Vv = volume of saturated vapor in condensate Vl = volume of saturated liquid in condensate
15
The quality and the volume fraction of the condensate downstream of the trap can also be estimated from Equations (25) and (26), respectively.
g/sa,b
Condensate Line Slope 0.5% 5
1%
2%
4%
7
10
13
20
10
14
20
29
25
19
27
39
54
32
40
57
80
113
40
60
85
121
171
50
117
166
235
332
65
189
267
377
534
80
337
476
674
953
100
695
983
1390
1970
125
1270
1800
2540
3590
150
2070
2930
4150
5860
(24)
where
Table 35 Vented Dry Condensate Return for Gravity Flow Based on Manning Equation Nominal Diameter, mm
(23)
h1 – hf 2 x = ------------------hg – hf 2
(25)
2
xvg 2 Vc = --------------------------------------vf 1 – x + xvg 2
(26)
2
where h1 = enthalpy of liquid condensate entering trap evaluated at supply pressure for saturated condensate or at saturation pressure corresponding to temperature of subcooled liquid condensate hf2 = enthalpy of saturated liquid at return or downstream pressure of trap hg2 = enthalpy of saturated vapor at return or downstream pressure of trap vf2 = specific volume of saturated liquid at return or downstream pressure of trap vg2 = specific volume of saturated vapor at return or downstream pressure of trap.
a Flow
is in g/s of 82°C water for Schedule 40 steel pipes. b Flow was calculated from Equation (22) and rounded.
Table 36 Vented Wet Condensate Return for Gravity Flow Based on Darcy-Weisbach Equation Condensate Flow, g/sa,b Nominal Diameter, mm 15 20 25 32 40 50 65 80 100 125 150 a Flow
Condensate Pressure, Pa/m 50
100
150
200
250
300
350
400
13 28 54 114 172 334 536 954 1 960 3 560 5 770
19 41 79 165 248 482 773 1 370 2 810 5 100 8 270
24 51 98 204 308 597 956 1 700 3 470 6 290 10 200
28 60 114 238 358 694 1 110 1 970 4 030 7 290 11 800
32 68 129 267 402 779 1 250 2 210 4 520 8 180 13 200
35 74 142 294 442 857 1 370 2 430 4 960 8 980 14 500
38 81 154 318 479 928 1 480 2 630 5 370 9 720 15 700
41 87 165 341 513 994 1 590 2 810 5 750 10 400 16 800
is in g/s of 82°C water for Schedule 40 steel pipes.
b Flow
calculated from Equation (1) and rounded.
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Pipe Design Table 38
22.35
Supply Pressure, kPa (gage)
Return Pressure, kPa (gage)
35 103 207 345 690 1030 690 1030
0 0 0 0 0 0 103 103
Table 39
x, Fraction Vapor, Mass Basis
Vc , Fraction Vapor, Volume Basis
0.016 0.040 0.065 0.090 0.133 0.164 0.096 0.128
0.962 0.985 0.991 0.994 0.996 0.997 0.989 0.992
One-Pipe Systems Gravity one-pipe air vent systems in which steam and condensate flow in the same pipe, frequently in opposite directions, are considered obsolete and are no longer being installed. Chapter 33 of the 1993 ASHRAE Handbook—Fundamentals or earlier ASHRAE Handbook volumes include descriptions of and design information for one-pipe systems.
Estimated Return Line Pressures Pressure in Return Line, Pa (gage)
Pressure Drop, Pa/m
200 kPa (gage) Supply
1000 kPa (gage) Supply
30 60 120 180 240 480
3.5 7 14 21 28 —
9 18 35 52 70 138
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Table 38 presents some values for quality and volume fraction for typical supply and return pressures in heating and ventilating systems. Note that the percent of vapor on a mass basis x is small, although the percent of vapor on a volume basis Vc is very large. This indicates that the return pipe cross section is predominantly occupied by vapor. Figure 21 is a working chart to determine the quality of condensate entering the return line from the trap for various combinations of supply and return pressures. If the liquid is subcooled entering the trap, the saturation pressure corresponding to the liquid temperature should be used for the supply or upstream pressure. Typical pressures in the return line are given in Table 39.
Flash Steam from Steam Trap on Pressure Drop
3.6
GAS PIPING
Piping for gas appliances should be of adequate size and installed so that it provides a supply of gas sufficient to meet the maximum demand without undue loss of pressure between the point of supply (the meter) and the appliance. The size of gas pipe required depends on (1) maximum gas consumption to be provided, (2) length of pipe and number of fittings, (3) allowable pressure loss from the outlet of the meter to the appliance, and (4) density of the gas. Insufficient gas flow from excessive pressure losses in gas supply lines can cause inefficient operation of gas-fired appliances and sometimes create hazardous operations. Gas-fired appliances are normally equipped with a data plate giving information on maximum gas flow requirements or input as well as inlet gas pressure requirements. The local gas utility can give the gas pressure available at the utility’s gas meter. Using this information, the required size of gas piping can be calculated for satisfactory operation of the appliance(s). Table 40 gives pipe capacities for gas flow for up to 60 m of pipe based on a gas density of 0.735 kg/m3. Capacities for pressures less than 10 kPa may also be determined by the following equation from NFPA/IAS National Fuel Gas Code (NFPA Standard 54/ANSI Standard Z223.1): Q = 0.0001d 2.623(p/CL)0.541
Fig. 21
Working Chart for Determining Percentage of Flash Steam (Quality) Table 40
Nominal Iron Internal Pipe Size, Diameter, mm mm 8 10 15 20 25 32 40 50 65 80 100
9.25 12.52 15.80 20.93 26.14 35.05 40.89 52.50 62.71 77.93 102.26
(27)
where Q = flow rate at 15°C and 101 kPa, L/s
Maximum Capacity of Gas Pipe in Litres per Second Length of Pipe, m
5 0.19 0.43 0.79 1.65 2.95 6.4 9.6 18.4 29.3 51.9 105.8
10
15
20
25
30
35
40
45
50
55
60
0.13 0.29 0.54 1.13 2.03 4.4 6.6 12.7 20.2 35.7 72.7
0.11 0.24 0.44 0.91 1.63 3.5 5.3 10.2 16.2 28.6 58.4
0.09 0.20 0.37 0.78 1.40 3.0 4.5 8.7 13.9 24.5 50.0
0.08 0.18 0.33 0.69 1.24 2.7 4.0 7.7 12.3 21.7 44.3
0.07 0.16 0.30 0.63 1.12 2.4 3.6 7.0 11.1 19.7 40.1
0.07 0.15 0.28 0.58 1.03 2.2 3.3 6.4 10.2 18.1 36.9
0.06 0.14 0.26 0.54 0.96 2.1 3.1 6.0 9.5 16.8 34.4
0.06 0.13 0.24 0.50 0.90 1.9 2.9 5.6 8.9 15.8 32.2
0.06 0.12 0.23 0.47 0.85 1.8 2.8 5.3 8.4 14.9 30.4
0.05 0.12 0.22 0.45 0.81 1.7 2.6 5.0 8.0 14.2 28.9
0.05 0.11 0.21 0.43 0.77 1.7 2.5 4.8 7.7 13.5 27.6
Note: Capacity is in litres per second at gas pressures of 3.5 kPa (gage) or less and pressure drop of 75 kPa; density = 0.735 kg/m3.
Copyright by American Gas Association and National Fire Protection Association. Used by permission of copyright holders.
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22.36
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 22 Typical Oil Circulating Loop d = inside diameter of pipe, mm p = pressure drop, Pa C = factor for viscosity, density, and temperature = 0.00223(t + 273)s0.8480.152 t = temperature, °C s = ratio of density of gas to density of air at 15°C and 101 kPa = viscosity of gas, Pa·s (12 for natural gas, 8 for propane) L = pipe length, m
Gas service in buildings is generally delivered in the lowpressure range of 1.7 kPa (gage). The maximum pressure drop allowable in piping systems at this pressure is generally 125 Pa but is subject to regulation by local building, plumbing, and gas appliance codes [see also the NFPA/IAS National Fuel Gas Code (NFPA Standard 54/ANSI Standard Z223.1)]. Where large quantities of gas are required or where long lengths of pipe are used (e.g., in industrial buildings), low-pressure limitations result in large pipe sizes. Local codes may allow (and local gas companies may deliver) gas at higher pressures [e.g., 15, 35, or 70 kPa (gage)]. Under these conditions, an allowable pressure drop of 10% of the initial pressure is used, and pipe sizes can be reduced significantly. Gas pressure regulators at the appliance must be specified to accommodate higher inlet pressures. NFPA/IAS (2012) provides information on pipe sizing for various inlet pressures and pressure drops at higher pressures. More complete information on gas piping can be found in the Gas Engineers’ Handbook (1970).
3.7
FUEL OIL PIPING
The pipe used to convey fuel oil to oil-fired appliances must be large enough to maintain low pump suction pressure and, in the case of circulating loop systems, to prevent overpressure at the burner oil pump inlet. Pipe materials must be compatible with the fuel and must be carefully assembled to eliminate all leaks. Leaks in suction lines can cause pumping problems that result in unreliable burner operation. Leaks in pressurized lines create fire hazards. Cast-iron or aluminum fittings and pipe are unacceptable. Pipe joint compounds must be selected carefully. Oil pump suction lines should be sized so that at maximum suction line flow conditions, the maximum vacuum will not exceed 34 kPa for distillate grade fuels and 50 kPa for residual oils. Oil supply lines to burner oil pumps should not be pressurized by circulating loop systems or aboveground oil storage tanks to more than
Table 41 Recommended Nominal Size for Fuel Oil Suction Lines from Tank to Pump (Residual Grades No. 5 and No. 6) Pumping Rate, L/h 10 50 100 200 300 400 500 600 700 800
40 40 40 50 50 50 65 65 65
Length of Run in Metres at Maximum Suction Lift of 4.5 kPa 20
30
40
50
40 40 50 50 50 65 65 65 65
40 50 50 65 65 65 65 65 80
50 50 50 65 65 65 80 80 80
50 65 65 65 80 80 80 80 100
60
70
80
90
100
50 65 65 65 80 65 65 65 80 80 65 65 80 80 80 80 80 80 80 80 80 80 80 80 100 80 80 80 100 100 80 100 100 100 100 100 100 100 100 100 100 100 100 100 100
Notes: 1. Sizes (in millimetres) are nominal. 2. Pipe sizes smaller than 25 mm ISO are not recommended for use with residual grade fuel oils. 3. Lines conveying fuel oil from pump discharge port to burners and tank return may be reduced by one or two sizes, depending on piping length and pressure losses.
34 kPa, or pump shaft seals may fail. A typical oil circulating loop system is shown in Figure 22. In assembling long fuel pipe lines, be careful to avoid air pockets. On overhead circulating loops, the line should vent air at all high points. Oil supply loops for one or more burners should be the continuous circulation type, with excess fuel returned to the storage tank. Dead-ended pressurized loops can be used, but air or vapor venting is more problematic. Where valves are used, select ball or gate valves. Globe valves are not recommended because of their high pressure drop characteristics. Oil lines should be tested after installation, particularly if they are buried, enclosed, or otherwise inaccessible. Failure to perform this test is a frequent cause of later operating difficulties. A suction line can be hydrostatically tested at 1.5 times its maximum operating pressure or at a vacuum of not less than 70 kPa. Pressure or vacuum tests should continue for at least 60 min. If there is no noticeable drop in the initial test pressure, the lines can be considered tight.
Pipe Sizes for Heavy Oil Tables 41 and 42 give recommended pipe sizes for handling No. 5 and No. 6 oils (residual grades) and No. 1 and No. 2 oils (distillate
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Pipe Design
22.37
Table 42 Recommended Nominal Size for Fuel Oil Suction Lines from Tank to Pump (Distillate Grades No. 1 and No. 2) Pumping Rate, L/h 10 50 100 200 300 400 500 600 700 800
15 15 15 15 20 20 20 20 20
Length of Run in Metres at Maximum Suction Lift of 9.0 kPa 20 30 40 50 60 70 80 90
100
15 15 20 20 20 25 25 25 25
25 25 25 32 32 32 50 50 50
15 15 20 20 20 25 25 25 25
15 15 20 20 20 25 25 25 25
15 20 20 20 25 25 25 25 32
20 20 20 25 25 25 32 32 32
20 20 25 25 25 32 32 32 32
20 20 25 25 25 32 32 32 32
25 25 25 25 32 32 32 50 50
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Note: Sizes (in millimetres) are nominal.
grades), respectively. Storage tanks and piping and pumping facilities for delivering the oil from the tank to the burner are important considerations in the design of an industrial oil-burning system. The construction and location of the tank and oil piping are usually subject to local regulations and National Fire Protection Association (NFPA) Standards 30 and 31.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ASHRAE. 2013. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2013. ASHRAE. 2013. Energy standard for buildings except low-ride residential buildings. ANSI/ASHRAE/IES Standard 90.1-2013. ASME. 2013. Pipe threads, general purpose, inch. Standard B1.20.1-2013. American Society of Mechanical Engineers, New York. ASME. 2006. Pipe threads, 60 deg. general purpose (metric). Standard B1.20.2M-2006. American Society of Mechanical Engineers, New York. ASME. 2015. Gray iron pipe flanges and flanged fittings: Classes 25, 125, and 250. Standard B16.1-2015. American Society of Mechanical Engineers, New York. ASME. 2011. Malleable iron threaded fittings: Classes 150 and 300. Standard B16.3-2011. American Society of Mechanical Engineers, New York. ASME. 2011. Gray iron threaded fittings: Classes 125 and 250. Standard B16.4-2011. American Society of Mechanical Engineers, New York. ASME. 2009. Pipe flanges and flanged fittings: NPS 1/2 through NPS 24 metric/inch standard. Standard B16.5-2009. American Society of Mechanical Engineers, New York. ASME. 2012. Factory made wrought buttwelding fittings. Standard B16.92012. American Society of Mechanical Engineers, New York. ASME. 2009. Forged fittings, socket-welding and threaded. Standard B16.11-2009. American Society of Mechanical Engineers, New York. ASME. 2009. Cast iron threaded drainage fittings. Standard B16.12-2009. American Society of Mechanical Engineers, New York. ASME. 2013. Cast copper alloy threaded fittings: Classes 125 and 250. Standard B16.15-2013. American Society of Mechanical Engineers, New York. ASME. 2012. Cast copper alloy solder joint pressure fittings. Standard B16.18-2012. American Society of Mechanical Engineers, New York. ASME. 2013. Wrought copper and copper alloy solder-joint pressure fittings. Standard B16.22-2013. American Society of Mechanical Engineers, New York. ASME. 2011. Cast copper alloy solder joint drainage fittings: DWV. Standard B16.23-2011. American Society of Mechanical Engineers, New York. ASME. 2011. Cast copper alloy pipe flanges and flanged fittings: Classes 150, 300, 600, 900, 1500, and 2500. Standard B16.24-2011. American Society of Mechanical Engineers, New York.
ASME. 2011. Cast copper alloy fittings for flared copper tubes. Standard B16.26-2011. American Society of Mechanical Engineers, New York. ASME. 2012. Wrought copper and wrought copper alloy solder-joint drainage fittings—DWV. Standard B16.29-2012. American Society of Mechanical Engineers, New York. ASME. 2011. Ductile iron pipe flanges and flanged fittings: Classes 150 and 300. Standard B16.42-2011. American Society of Mechanical Engineers, New York. ASME. 2016. Power piping. Standard B31.1-2016. American Society of Mechanical Engineers, New York. ASME. 2016. Refrigeration piping and heat transfer components. Standard B31.5-2016. American Society of Mechanical Engineers, New York. ASME. 2014. Building services piping. Standard B31.9-2014. American Society of Mechanical Engineers, New York. ASME. 2015. Welded and seamless wrought steel pipe. Standard B36.10M2015. American Society of Mechanical Engineers, New York. ASME. 2015. Qualification standard for welding and brazing procedures, welders, brazers, and welding and brazing operators. Boiler and Pressure Vessel Code, Section IX. American Society of Mechanical Engineers, New York. ASTM. 2012. Standard specification for pipe, steel, black and hot-dipped, zinc-coated, welded, and seamless. Standard A53. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for seamless carbon steel pipe for hightemperature service. Standard A106. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2014. Standard specification for seamless copper water tube. Standard B88. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2016. Standard specification for seamless copper tube for air conditioning and refrigeration field service. Standard B280. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard specification for rigid poly(vinyl chloride) (PVC) compounds and chlorinated poly(vinyl chloride) (CPVC) compounds. Standard D1784. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for poly(vinyl chloride) (PVC) plastic pipe, schedules 40, 80, and 120. Standard D1785. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard test method for determining dimensions of thermoplastic pipe and fittings. Standard D2122. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2012. Standard specification for polyethylene (PE) plastic pipe (SIDR-PR) based on controlled inside diameter. Standard D2239. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for threaded poly(vinyl chloride) (PVC) plastic pipe fittings, schedule 80. Standard D2464. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for poly(vinyl chloride) (PVC) plastic pipe fittings, schedule 40. Standard D2466. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for poly(vinyl chloride) (PVC) plastic pipe fittings, schedule 80. Standard D2467. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2012. Standard specification for solvent cements for poly(vinyl chloride) (PVC) plastic piping systems. Standard D2564. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2014. Standard specification for acrylonitrile-butadiene-styrene (ABS) schedule 40 plastic drain, waste, and vent pipe and fittings. Standard D2661. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2014. Standard specification for poly(vinyl chloride) (PVC) plastic drain, waste, and vent pipe and fittings. Standard D2665. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2013. Standard test method for obtaining hydrostatic design basis for thermoplastic pipe materials or pressure design basis for thermoplastic pipe products. Standard D2837-13e1. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2012. Standard practice for obtaining hydrostatic or pressure design basis for “fiberglass” (glass-fiber-reinforced thermosetting-resin) pipe and fittings. Standard D2992. American Society for Testing and Materials, West Conshohocken, PA.
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22.38 ASTM. 2014. Standard specification for polyethylene plastics pipe and fittings materials. Standard D3350. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2016. Standard classification system and basis for specifications for rigid acrylonitrile-butadiene-styrene (ABS) materials for pipe and fittings. Standard D3965. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for threaded chlorinated poly(vinyl chloride) (CPVC) plastic pipe fittings, schedule 80. Standard F437. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for socket-type chlorinated poly(vinyl chloride) (CPVC) plastic pipe fittings, schedule 40. Standard F438. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2013. Standard specification for chlorinated poly(vinyl chloride) (CPVC) plastic pipe fittings, schedule 80. Standard F439. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for crosslinked polyethylene (PEX) tubing. Standard F876. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard specification for crosslinked polyethylene (PEX) hot- and cold-water distribution systems. Standard F877. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2014. Standard specification for solvent cements for chlorinated poly(vinyl chloride) (CPVC) plastic pipe and fittings. Standard F493. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard test method for evaluating the oxidative resistance of crosslinked polyethylene (PEX) pipe, tubing and systems to hot chlorinated water. Standard F2023. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for pressure-rated polypropylene (PP) piping systems. Standard F2389. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard specification for manufacture and joining of polyethylene (PE) gas pressure pipe with a peelable polypropylene (PP) outer layer. Standard F2830. American Society for Testing and Materials, West Conshohocken, PA. AWWA. 2014. Thickness design of ductile-iron pipe. Standard C150/ A21.50. American Water Works Association, Denver. Ball, E.F., and C.J.D. Webster. 1976. Some measurements of water-flow noise in copper and ABS pipes with various flow velocities. The Building Services Engineer 44(2):33. Carrier. 1960. Piping design. In System design manual. Carrier Air Conditioning Company, Syracuse, NY. CDA. 2010. The copper tube handbook. Copper Development Association, New York. Crane Co. 1976. Flow of fluids through valves, fittings and pipe. Technical Paper 410. Crane Company, New York. Crane Co. 1988. Flow of fluids through valves, fittings and pipe. Technical Paper 410M. Crane Company, New York. Dawson, F.M., and J.S. Bowman. 1933. Interior water supply piping for residential buildings. University of Wisconsin Experiment Station Bulletin 77. Ding, C., L. Carlson, C. Ellis, and O. Mohseni. 2005. Pressure loss coefficients in 6, 8, and 10 inch steel pipe fittings. ASHRAE Research Project TRP-1116, Final Report. University of Minnesota, Saint Anthony Falls Laboratory. Eshbach, O.W. 2009. Eshbach’s handbook of engineering fundamentals, 5th ed. M. Kutz, ed. John Wiley & Sons, New York. Freeman, J.R. 1941. Experiments upon the flow of water in pipes. American Society of Mechanical Engineers, New York. Gas Engineers’ Handbook. 1970. Industrial Press, New York. Giesecke, F.E. 1926. Friction of water elbows. ASHVE Transactions 32:303. Giesecke, F.E., and W.H. Badgett. 1931. Friction heads in one-inch standard cast-iron tees. ASHVE Transactions 37:395. Giesecke, F.E., and W.H. Badgett. 1932a. Loss of head in copper pipe and fittings. ASHVE Transactions 38:529. Giesecke, F.E., and W.H. Badgett. 1932b. Supplementary friction heads in one-inch cast-iron tees. ASHVE Transactions 38:111. Grinnell Company. 1951. Piping design and engineering. Grinnell Company, Cranston, RI. HDR design guide. 1981. Hennington, Durham and Richardson, Omaha, NE.
2017 ASHRAE Handbook—Fundamentals (SI) Heald, C.C. 2002. Cameron hydraulic data, 19th ed. Flowserve Corporation, Irving, TX. Hegberg, R.A. 1995. Where did the k-factors for pressure loss in fittings come from? ASHRAE Transactions 101(1):1264-78. Paper CH-95-20-3. Hunter, R.B. 1940. Methods of estimating loads in plumbing systems. NBS Report BMS 65. National Institute of Standards and Technology, Gaithersburg, MD. Hunter, R.B. 1941. Water distributing systems for buildings. NBS Report BMS 79. National Institute of Standards and Technology, Gaithersburg, MD. Hydraulic Institute. 1990. Engineering data book. Hydraulic Institute, Parsippany, NJ. ICC. 2012. International plumbing code®. International Code Council, Washington, D.C. Idelchik, I.E. 1986. Handbook of hydraulic resistance. Hemisphere Publishing, New York. ISA. 2007. Flow equations for sizing control valves. Standard 75.01.01-07. International Society of Automation, Research Triangle Park, NC. Laschober, R.R., G.Y. Anderson, and D.G. Barbee. 1966. Counterflow of steam and condensate in slightly pitched pipes. ASHRAE Transactions 72(1):157. Marseille, B. 1965. Noise transmission in piping. Heating and Ventilating Engineering (June):674. MSS. 2009. Pipe hangers and supports—Materials, design, manufacture, selection, application, and installation. ANSI/MSS Standard SP-58. Manufacturers Standardization Society of the Valve and Fittings Industry, Vienna, VA. MSS. 2003. Pipe hangers and supports—Selection and application. Standard SP-69. Manufacturers Standardization Society of the Valve and Fittings Industry, Vienna, VA. Nayyar, M. 1999. Piping handbook. McGraw-Hill, New York. NFPA. 2010. Installation of sprinkler systems. Standard 13. National Fire Protection Association, Quincy, MA. NFPA/AGA. 2012. National fuel gas code. ANSI/NFPA Standard 54. National Fire Protection Association, Quincy, MA. ANSI/AGA Standard Z223.1-2002. American Gas Association, Arlington, VA. NSF/ANSI. 2016. Plastics piping system components and related materials. ANSI/NSF Standard 14-2016. NSF International, Ann Arbor, MI. NSF/ANSI. 2016. Drinking water system components—Health effects. ANSI/NSF Standard 61. NSF International, Ann Arbor, MI. Obrecht, M.F., and M. Pourbaix. 1967. Corrosion of metals in potable water systems. AWWA 59:977. American Water Works Association, Denver, CO. PHCC. 2012. National standard plumbing code. Plumbing-Heating-Cooling Contractors Association, Falls Church, VA. Plastic Pipe Institute. 1971. Water flow characteristics of thermoplastic pipe. Plastic Pipe Institute, New York. PPFA. 2009. PVC piping systems for commercial and industrial applications design guide. Plastic Pipe and Fitting Association, Glenn Ellyn, IL Rahmeyer, W.J. 1999a. Pressure loss coefficients of threaded and forged weld pipe fittings for ells, reducing ells, and pipe reducers (RP-968). ASHRAE Transactions 105(2):334-354. Paper 4308. Rahmeyer, W.J. 1999b. Pressure loss coefficients of pipe fittings for threaded and forged weld pipe tees (RP-968). ASHRAE Transactions 105(2):355385. Paper 4309. Rahmeyer, W.J. 2002a. Pressure loss data for large pipe ells, reducers, and expansions. ASHRAE Transactions 108(1):360-375. Paper 4533. Rahmeyer, W.J. 2002b. Pressure loss data for large pipe tees. ASHRAE Transactions 108(1):376-389. Paper 4534. Rahmeyer, W.J. 2002c. Pressure loss coefficients for close-coupled pipe ells. ASHRAE Transactions 108(1):390-406. Paper 4535. Rahmeyer, W.J. 2003a. Pressure loss data for PVC pipe elbows, reducers, and expansions (RP-1193). ASHRAE Transactions 109(2):230-251. Paper 4653. Rahmeyer, W.J. 2003b. Pressure loss data for PVC pipe tees (RP-1193). ASHRAE Transactions 109(2):252-271. Paper 4654. Rogers, W.L. 1953. Experimental approaches to the study of noise and noise transmission in piping systems. ASHVE Transactions 59:347-360. Rogers, W.L. 1954. Sound-pressure levels and frequencies produced by flow of water through pipe and fittings. ASHRAE Transactions 60:411-430. Rogers, W.L. 1956. Noise production and damping in water piping. ASHAE Transactions 62:39.
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Pipe Design
22.39 Williams, G.S., and A. Hazen. 1933. Hydraulic tables. John Wiley & Sons, New York.
BIBLIOGRAPHY ASTM. 2009. Standard specification for chlorinated poly(vinyl chloride) (CPVC) plastic pipe, Schedules 40 and 80. Standard F441/F441M. American Society for Testing and Materials, West Conshohocken, PA.
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Sanks, R.L. 1978. Water treatment plant design for the practicing engineer. Ann Arbor Science, Ann Arbor, MI. Smith, T. 1983. Reducing corrosion in heating plants with special reference to design considerations. Anti-Corrosion Methods and Materials 30 (October):4. Stewart, W.E., and C.L. Dona. 1987. Water flow rate limitations (RP-450). ASHRAE Transactions 93(2):811-825. Paper 3106. Williams, G.J. 1976. The Hunter curves revisited. Heating/Piping/Air Conditioning (November):67.
Related Commercial Resources
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Related Commercial Resources CHAPTER 23
INSULATION FOR MECHANICAL SYSTEMS Design Objectives and Considerations ......................................................................................... 23.1 Materials and Systems................................................................................................................... 23.9 Installation .................................................................................................................................. 23.13 Design Data ................................................................................................................................ 23.18 Project Specifications.................................................................................................................. 23.19
HIS chapter deals with applications of thermal and acoustical insulation for mechanical systems in residential, commercial, and industrial facilities. Applications include pipes, tanks, vessels and equipment, and ducts. Thermal insulation is primarily used to limit heat gain or loss from surfaces operating at temperatures above or below ambient temperature. Insulation may be used to satisfy one or more of the following design objectives:
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T
• Energy conservation: minimizing unwanted heat loss/gain from building HVAC systems, as well as preserving natural and financial resources • Economic thickness: selecting the thickness of insulation that yields the minimum total life-cycle cost • Personnel protection: controlling surface temperatures to avoid contact burns (hot or cold) • Condensation control: minimizing condensation by keeping surface temperature above the dew point of surrounding air • Process control: minimizing temperature change in process fluids where close control is needed • Freeze protection: minimizing energy required for heat tracing systems and/or extending the time to freezing in the event of system failure or when the system is purposefully idle • Noise control: reducing/controlling noise in mechanical systems • Fire safety: protecting critical building elements and slowing the spread of fire in buildings Fundamentals of thermal insulation are covered in Chapter 25; applications in insulated assemblies are discussed in Chapter 27; and data on thermal and water vapor transmission data are in Chapter 26.
1.
awards points based on satisfying performance criteria in several different categories. Different levels of green building certification are awarded based on the total points earned. The role of mechanical insulation in reducing energy usage, along with the associated greenhouse gas emissions, can help to contribute to LEED certification and should be considered when designing an insulation system.
Economic Thickness Economics can be used to (1) select the optimum insulation thickness for a specific insulation, or (2) evaluate two or more insulation materials for least cost for a given level of thermal performance. In either case, economic considerations determine the most cost-effective solution for insulating over a specific period. Life-cycle costing considers the initial cost of the insulation system plus the ongoing value of energy savings over the expected service lifetime. The economic thickness is defined as the thickness that minimizes the total life-cycle cost. Labor and material costs of installed insulation increase with thickness. Insulation is often applied in multiple layers (1) because materials are not manufactured in single layers of sufficient thickness and (2) in many cases, to accommodate expansion and contraction of insulation and system components. Figure 1 shows installed costs for a multilayer application. The slope of the curves is discontinuous and increases with the number of layers because labor and material costs increase more rapidly as thickness increases. Figure 1 shows curves of total cost of operation, insulation costs, and lost energy costs. Point A on the total cost curve corresponds to the economic insulation
DESIGN OBJECTIVES AND CONSIDERATIONS
Energy Conservation Thermal insulation is commonly used to reduce energy consumption of HVAC systems and equipment. Minimum insulation levels for ductwork and piping are often dictated by energy codes, many of which are based on ASHRAE Standards 90.1 and 90.2. In many cases, it may be cost-effective to go beyond the minimum levels dictated by energy codes. Thicknesses greater than the optimum economic thickness may be required for other technical reasons such as condensation control, personnel protection, or noise control. Tables 1 to 3 contain minimum insulation levels for ducts and pipes, excerpted from ANSI/ASHRAE Standard 90.1-2010. Interest in green buildings (i.e., those that are environmentally responsible and energy efficient, as well as healthier places to work) is increasing. The LEED® (Leadership in Energy and Environmental Design) Green Building Rating System™, created by the U.S. Green Building Council, is a voluntary rating system that sets out sustainable design and performance criteria for buildings. It evaluates environmental performance from a whole-building perspective and The preparation of this chapter is assigned to TC 1.8, Mechanical Systems Insulation.
23.1 Copyright © 2017, ASHRAE
Fig. 1 Determination of Economic Thickness of Insulation
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23.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Minimum Duct Insulation R-Value,a Cooling and Heating Only Supply Ducts and Return Ducts
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Duct Location Unvented Attic Above Insulated Ceiling
Climate Zoned
Exterior
Ventilated Attic
1, 2 3 4 5 6 7 8
none R-0.62 R-0.62 R-1.06 R-1.06 R-1.41 R-1.41
none none none R-0.62 R-1.06 R-1.06 R-1.41
none none none none R-0.62 R-1.06 R-1.06
1 2 3 4 5, 6 7, 8
R-1.06 R-1.06 R-1.06 R-0.62 R-0.62 R-0.34
R-1.06 R-1.06 R-1.06 R-0.62 R-0.34 R-0.34
R-1.41 R-1.06 R-1.06 R-1.06 R-0.62 R-0.34
1 to 8
R-0.62
aInsulation
R-0.62
R-0.62
Unvented Attic with Roof Insulationa
Unconditioned Spaceb
Indirectly Conditioned Spacec
Buried
Heating-Only Ducts none none none none none none none
none none none none none R-0.62 R-1.06
none none none none none none none
none none none R-0.62 R-0.62 R-0.62 R-1.06
Cooling-Only Ducts R-0.62 R-0.62 R-0.62 R-0.34 R-0.34 R-0.34
R-0.62 R-0.62 R-0.64 R-0.34 R-0.34 R-0.34
none none none none none none
R-0.62 R-0.62 none none none none
Return Ducts none
none
none
none
(m2 ·K)/W,
R-values, measured in are for the insulation as installed and do not include film resistance. The required minimum thicknesses do not consider water vapor transmission and possible surface condensation. Where exterior walls are used as plenum walls, wall insulation must be as required by the most restrictive condition of Section 6.4.4.2 or Section 5 of 90.1-2010. Insulation resistance measured on a horizontal plane in accordance with ASTM C518 at a mean temperature of 23.9°C at the installed thickness. bIncludes crawlspaces, both ventilated and nonventilated. cIncludes return air plenums with or without exposed roofs above. dClimate zones for the continental United States defined in ASHRAE Standard 90.1-2010.
Table 2
Minimum Pipe Insulation Thickness,a mm
Insulation Conductivity Fluid Design Operating Temp. Range, °C >177 122 to 177 94 to 121 61 to 93 41 to 60 4 to 16 <4
Conductivity, W/(m·K)
Mean Rating Temp., °C
Nominal Pipe or Tube Size, mm 100 to <200
200
Heating Systems (Steam, Steam Condensate, Hot Water, and Domestic Hot Water)b,c 0.046 to 0.049 121 114 127 127 0.042 to 0.046 93 89 102 114 0.039 to 0.043 66 64 64 76 0.036 to 0.042 52 38 38 51 0.032 to 0.040 38 25 25 38
127 114 76 51 38
127 114 76 51 38
Cooling Systems (Chilled Water, Brine, and Refrigerant)d 24 13 13 10 13 25
25 25
25 38
0.032 to 0.040 0.032 to 0.040
aFor
insulation outside stated conductivity range, determine minimum thickness T as follows: T = r{(1 + t/r)K/k – 1} where T = minimum insulation thickness (mm), r = actual outside radius of pipe (mm), t = insulation thickness listed in this table for applicable fluid temperature and pipe size, K = conductivity of alternative material at mean rating temperature indicated for applicable fluid temperature [W/(m·K)]; and k = upper value of conductivity range listed in this table for applicable fluid temperature.
thickness, which, in this example, is in the double-layer range. Viewing the calculated economic thickness as a minimum thickness provides a hedge against unforeseen fuel price increases and conserves energy. Initially, as insulation is applied, the total life-cycle cost decreases because the value of incremental energy savings is greater than the incremental cost of insulation. Additional insulation reduces total cost up to a thickness where the change in total cost is equal to zero. At this point, no further reduction can be obtained; beyond it, incremental insulation costs exceed the additional energy savings derived by adding another increment of insulation. Economic analysis should also consider the time value of money, which can be based on a desired rate of return for the insulation
<25
25 to <40
40 to <100
25 25
bThese
thicknesses are based on energy efficiency considerations only. Additional insulation is sometimes required relative to safety issues/surface temperature.
cPiping insulation is not required between the control valve and coil on run-outs when
the control valve is located within 1.2 m of the coil and the pipe size is 25 mm or less. thicknesses are based on energy efficiency considerations only. Issues such as water vapor permeability or surface condensation sometimes require vapor retarders or additional insulation.
dThese
investment. Energy costs are volatile, and a fuel cost inflation factor is sometimes included to account for the possibility that fuel costs may increase more quickly than general inflation. Insulation system maintenance costs should also be included, along with cost savings associated with the ability to specify lower capacity equipment, resulting in lower first costs. Chapter 37 of the 2015 ASHRAE Handbook—HVAC Applications has more information on economic analysis.
Personnel Protection In many applications, insulation is provided to protect personnel from burns. The potential for burns to human skin is a complex function of surface temperature, surface material, and time of contact.
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Insulation for Mechanical Systems
23.3
Table 3 Minimum Duct Insulation R-Value,a Combined Heating and Cooling Supply Ducts and Return Ducts Duct Location
Climate Zone
Exterior
Ventilated Attic
1 2 3 4 5 6 7 8
R-1.06 R-1.06 R-1.06 R-1.06 R-1.06 R-1.41 R-1.41 R-1.41
R-1.06 R-1.06 R-1.06 R-1.06 R-1.06 R-1.06 R-1.06 R-1.41
1 to 8
R-0.62
R-0.62
Unvented Attic Above Insulated Ceiling R-1.41 R-1.06 R-1.06 R-1.06 R-1.06 R-1.06 R-1.06 R-1.41
Unvented Attic with Roof Insulationa
Supply Ducts R-0.62 R-0.62 R-0.62 R-0.62 R-0.34 R-0.34 R-0.34 R-0.34 Return Ducts R-0.62 none
Unconditioned Spaceb
Indirectly Conditioned Spacec
Buried
R-0.62 R-0.62 R-0.62 R-0.62 R-0.62 R-0.62 R-0.62 R-1.06
none none none none none none none none
R-0.62 R-0.62 R-0.62 R-0.62 R-0.62 R-0.62 R-0.62 R-1.06
none
none
none
aInsulation R-values, measured in (m2 ·K)/W, are for the insulation as installed and do not include film resistance. The required minimum thicknesses do not consider water vapor
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transmission and possible surface condensation. Where exterior walls are used as plenum walls, wall insulation must be as required by the most restrictive condition of Section 6.4.4.2 or Section 5 of 90.1-2010. Insulation resistance measured on a horizontal plane in accordance with ASTM C518 at a mean temperature of 23.9°C at the installed thickness. bIncludes crawlspaces, both ventilated and nonventilated. cIncludes return air plenums with or without exposed roofs above.
ASTM Standard C1055 has a good discussion of these factors. Standard industry practice is to specify a maximum temperature of 60°C for surfaces that may be contacted by personnel. For indoor applications, maximum air temperatures depend on the facility and location, and are typically lower than design outdoor conditions. For outdoor installations, base calculations on summer design ambient temperatures with no wind (i.e., the worst case). Surface temperatures increase because of solar loading, but are usually neglected because of variability in orientation, solar intensity, and many other complicating factors. Engineering judgment must be used in selecting ambient and operating temperatures and wind conditions for these calculations. Note that the choice of jacketing strongly affects a surface’s relative safety. Higher-emittance jacketing materials (e.g., plastic, painted metals) can be selected to minimize the surface temperature. Jacketing material also affects the relative safety at a given surface temperature. For example, at 80°C, a stainless steel jacket blisters skin more severely than a nonmetallic jacket at equal contact time.
Table 4
Insulation Thickness Required to Prevent Surface Condensation
Relative Humidity, %
Thickness, mm
20 30 40 50 60 70 80 90 95
— 2.5 5.0 7.5 13 18 33 74 150
Note: Calculated using Equation (14), assuming surface conductance of 6.8 W/(m2 ·K) and insulation with thermal conductivity of 0.043 W/(m·K). Different assumed values yield different results.
Condensation Control For below-ambient systems, condensation control is often the overriding design objective. The design problem is best addressed as two separate issues: (1) avoiding surface condensation on the outer surface of the insulation system and (2) minimizing or managing water vapor intrusion. Avoiding surface condensation is desirable because it (1) prevents dripping, which can wet surfaces below; (2) minimizes mold growth by eliminating the liquid water many molds require; and (3) avoids staining and possible damage to exterior jacketing. The design goal is to keep the surface temperature above the dew-point temperature of surrounding air. Calculating surface temperature is relatively simple, but selecting the appropriate design conditions is often confusing. The appropriate design condition is normally the worst-case condition expected for the application. For condensation control, however, a design that satisfies the worst case is sometimes impossible. To illustrate, Table 4 shows insulation thicknesses required to prevent condensation on the exterior surface of a hypothetical insulated tank containing a liquid held at 4°C in a mechanical room with a temperature of 27°C. Note that, at high relative humidities, the thickness required to prevent surface condensation increases dramatically, and becomes impractical above 90% rh.
Fig. 2 Relative Humidity Histogram for Charlotte, NC For outdoor applications (or for unconditioned spaces vented to outdoor air), there are always some hours per year where the ambient air is saturated or nearly saturated. For these times, no amount of insulation will prevent surface condensation. Figure 2 shows the frequency distribution of outdoor relative humidity based on typical meteorological year weather data for Charlotte, North Carolina (Marion and Urban 1995). Note that there are over 1200 h per year when the relative humidity is equal to or greater than 90%, and nearly 600 h per year when the relative humidity is equal to or greater than 95%. For outdoor applications and mechanical rooms vented to outdoor conditions, it is suggested to design for a relative humidity of
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23.4
2017 ASHRAE Handbook—Fundamentals (SI) Table 5 Design Weather Data for Condensation Control
City New Orleans, LA Houston, TX Miami, FL Tampa, FL Savannah, GA Norfolk, VA San Antonio, TX Charlotte, NC Honolulu, HI Columbus, OH Minneapolis, MN Seattle, WA
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Fig. 3 ASHRAE Psychrometric Chart No. 1 90%. Appropriate water-resistant vapor-retarder jacketing or mastics must then be specified to protect the system from the inevitable surface condensation. Table 5 summarizes design weather data for a select number of cities. The design dew-point temperature and the corresponding dry-bulb temperatures at 90% rh are given, along with the number of hours per year that the relative humidity would exceed 90%. Additional design dew-point data can be found in Chapter 14. Design Example: Tampa, Florida. Chilled-water supply piping is to be located outdoors to serve a commercial building expansion in Tampa, Florida. The supply piping is 150 mm NPS steel and the design temperature of the chilled-water supply is 4.4°C. Determine the appropriate design ambient conditions for this installation. From Table 5, the design dew-point temperature for Tampa is 25.6°C. The design conditions are best visualized using a psychrometric chart, which graphically represents the properties of moist air. The horizontal axis is dry-bulb temperature, and the vertical axis is humidity ratio (kg of water vapor per kg of dry air). The chart includes the saturation curve (relative humidity = 100%) as well as parallel curves for other values of constant relative humidity. Lines of constant dew-point temperature are horizontal on the psychrometric chart. Using Figure 3, enter the chart on the saturation curve at a dew point of 25.6°C (point A. in the figure) and draw a horizontal line. The design point is located where this horizontal line intersects the 90% rh curve. The dry-bulb temperature associated with this design point is read from the horizontal axis at point C, which for this example is approximately 27°C. The insulation system should therefore be designed for an operating temperature of 4.4°C, an ambient temperature of 27°C, and an ambient relative humidity of 90%.
This section is based on WDBG (2012). For indoor designs in conditioned spaces, care is needed when selecting design conditions. Often, the HVAC system is sized to provide indoor conditions of 24°C/50% rh on a design summer day. However, those indoor conditions do not represent the worst-case indoor conditions for insulation design. Part-load conditions could result in higher humidity levels, or night and/or weekend shutdown could result in more severe conditions. In addition to avoiding condensation on the exposed surface, another important design consideration is minimizing or managing water vapor intrusion, which is extremely important for piping and equipment operating at below-ambient temperatures. Water-related problems include thermal performance loss, health and safety issues, structural degradation, and aesthetic issues. Water entry into
Design Dew- Corresponding Point Temp., Dry-Bulb Temp. Hours per Year °C at 90% rh, °C >90% rh 26.1 25.6 25.6 25.6 25.0 24.4 24.4 23.3 23.3 22.8 22.8 15.6
27.8 27.2 27.2 27.2 26.7 26.1 26.1 25.0 25.0 24.4 24.4 17.2
1253 2105 633 992 1560 1279 932 1233 166 531 619 1212
the insulation system may be through diffusion of water vapor, air leakage carrying water vapor, and leakage of surface water. When the operating temperature is below the dew point of the surrounding ambient air, there is a difference in water vapor pressure across the insulation system. This vapor-pressure difference drives diffusion of water vapor from the ambient toward the cold surface. Piping and equipment typically create an absolute barrier to the passage of water vapor, so any vapor-pressure difference imposed across the insulation system results in the potential for condensation either in the insulation or at the cold surface. The vaporpressure difference can range from below 345 Pa for a supply air duct operating in the return air plenum of a commercial building, to over 4 kPa for a cryogenic system operating outdoors near the U.S. Gulf Coast. Although these pressure differences seem small, the effect over many operating hours can be significant. Several fundamental design principles are used in managing water vapor intrusion. One method is to reduce the driving force by reducing the moisture content of the surrounding air. The insulation designer typically does not have control of the location of the piping, ductwork, or equipment to be insulated, but there are opportunities for the mechanical engineer to influence ambient conditions. Certainly, locating cold piping, ductwork, and equipment in unconditioned portions of buildings should be minimized. Consider conditioning mechanical rooms if feasible. Another common method is moisture blocking, wherein passage of water vapor is eliminated or minimized to an insignificant level. The design must incorporate the following: (1) a vapor retarder with suitably low permeance; (2) a joint and seam sealing system that maintains vapor retarding system integrity; and (3) accommodation for future damage repair, joint and seam resealing, and reclosing after maintenance. A vapor retarder is a material or system that adequately reduces transmission of water vapor through the insulation system. The vapor retarder system is seldom intended to resist entry of surface water or prevent air leakage, but can occasionally be considered the second line of defense for these moisture sources. An effective vapor retarder material or system is essential for blocking systems to perform adequately. Mumaw (2001) showed that the design, installation, and performance of vapor retarder systems are key to the ability of an insulation system to minimize water vapor ingress. Performance of the vapor retarder material or system is characterized by the water vapor permeance: the lower the permeance, the better. The water vapor permeance can be evaluated using procedures outlined in ASTM Standard E96. In this test, a vapor pressure difference is imposed across vapor retarder material that has been sealed to a test cup, and the moisture gain or loss is measured gravimetrically. The insulation system should be dry before applying a vapor retarder to prevent trapping water vapor in the insulation system.
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Insulation for Mechanical Systems
23.5
The insulation system also must be protected from undue weather exposure that could introduce moisture into the insulation before the system is sealed. Faulty application techniques can impair vapor retarder performance. The effectiveness of installation and application techniques must be considered during selection. Factors such as vapor retarder structure, number of joints, mastics and adhesives that are used, as well as inspection procedures affect system performance and durability. When selecting a vapor retarder, the vapor-pressure difference across the insulation system should be considered. Higher vaporpressure differences typically require a vapor retarder with a lower permeance to control the overall moisture pickup of the insulated system. Service conditions affect the direction and magnitude of the vapor pressure difference: unidirectional flow exists when the water vapor pressure is constantly higher on one side of insulation system, whereas reversible flow exists when vapor pressure may be higher on either side (typically caused by diurnal or seasonal changes on one side of the insulation system). Properties of the insulation system materials should be considered. All materials reduce the flow of water vapor; the low permeance of some insulation materials can add to the overall resistance to water vapor transport of the insulation system. All vapor retarder joints should be tightly sealed with manufacturer-recommended sealants. Another fundamental design principle is moisture storage design. In many systems, some condensation can be tolerated, the amount depending on the water-holding capacity or tolerance of a particular system. The moisture storage principle allows accumulation of water in the insulation system, but at a rate designed to prevent harmful effects. This concept is applicable when (1) unidirectional vapor flow occurs, but accumulations during severe conditions can be adequately expelled during less severe conditions; or (2) reverse flow regularly occurs on a seasonal or diurnal cycle. Design solutions using this principle include (1) periodically flushing the cold side with low-dew-point air (requires a supply of conditioned air and a means for distribution), and (2) using an insulation system supplemented by selected vapor retarders and absorbent materials such that an accumulation of condensation is of little importance. Such a design must ensure sufficient expulsion of accumulated moisture. ASTM Standard C755 discusses various design principles. Chapters 25 to 27 of this volume thoroughly describe the physics associated with water vapor transport. Additional information is found in Chapter 10 of the 2014 ASHRAE Handbook—Refrigeration, and in ASTM (2001).
Freeze Prevention It is important to recognize that insulation retards heat flow; it does not stop it completely. If the surrounding air temperature remains low enough for an extended period, insulation cannot prevent freezing of still water or of water flowing at a rate insufficient for the available heat content to offset heat loss. Insulation can prolong the time required for freezing, or prevent freezing if flow is maintained at a sufficient rate. To calculate time (in hours) required for water to cool to 0°C with no flow, use the following equation: = (1/3600)Cp(D1/2)2 RT ln[(ti – ta)/(tf –ta)] where time to freezing, h density of water = 1000 kg/m3 specific heat of water = 4200 J/(kg·K) inside diameter of pipe, m (see Figure 4) combined thermal resistance of pipe wall, insulation, and exterior air film (for a unit length of pipe) ti = initial water temperature, °C ta = ambient air temperature, °C
Cp D1 RT
= = = = =
(1)
Fig. 4 Time to Freeze Nomenclature Table 6 Time to Cool Water to Freezing, h Insulation Thickness, mm
Nominal Pipe Size, mm
13
25
38
50
75
100
15 25 40 50 80 100 125 150 200 250 300
0.1 0.3 0.4 0.6 0.9 1.3 1.6 1.9 — — —
0.2 0.4 0.8 1.1 1.7 2.4 3.0 3.7 5.3 6.5 8.8
0.2 0.5 1.0 1.4 2.3 3.3 4.3 5.3 7.6 10.2 12.5
0.3 0.6 1.3 1.7 2.9 4.1 5.4 6.9 9.6 12.9 15.8
— 0.8 1.5 2.2 3.7 5.5 7.4 9.4 13.7 17.9 22.1
— — — 2.5 4.5 6.6 9.1 11.7 16.9 22.3 27.7
Note: Assumes initial temperature = 5.5°C, ambient air temperature = –28°C, and insulation thermal conductivity = 0.0432 W/(m·K). Thermal resistances of pipe and air film are neglected. Different assumed values yield different results.
tf = freezing temperature, °C
As a conservative assumption for insulated pipes, thermal resistances of pipe walls and exterior air film are usually neglected. Resistance of the insulation layer for a unit length of pipe is calculated as RT = ln(D3/D2)/(2k)
(2)
where D3 = outer diameter of insulation, m D2 = inner diameter of insulation, m k = thermal conductivity of insulation material, W/(m·K)
Table 6 shows estimated time to freezing, calculated using these equations for the specific case of still water with ti = 6°C and ta = –28°C. When unusual conditions make it impractical to maintain protection with insulation or flow, a hot trace pipe or electric resistance heating cable is required along the bottom or top of the water pipe. The heating system then supplies the heat lost through the insulation. Clean water in pipes usually supercools several degrees below freezing before any ice is formed. Then, upon nucleation, dendritic ice forms in the water and the temperature rises to freezing. Ice can be formed from water only by the release of the latent heat of fusion (335 kJ/kg) through the pipe insulation. Well-insulated pipes may greatly retard this release of latent heat. Gordon (1996) showed that water pipes burst not because of ice crystal growth in the pipe, but because of elevated fluid pressure within a confined pipe section occluded by a growing ice blockage.
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23.6
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Noise Control Duct Insulation. Without insulation, the acoustical environment of mechanically conditioned buildings can be greatly compromised, resulting in reduced productivity and a decrease in occupant comfort. HVAC ducts act as conduits for mechanical equipment noise, and also carry office noise between occupied spaces. Additionally, some ducts can create their own noise through duct wall vibrations or expansion and contraction. Lined sheet metal ducts and fibrous glass rigid ducts can greatly reduce transmission of HVAC noise through the duct system. The insulation also reduces cross-talk from one room to another through the ducts. A good discussion of duct acoustics is provided in Chapter 48 of the 2015 ASHRAE Handbook—HVAC Applications. Duct insulation can be used to provide both attenuation loss and breakout noise reduction. Attenuation loss is noise absorbed within the duct. In uninsulated ducts, it is a function of duct geometry and dimensions as well as noise frequency. Internal insulation liners are generally available for most duct geometries. Chapter 48 in the 2015 ASHRAE Handbook—HVAC Applications provides attenuation losses for square, rectangular, and round ducts lined with fibrous glass, and also gives guidance on use of insulation in plenums to absorb duct system noise. Internal linings can be very effective in fittings such as elbows, which can have 2 to 8 times more attenuation than an unlined elbow of the same size. For alternative lining materials, consult individual manufacturers. It is difficult to write specifications for sound attenuation because it changes with every duct dimension and configuration. Thus, insulation materials are generally selected for attenuation based on sound absorption ratings. Sound absorption tests are run per ASTM Standard C423 in large reverberation rooms with random sound incidence. The test specimens are laid on the chamber floor per ASTM Standard E795, type A mounting. This mode of sound exposure is different from the exposure of internal linings installed in an air duct; therefore, sound absorption ratings for materials can only be used for general comparisons of effectiveness when used in air ducts of varying dimensions (Kuntz and Hoover 1987). Breakout noise is from vibration of the duct wall caused by air pressure fluctuations in the duct. Absorptive insulation can be used in combination with mass-loaded jacketing materials or mastics on the duct exterior to reduce breakout noise. This technique is only minimally effective on rectangular ducts, which require the insulation and mass composite to be physically separated from the duct wall to be very effective. For round ducts, as with pipes, absorptive insulation and mass composite can be effective even when directly applied to the duct surface. Chapter 48 in the 2015 ASHRAE Handbook—HVAC Applications provides breakout noise guidance data. Noise Radiating from Pipes. Noise from piping can be reduced by adding an absorptive insulation and jacketing material. By knowing the sound insertion loss of insulation and jacketing material combinations, the expected level of noise reduction in the field can be estimated. A range of jacket weights and insulation thicknesses can be used to reduce noise. Jackets used to reduce noise are typically referred to as being mass filled. Some products for outdoor applications use mass-filled vinyl (MFV) in combination with aluminum. Pipe insertion loss is a measurement (in dB) of the reduction in sound pressure level from a pipe as a result of application of insulation and jacketing. Measured at different frequencies, the noise level from the jacketed pipe is subtracted from that of the bare pipe; the larger the insertion loss number, the larger the amount of noise reduction. ASTM Standard E1222 describes how to determine insertion loss of pipe jacketing systems. A band-limited white noise test signal is produced inside a steel pipe located in a reverberation room, using a loudspeaker or acoustic driver at one end of the pipe to produce the noise. Average sound pressure levels are measured in the room for
Fig. 5 Insertion Loss Versus Mass of Jacket two conditions: with sound radiating from a bare pipe, and with the same pipe covered with a jacketing system. The insertion loss of the jacketing system is the difference in the sound pressure levels measured, adjusted for changes in room absorption caused by the jacketing system’s presence. Results may be obtained in a series of 100 Hz wide bands or in one-third octave bands from 500 to 5000 Hz. Table 7 gives measured insertion loss values for several pipe insulation and jacket combinations. The mass of the jacket material significantly affects insertion loss of pipe insulation systems. Figure 5 represents insertion loss of typical fibrous pipe insulations with various weights of jacketing (Miller 2001). It is very important that sound sources be well identified in industrial settings. It is possible to treat a noisy pipe very effectively and have no significant influence on ambient sound measurement after treatment. All sources of noise above desired levels must receive acoustical treatment, beginning with the largest source, or no improvement will be observed.
Fire Safety Materials used to insulate mechanical equipment generally must meet the requirements of local codes adopted by governmental entities having jurisdiction over the project. In the United States, most local codes incorporate or are patterned after model codes developed and maintained by organizations such as the National Fire Protection Association (NFPA) and International Code Council (International Codes). Refer to local codes to determine specific requirements. Most codes related to insulation product fire safety refer to the surface burning characteristics as determined by the Steiner tunnel test (ASTM Standard E84, UL Standard 723, or CAN/ULC Standard S-102). These similar test methods evaluate the flame spread and smoke developed from samples mounted in a 7.62 m long tunnel and subsequently exposed to a controlled flame. Results are given in terms of flame spread and smoke developed indices, which are relative to a baseline index and calibration standards of inorganic reinforced cement board (0) and select-grade red oak flooring (100). Samples are normally mounted with the exposed surface face down in the ceiling of the tunnel. Upon ignition, the progress of the flame front is timed while being tracked visually for distance down the tunnel, with the results used to calculate the flame spread index. Smoke index is determined by measuring smoke density with a light cell mounted in the exhaust stream. Using supporting materials on the underside of the test specimen can lower the flame spread index. Materials that melt, drip, or delaminate to such a degree that the continuity of the flame front is destroyed give low flame spread indices that do not relate directly to indices obtained by testing materials that remain in place. Alternative
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Insulation for Mechanical Systems Table 7
Insertion Loss for Pipe Insulation Materials, dB Frequency, Hz
Pipe Size, mm
Insulation Material
Insulation Thickness, mm
Jacket
500
1000
2000
4000
150
Fibrous glass
50 50 50 100 100 13 25 13 25
ASJa 0.5 mm aluminum 5 kg/m2 MFVb with Al ASJ 0.5 mm aluminum None None 5 kg/m2 MFV with Al 5 kg/m2 MFV with Al
2 3 13 4 3 0 0 0 0
9 16 20 21 17 2 2 14 16
14 24 32 27 27 5 5 18 20
16 33 40 33 42 10 10 20 26
50 50 100 100 100 50 75
ASJ 0.5 mm aluminum ASJ 0.5 mm aluminum 5 kg/m2 MFV with Al 0.4mm aluminum 0.4mm aluminum
0 4 8 12 14 1 0
12 19 16 22 23 9 14
19 25 22 30 31 18 19
23 26 26 32 31 28 30
Flexible elastomeric
300
Fibrous glass
Mineral wool aASJ
23.7
= all-service jacket, a typical factory-applied vapor retarder applied to many products. = mass filled vinyl, a field-installed jacket, which has considerably more mass than ASJ.
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bMFV
means of testing may be necessary to fully evaluate some of these materials. For pipe and duct insulation products, samples are prepared and mounted in the tunnel per ASTM Standard E2231, which directs that “the material, system, composite, or assembly tested shall be representative of the completed insulation system used in actual field installations, in terms of the components, including their respective thicknesses.” Samples are constructed to mimic, as closely as possible, the products as they will be used, including any facings and adhesives as appropriate. Duct insulation generally requires a flame spread index of not more than 25 and a smoke developed index of not more than 50, when tested in accordance with ASTM Standard E84. Codes often require factory-made duct insulations (e.g., insulated flexible ducts, rigid fibrous glass ducts) to be listed and labeled per UL Standard 181. This standard specifies several other fire tests (e.g., flame penetration and low-energy ignition) as part of the listing requirements. Some building codes require that duct insulations meet the fire hazard requirements of NFPA Standard 90A or 90B, to restrict spread of smoke, heat, and fire through duct systems, and to minimize ignition sources. Local code authorities should also be consulted for specific requirements. For pipe insulation, the requirement is generally a maximum flame spread index of 25 and a maximum smoke developed index of 450 in nonplenum spaces (in plenums, less than or equal to 25 and 50, respectively). Consult local code authorities for specific requirements. The term noncombustible, as defined by building codes, refers to materials that pass the requirements of ASTM Standard E136. This test method involves introducing a small specimen of the material into a furnace initially maintained at a temperature of 750C. The temperature rise of the furnace is monitored and the specimen is observed for any flaming. Criteria for passing include limits on temperature rise, flaming, and weight loss of the specimen. Some building codes accept as noncombustible a composite material having a structural base of noncombustible material and a surfacing not more than 3 mm thick that has a flame spread index not greater than 50. A related term sometimes referenced in building codes is limited combustible, which is an intermediate category that considers the potential heat content of materials determined per the testing requirements of NFPA Standard 259. Mechanical insulation materials are often used as a component in systems or assemblies designed to protect buildings and equip-
ment from the effects or spread of fire (i.e., fire-resistance assemblies). They can include walls, roofs, floors, columns, beams, partitions, joints, and through-penetration fire stops. Specific designs are tested and assigned hourly ratings based on performance in full-scale fire tests. Note that insulation materials alone are not assigned hourly fire resistance ratings; ratings are assigned to a system or assembly that may include specific insulation products, along with other elements such as framing members, fasteners, wallboard, etc. Fire resistance ratings are often developed using ASTM Standard E119. This test exposes assemblies (walls, partitions, floor or roof assemblies, and through-penetration fire stops) to a standard fire exposure controlled to achieve specified temperatures throughout a specified time period. The time/temperature curve is intended to represent building fires where the primary fuel is solid, and specifies a temperature of 540°C at 5 min, 930°C at 1 h, and 1260°C at 8 h. In the hydrocarbon processing industry, liquid hydrocarbonfueled pool fires are a concern; fire resistance ratings for these applications are tested per ASTM Standard E1529. This time/temperature curve rises rapidly to 1090°C within 5 min, and remains there for the duration of the test. Fire-resistant rated designs can be found in the directories of listing agencies. Examples of such agencies include Underwriters Laboratories, Factory Mutual, UL Canada, and Intertek. The following standard cross-references are provided for products to be tested in Canada. Although these are parallel Canadian standards, the requirements may differ from those of ASTM standards. ASTM Standard CAN/ULC Standard S102, Standard Method of E84 Test for Surface Burning Characteristics of Building Materials and Assemblies ASTM Standard CAN/ULC Standard S101-M, Standard Methods E119 of Fire Endurance Tests of Building Construction and Materials ASTM Standard CAN Standard 4-S114, Standard Method of E136 Test for Determination of Noncombustibility in Building Materials ASTM Standard There is no Canadian equivalent standard on this E1529 subject
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23.8
Licensed for single user. © 2017 ASHRAE, Inc.
Corrosion Under Insulation Corrosion of metal pipe, vessels, and equipment under insulation, though not typically caused by the insulation, is still a significant issue that must be considered during the design of any mechanical insulation system. The propensity for corrosion depends on many factors, including the ambient environment, operating temperature of the metal, proper installation, and maintenance of the insulated system. Corrosion under insulation (CUI) is most prevalent in outdoor industrial environments such as refineries and chemical plants. Corrosion can be very costly because of forced downtime of processes and can be a health and safety hazard as well. Although insulation itself may not necessarily be the cause of corrosion, it can be a passive component because it is in direct contact with the pipe or equipment surface. Very little information is published on corrosion in commercial environments. Although corrosion under insulation is less likely to be a major concern for most insulated surfaces located indoors, it may be a factor on indoor systems that are frequently washed down, such as in the food processing industry. Water from condensation on cold surfaces can be present on both indoor and outdoor insulation systems if there is damage to the vapor retarder. Hot processes can also be subject to condensation during periods of system shutdown. The following factors may lead to corrosion under insulation. • Water must be present. Water ingress may occur at some point on insulated surfaces. The entry point for water is through breaks in weatherproofing materials such as lagging, mastic, caulk, or adhesives. • A general lack of inspection and maintenance increases the potential for corrosion. • Temperature affects the rate of corrosion. In general, temperatures up to 175°C increase the corrosion rate (NACE Standard SP0198). • Contaminants in the plant environment can accelerate corrosion. For instance, chlorides, sulfates, and other corrosion-causing ions could reside on the insulation’s exterior jacket, then be washed into the insulation system by rain or washdown. Other sources of corrosive ions include rainwater, ocean mist, and cooling tower spray, each of which can provide a major and virtually inexhaustible supply of ions. Even if the level of ions in the water is low, significant amounts of ions can accumulate at the pipe surface by a continuing cycle of water penetrations and evaporation. Chloride and other ions only contribute to stress corrosion cracking of stainless steel when water (liquid or vapor) and these ions are present at the surface of a pipe at temperatures above ambient, usually when the surface is above about 60°C and below about 150°C; water and corrosive ions on the pipe surface may contribute to normal oxidative (rusting) corrosion of carbon steel at temperatures between –4 and 175°C [see Kalis (1999) and NACE Standard SP0198]. Exposure of the insulation system to water from some outside source is inevitable, so the key to eliminating stress corrosion cracking lies in preventing moisture and ions, even in small amounts, from reaching the metal surface. • Insulation can contain leachable corrosive agents. Information on the potential corrosion of carbon steel by insulation materials is available from tests conducted using ASTM Standard C1617. The information may be available from the insulation ASTM material standard or from the insulation manufacturer. • Austenitic stainless steels are particularly susceptible to attack from chlorides. Austenitic stainless steels are generally classified as “18-8s”: austenitic alloys containing approximately 18% chromium, 8% nickel, and the balance iron. Besides the basic alloy UNS S30400, these stainless alloys include molybdenum (UNS S31600 and S31700), carbon-stabilized (UNS S321000 and
2017 ASHRAE Handbook—Fundamentals (SI) S347000), and low-carbon grades (UNS S30403 and S31603) (NACE Standard SP0198). For outdoor applications, to prevent ingress of moisture and corrosive ions from precipitation, a properly designed, installed, and maintained weather-protective jacket is recommended. For more specific guidelines, consult the insulation material manufacturer. If process temperatures are lower than ambient (even for short periods of time such as during shutdowns), a vapor retarder is also required. Even with a protective jacket and vapor retarder, it is likely that some moisture and ions will eventually enter the system because of abuse, wear, age, or improper installation. No installation is ideal in any real-world setting. Because most people only consider the chlorides arising from the insulation material, the crucial issue of water and ions infiltrating the insulation system from the environment and yielding metal corrosion remains largely unaddressed. Thus, painting the pipe is the second and most important line of defense, and is necessary if pipe temperature is in the 60 to 150°C range for significant periods of time. To minimize the potential for corrosion, metal pipe should be primed with, for example, an epoxy coating. This alternative offers superior protection against corrosion, because priming protects against ions arising from the insulation and, more importantly, from ions that enter the system from the environment. To minimize corrosion, • Design, install, and maintain insulation systems to minimize ponding water or penetration of water into the system. Flat sections should be designed with a pitch to shed water. Top sections should overlap the sides to provide a watershed effect on ducts, preventing water penetration in the seam. Jacketing joints should be oriented so as to shed water. Design should always minimize penetrations; necessary protrusions (e.g., supports, valves, flanges) should be designed to shed rather than capture water. Water from external sources can enter at any discontinuity in the insulation system. • Insulation should be appropriate for its intended application and service temperature. NACE Standard RP0198 states, “CUI of carbon steel is possible under all types of insulation. The insulation type may only be a contributing factor. The insulation characteristics with the most influence on CUI are (1) water-leachable salt content in insulation that may contribute to corrosion, such as chloride, sulfate, and acidic materials in fire retardants; (2) water retention, permeability, and wettability of the insulation; and (3) foams containing residual compounds that react with water to form hydrochloric or other acids. Because CUI is a product of wet metal exposure duration, the insulation system that holds the least amount of water and dries most quickly should result in the least amount of corrosion damage to equipment.” • Ancillary materials used for weatherproofing (e.g., sealants, caulks, weather stripping, adhesives, mastics) should be appropriate for the application, and be applied following the manufacturer’s recommendations. • Maintenance should monitor for and immediately repair compromises in the protective jacketing system. Because water may infiltrate the insulation system, inspection ports should be used to facilitate inspection without requiring insulation removal. This is particularly important on subambient systems. • Because some water will eventually enter the system, a protective pipe coating is necessary for good design. The type of coating depends on temperature (see NACE Standard RP0198 for coating guidelines). In Europe, essentially all piping is coated for corrosion protection. This is not necessarily the case in the United States, but should be considered as part of good design practice. • When using austenitic stainless steel, all insulation products and accessories should meet the requirements of ASTM Standard C795 if the system will operate at or spend time between 60 and
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Insulation for Mechanical Systems 150°C. Likewise, any ancillary weatherproofing materials should have low chloride content.
2.
MATERIALS AND SYSTEMS
Categories of Insulation Materials
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Turner and Malloy (1981) categorized insulation materials into four types: • Fibrous insulations are composed of small-diameter fibers that finely divide the air space. The fibers may be organic or inorganic, and are normally (but not always) held together by a binder. • Granular insulations are composed of small nodules that contain voids or hollow spaces. These materials are sometimes considered open-cell materials, because gases can transfer between the individual spaces. • Cellular insulations are composed of small, individual cells, either interconnecting or sealed from each other, to form a cellular structure. Glass, plastics, and rubber may comprise the base material, and various foaming agents are used. Cellular insulations are often further classified as either opencell (i.e., cells are interconnecting) or closed-cell (i.e., cells sealed from each other). Generally, materials with greater than 90% closed-cell content are considered to be closed-cell materials. • Reflective insulations and treatments are added to surfaces to lower long-wave emittance, thereby reducing radiant heat transfer from the surface. Low-emittance jackets and facings are often used in combination with other insulation materials. Another material sometimes called thermal insulating paint or coating is available for use on pipes ducts and tanks. These products’ performance must be clearly understood before using them as thermal insulation. These paints and coatings have not been extensively tested and additional research is needed to verify their performance. Further discussion of these products can be found in Hart (2006).
Physical Properties of Insulation Materials Selecting an insulation material for a particular application requires understanding the various physical properties associated with available materials. Operating temperature is often the primary consideration. Maximum temperature capability is normally assessed using ASTM Standard C411 by exposing samples to hot surfaces for an extended time, and assessing the materials for any changes in properties. Evidence of warping, cracking, delamination, flaming, melting, or dripping are indications that the maximum use temperature of the material has been exceeded. There is currently no industry-accepted test method for determining the minimum operating temperature of an insulation material, but minimum temperatures are normally determined by evaluating the material’s integrity and physical properties after exposure to low temperatures. Thermal conductivity of insulation materials is a function of temperature. Many specifications call for insulation conductivity values evaluated at a mean temperature of 24°C. Most manufacturers provide conductivity data over a range of temperatures to allow evaluations closer to actual operating conditions. Conductivity of flat product is generally measured per ASTM Standards C177 or C518, whereas pipe insulation conductivity is generally determined using ASTM Standard C335. Additional information on this property is presented in the following section. Compressive resistance is important where the insulation must support a load without crushing (e.g., insulation inserts in pipe hangers and supports). When insulation is used in an expansion or contraction joint to take up a dimensional change, lower values of compressive resistance are desirable. ASTM Standard C165 is used
23.9 to measure compressive resistance for fibrous materials, and ASTM Standard D1621 is used for foam plastic materials. Water vapor permeability is the water vapor flux through a material induced by a unit vapor pressure gradient across the material. For insulating materials, it is commonly expressed in units of ng/(s·m·Pa). A related and often-confused term is water vapor permeance [in ng/(s·m2 ·Pa)], which measures water vapor flux through a material of specific thickness and is generally used to define vapor retarder performance. In below-ambient applications, it is important to minimize the rate of water vapor flow to the cold surface. This is normally accomplished by using vapor retarders or insulation materials (e.g., cellular glass insulation) with a permeance less than or equal to 1.15 ng/(s·m2·Pa), or both. However, some flexible closed-cell insulation materials have been used successfully without a separate vapor retarder material. ASTM Standard E96 is used to measure water vapor transmission properties of insulation materials. Water absorption is generally measured by immersing a sample of material under a specified pressure of water for a specified time period. It is a useful measure of the amount of liquid water absorbed from water leaks in weather barriers or during construction. Typical physical properties of interest are given in Table 8. Values in this table are taken from the relevant ASTM material specification with permission from ASTM International. Within each material category, a variety in types and grades of materials exist. A representative type and grade are listed in Table 8 for each material category; refer to ASTM standards or to manufacturers for specific data. Thermal Conductivity of Below-Ambient Pipe Insulation Systems. Mechanical pipe insulation systems are installed around cold cylindrical surfaces, such as chilled pipes, and work below ambient temperature in several industrial and commercial building applications. Thermal performance of a pipe insulation system is affected by ambient temperature and humidity and might vary gradually with time. For below-ambient temperature applications of pipe insulation, most published data are extrapolated from flat slab configurations of insulation material, and may not be accurate for cylindrical pipe insulation systems because of radial configuration and longitudinal split joints. Thus, ASHRAE research project RP1356 (Cremaschi et al. 2012) developed an experimental apparatus to measure the thermal conductivity of mechanical pipe insulation systems below ambient temperature. Thermal conductivities of five pipe insulation systems under low-humidity, noncondensing conditions are provided in Table 9 at mean insulation temperatures of 13 and 24°C; the insulation was installed on a 75 mm nominal pipe size diameter aluminum pipe, and the test specimens were 1 m long. Radial heat flux was inward and ranged from 7.6 to 33.3 W/m. Nominal wall thickness of the pipe insulation systems varied from 250 to 500 mm. Vapor barriers on the outer surface of the pipe insulation were not installed. The dry test were performed with the aluminum pipe surface temperature at 4.7°C ± 0.3 K, ambient temperature from 22.8 to 43.3°C, and air dew-point temperature below 4.4°C. For these test conditions, water vapor does not condense on the aluminum pipe surface. The relation between the pipe insulation’s thermal conductivity and its mean temperature is linear, and the thermal conductivity of pipe insulation system had a weak dependence on the nominal wall thickness. Note that, for some cases, joint sealant is recommended for the installation of the pipe insulation. Values in Table 9 represent a combined thermal conductivity of the pipe insulation with a certain type of joint sealant applied on the longitudinal joints, where two C-shells come in contact with each other. The combined thermal conductivity of the pipe insulation might be higher than the thermal conductivity of pipe insulation C-shells that are mechanically joint together without sealant on the longitudinal joints. For example, if a butyl rubber sealant with a layer thickness ranging
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23.10
2017 ASHRAE Handbook—Fundamentals (SI) Table 8 Performance Property Guide for Insulation Materials Cellular Polystyrene
Cellular Polyisocyanurate
C547, C553, C612 C552 C578 Type IVB Type I Type XIII Category 1 Grade 1 649 427 74 –18 –268 –183 N/S 45 at failure 138 at 10%
Calcium Flexible Silicate Elastomeric ASTM Standard Type/grade listed
C533 Type I
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Max. operating temperature, °C Min. operating temperature, °C Min. compressive resistance, kPa Max. thermal conductivity, W/(m·K) –18°C mean –4°C 24°C 93°C 204°C 316°C Maximum water vapor permeability, ng/ (s·m·Pa) Maximum liquid water absorption, % volume Maximum water vapor sorption, % mass Maximum surface burning characteristics
Mineral Fiber
Cellular Glass
Cellular Phenolic
Cellular Polyolefin
C591 Type IV Grade 2 150 –183 145 at 10%
C1126 Type III 125 –179 124
C1427 Type I Grade 1 93 –101 N/S
649 60 688 at 5%
C534 Type I Grade 1 104 –57 N/S
N/A N/A N/A 0.065 0.079 0.095 N/S
0.038 N/S 0.040 N/A N/A N/A 0.15
N/A 0.034 0.035 0.043 0.061 0.091 N/A
0.039 N/S 0.045 0.058 0.084 N/A 0.007
0.032 0.034 0.037 N/A N/A N/A 2.2
0.026 N/S 0.026 0.035 N/A N/A 5.8
0.02 N/S 0.02 0.04 N/A N/A 0.72
0.048 N/S 0.050 N/A N/A N/A 0.07
N/S
0.2
N/S
0.5
0.5 (24 h)
0.5 (24 h)
0.43
0.2
N/S 0/0
N/S 25/50
5 25/50
N/S 5/0
N/S N/S
N/S N/S
N/S 25/50
N/S N/S
Note: N/A = not applicable. N/S = not stated (i.e., ASTM standards do not include a value for this property). Properties not stated do not necessarily indicate that material is not appropriate for a given application depending on that property. See previous editions of ASHRAE Handbook—Fundamentals for data on historical insulation materials.
Table 9
Pipe Insulation Material Cellular glass PIR Glass fiber Elastomeric rubber Phenolic
Thermal Conductivities of Cylindrical Pipe Insulation at 12.8 and 24°C
Nominal Wall Thickness mm 25.4 50.8 25.4 50.8 50.8 25.4 50.8
Thermal Conductivity Joint Sealant Type
Butyl rubber Butyl rubber Butyl rubber Contact cement Contact cement Butyl rubber Butyl rubber
at 12.8°C, W/(m·K)
at 24°C, W/(m·K)
0.0432 0.0399 0.0282 0.0335 0.0347
0.0454 0.0467 0.0293 0.0345 0.0358
0.0322 0.0271
0.0344 0.0305
from 1.59 to 2.54 mm is used, it is possible that the combined thermal conductivity increases up to 15% with respect to the value of the same pipe insulation system without joint sealant. ASHRAE research project RP-1356 also measured the thermal conductivity of mechanical pipe insulation systems below ambient temperature in high humidity without vapor retarders, resulting in rapid moisture ingress. Two types of pipe insulation, installed on a 75 mm nominal pipe size diameter aluminum pipe that was 1 m long, were exposed for less than a month to a warm, humid environment, resulting in water vapor condensation in the insulation samples and increased thermal conductivities. The nominal wall thickness of the pipe insulation systems was 50 mm. Vapor barriers were not installed on the outer surface of the pipe insulation, and the thermal conductivities increased as result of condensed water being retained into the insulation systems. For one type of pipe insulation, the thermal conductivity increased by 3.15 times when the moisture content was about 11% volume. For the other insulation, the thermal conductivity was 1.55 times of the original dry value when the moisture content reached 5% by volume. Each test was run for less than 1 month at high ambient humidity (>80% rh at 35°C. The test conditions were intentionally different from each other, and the thermal performance of the two
pipe insulation systems tested in warm, high-humidity conditions should not be compared. Caution is needed in using this data. Only two types of pipe insulation were tested, and neither had a vapor retarder. The manufacturers of these materials do not recommend these pipe insulation materials be installed in this manner for below-ambient applications. Nevertheless, these data are significant because they demonstrate both the necessity of installing an effective water vapor retarder and the negative impact of water retention in the insulation on its thermal conductivity. These results are only an example of this phenomenon, and any insulation that absorbs and retains water could be similarly affected.
Weather Protection Weather barriers, often referred to as jacketing, are extremely important. Premature failure can lead to insulation failure, with safety and economic consequences. Safety consequences • If insulation is installed for burn protection from a hot pipe or equipment, water entering the insulation system can vaporize into steam and cause a surface temperature well above the expected 60°C, the common design temperature for personnel protection. • Pipe or equipment can corrode, rupture, and release a hazardous material. Economic consequences • Wet insulation has higher thermal conductivity and lower insulation values. • On a hot system, 1 kg of water entering the system requires 1 kJ to revaporize. If this vapor cannot vent easily, it can condense, causing interior jacket corrosion; the weather barrier will begin consuming itself from the inside out. Consequently, the system cannot deliver the desired energy efficiency and will quickly require an expensive repair. • On a very cold system, improper vapor retarder selection allows moisture to migrate to the cold surface because of the continuous drive of vapor pressure. A hole in the system allows direct water influx. Either of these entry mechanisms results in ice formation,
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Insulation for Mechanical Systems which separates the insulation and weather protection barrier from the pipe or vessel surface and compromises thermal performance.
Licensed for single user. © 2017 ASHRAE, Inc.
Many more scenarios must be considered, especially when the broad range of features required of a weather barrier are considered. Turner and Malloy (1981) define a weather barrier as “a material or materials, which, when installed on the outer surface of thermal insulation, protects the insulation from . . . rain, snow, sleet, wind, solar radiation, atmospheric contamination and mechanical damage.” With this definition in mind, several service requirements must be considered. • Internal mechanical forces: Expansion and contraction of the pipe or vessel must be considered because the resulting forces are transferred to the external surface of the weather barrier. Ability to slide, elongate, or contract must be provided. • External mechanical forces: Mechanical abuse (i.e., tools being dropped, abrasion from wind-driven sand, personnel walking on the system) inflicted on a pipe or vessel needs to be considered in design. This may affect insulation type, as well as the weather barrier jacketing type. • Dimensional stability: Some cellular materials can show irreversible dimensional change after installation. Manufacturers of these materials provide installation guidelines to minimize the effects of dimensional change. If guidelines are not followed, failure of joint seals can occur, which can lead to system failure. • Chemical resistance: Some industrial environments may have airborne or spilled corrosive agents that accumulate on the weather barrier and chemically attack the pipe or vessel jacketing. Elements that create corrosive issues must be well understood and accounted for. Insulation design of coastal facilities should account for chloride attack. • Galvanic corrosion: Contacts between two different types of metal must be considered for galvanic corrosion potential. Similarly, water can act as an electrolyte, and galvanic corrosion can occur because of the different potential of the pipe or vessel and the metal jacketing. • Crevice/pitting corrosion: Water trapped against the interior surface of a metal weather barrier/jacket can lead to pitting/crevicetype corrosion on the interior surface of the jacket (Young 2011). • Insulation corrosivity: Some insulation materials can cause metal jacket corrosion or chemically attack some polymer films. Both of these situations shorten service life. • Thermal degradation: Hot systems are typically designed so that the surface temperature of the insulation and jacketing material do not exceed 60°C. The long-term effect of 60°C on the jacketing material must be considered. Additionally, there may be solar radiation load and perhaps parallel heat loss from an adjacent pipe. Turner and Malloy (1981) suggest that 120°C should be considered as the long-term operating temperature of the jacketing material selected. This is a critical design consideration, particularly for a nonmetal jacket. • Installation and application logistics: Often, the insulation contractor installs more insulation in a day than can be protected with jacket. If it rains, the exposed insulation gets saturated and, the next day, the jacket is installed over the wet insulation. This creates an obvious potential corrosion and performance issue before the installation is operational, and must be corrected immediately. It should also be understood that the size, shape, and adjacent space available for work may dictate the type of weather barrier used, even if it is a less desirable option. If this is the case, the maintenance schedule must recognize and accommodate for this. • Maintenance: The importance of a maintenance and inspection plan cannot be overemphasized to achieve the service life expected of the design.
23.11 Materials Used as Weather Barriers for Insulation. Metal rolls or sheets of various thicknesses are available with embossing, corrugation, moisture barriers, and different banding and closure methods. Elbows and tees are also available for piping. Typical metal jacketing materials are • • • • • • •
Bare aluminum Polymer-film-coated aluminum Painted aluminum Stainless steel Painted steel Galvanized steel Aluminum-zinc coated steel
All metal weather barrier/jacketing should have a 0.76 mm thick multiple-layer moisture barrier factory heat laminated to the interior surface to help prevent galvanic and pitting/crevice corrosion on the interior surface of the jacketing (Young 2011). Polymeric (plastic) rolls or sheets are available at various thicknesses. These materials are glued or solvent-welded, depending on the polymer. Elbows and tees are also available for piping for some type of polymers. Typical polymeric (plastic) jacketing materials include • • • •
Polyvinyl chloride (PVC) Polyvinyliedene chloride (PVDC) Polyisobutylene Multiple-layer composite materials (e.g., polymeric/foil/mesh laminates) • Fabrics (silicone-impregnated fiberglass) Numerous mastics are available. Mastics are often used with fiber-glass cloth or canvas to encapsulate pipes, tanks, or other vessels; they are also used at insulation terminations and at or around protrusions such as valves or supports. It is important to choose the correct mastic for the application, considering surface temperature, insulation type, fire hazard classification, water resistance, and vapor permeability requirements. Mastics are brushed, troweled, or sprayed on the surface at a thickness recommended by the manufacturer. Importance of Workmanship and Knowledge. Workmanship and knowledge are key to successful insulation weather barrier design. The importance of working with the installing contractor and material manufacturers regarding fitness for use of each material is paramount. The Midwest Insulation Contractors Association (MICA) publishes an excellent resource regarding materials used as weather barriers: the National Commercial and Industrial Insulation Manual (2011), in print and as a PDF file, is available from www.micainsulation.org/standards-manual.html.
Vapor Retarders Water vapor control is extremely important for piping and equipment operating below ambient temperatures. These systems are typically insulated to prevent surface condensation and control heat gain. Piping and equipment typically create an absolute barrier to passage of water vapor, so any vapor-pressure difference imposed across the insulation system results in the potential for condensation at the cold surface. A high-quality vapor retarder material or system is essential for these systems to perform adequately. Mumaw (2001) showed that the design, installation, and performance of the vapor retarder systems are key to an insulation system’s ability to minimize water vapor ingress. This research also suggests that in-place system vapor permeance may be greater than the rated performance based on standard material test methods. Moisture-related problems include thermal performance loss, health and safety issues, structural degradation, corrosion, and aesthetic issues. Water may enter the insulation system through water vapor diffusion, air leakage carrying water vapor, and leakage of surface water. A vapor retarder is a material or system that
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23.12 adequately reduces the transmission of water vapor through the insulation system. The vapor retarder system is seldom intended to resist the entry of surface water or prevent air leakage, but can occasionally be considered the second line of defense for these moisture sources. The performance of the vapor retarder material or system is characterized by its water vapor permeance. Chapter 25 has a thorough description of the physics associated with water vapor transport. Water vapor permeance can be evaluated per ASTM Standard E96, using either procedure A (desiccant method) or procedure B (water method), by imposing a vapor pressure difference across vapor retarder material that has been sealed to a test cup, and gravimetrically measuring the moisture gain or loss. It is recommended that both tests be performed and the results of both be evaluated. Faulty application can impair vapor retarder performance. The effectiveness of installation and application techniques must be considered when selecting a vapor retarder system. Factors such as vapor retarder structure, number of joints, mastics and adhesives that are used, and inspection procedures affect performance and durability. The insulation system should be dry before application of a vapor retarder to prevent trapping water vapor in the system. The system must be protected from undue weather exposure that could introduce moisture into the insulation before the system is sealed. When selecting a vapor retarder, the vapor pressure difference across the insulation system should be considered. Higher vapor pressure differences typically require a lower-permeance vapor retarder to control water vapor intrusion into the system. Service conditions affect the direction and magnitude of the vapor pressure difference. Unidirectional flow exists when water vapor pressure is constantly higher on one side of insulation system. Typically, in buildings located in humid climates, vapor pressure differences in unconditioned spaces are significantly greater than in conditioned spaces. For example, in a conditioned space at 24°C and 50% rh, the vapor pressure difference across the insulation system, on 5.6°C chilled-water lines, is less than 0.68 kPa. For belowambient, insulated pipes running in humid unconditioned spaces, the vapor pressure difference across the insulation system can be as great as 2.7 kPa in the continental United States, and even greater in places with tropical climates. Hence, for below-ambient pipes running in humid unconditioned spaces, the vapor retarder may require a lower vapor permeance than for those same pipes running in conditioned spaces. Reversible flow exists when vapor pressure may be higher on either side, typically caused by diurnal or seasonal changes on one side of the system. Properties of insulating materials used should be considered. All materials reduce water vapor flow, but low-permeance insulations can add to the overall water vapor transport resistance of the insulation system. Some low-permeability materials are considered to be vapor retarders without any additional jacket material. Vapor Retarder Jackets. There is some inconsistency in the nomenclature used for materials used as vapor retarders for pipe, tank, and equipment insulation. Designations such as jacket, jacketing, facing, and all-service jacket (ASJ ) are all applied to this component, sometimes interchangeably. On the other hand, the vapor retarder component of insulation for air-handling systems, such as duct wrap and duct board, is typically referred to only as facing. The term vapor diffusion retarder (VDR) is also used to generically describe these materials. In this chapter, vapor retarder denotes the vapor-retarding membrane of the system, but the reader should be aware of the various terms that may be encountered. In addition, some insulation materials are considered vapor retarders in themselves without any additional retarding membrane. Vapor Retarders for Pipe, Tank, and Equipment Insulation. Materials or combinations of materials used for vapor retarders can take many different forms. Necessarily, one component must be a
2017 ASHRAE Handbook—Fundamentals (SI) material that offers significant resistance to vapor passage. A commonly used preformed material for pipe, tank, and equipment vapor retarder applications is laminated white paper, reinforcing glassfiber scrim, and aluminum foil or metallized polyester film. These products are generally referred to as all-service jackets (ASJs), and meet the requirements of ASTM Test Method C1136 with a vapor permeance of 1.15 ng/(s·m2 ·Pa), per procedure A of ASTM Standard E96; C1136 is the accepted industry vapor retarder material standard for mechanical insulation applications. These facings are commonly used as the outer finish in low-abuse indoor areas; elsewhere, they should be covered by a protective jacket. Many types of insulation are supplied with factory-applied ASJ vapor retarders. Note that ASJs with exposed paper may have service limitations on below-ambient systems in wet environments, particularly in unconditioned spaces in hot, humid climates. In such spaces, during periods of high humidity, expect condensation to occur on the insulation’s surface some of the time (e.g., when relative humidity exceeds 90%). Condensation on the surface can degrade the ASJ by wetting the exposed kraft paper surface, leading to mold growth on the paper, degradation of the paper itself, and/or corrosion of the aluminum vapor retarder component. In addition to traditional ASJ vapor retarders, low-permeance monolayer plastic film and sheet, ASJ without exposed paper, laminates using aluminum foil, and other types of sheet structures are used in low- and very-low-temperature applications. They usually are water resistant; that is, when condensation occurs on their surface, they do not absorb the water. These are not always referred to as ASJ, and may be either factory applied by the insulation manufacturer or procured separately from the insulation and applied in fabrication shops or in the field. Examples of laminates include 3- to 13-ply sheet materials with thicknesses up to about 0.4 mm; these plies include at least one layer of aluminum foil and many have a permeance < 0.3 ng/(s·m2 ·Pa). There are also rubber and asphalt membranes, with an aluminum facing, with the same permeance. An example of one of the plastic films is polyvinylidene chloride (PVDC). This is typically used in more demanding applications, and often is covered by protective jacket. Many of these ASJ facings meet the requirements of ASTM Standard C1136, Type VII or VIII, or ASTM Standard C921, and generally have a permeance < 1.15 ng/(s·m2 ·Pa) per ASTM Standard E96, procedure A. The moisture-sensitive nature of paper and the relative frailty of uncoated aluminum foil can be problematic in the potentially highhumidity environment of unconditioned spaces with below-ambient applications. Exposure to water, either from condensation caused by inadequate insulation thickness or from ambient sources, can cause degradation and distortion of the paper, higher likelihood of mold growth, and foil corrosion, leading to vapor retarder failure. The presence of leachable chloride can promote corrosion of the foil or metallized film. The trend in vapor retarders for pipe insulation is toward structures without exposed paper, such as plastic films, coated metallized films, and better-protected foil laminates. Many have water vapor permeance ratings < 0.3 ng/(s·m2 ·Pa). For most common vapor retarder jackets, matching pressuresensitive tapes are available for making joint and puncture seals. With careful installation, these can be used to effectively seal joints in the vapor retarder system. In addition, vapor retarder mastics are available; these should be designated as such by their manufacturer and have a low vapor permeance rating, no greater than that of the vapor retarder membrane. Applying mastic thickly enough is critical to providing sufficient permeance. Vapor retarder mastics are typically used where fittings, supports, and other obstructions make a proper vapor seal difficult to achieve. Highly conformable tapes are sometimes used for this purpose, as well. Applied mastic systems are a vapor-retarding layer; they are often called vapor barriers by manufacturers, but their vapor resistance is a function of their permeability, thickness, and quality of the mastic application. Also,
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Insulation for Mechanical Systems some mastics may not be compatible with certain insulation types. For this reason, always consult the insulation manufacturer for recommendations on the correct type of vapor retarder to use in the application. Weather barrier mastics are not vapor retarder mastics and should not be used for below-ambient applications unless they also have a low vapor permeability or are used in conjunction with a separate vapor retarder. Below-ambient piping and equipment in general, and belowfreezing applications in particular, are the most demanding applications for an insulation vapor retarder. Even though extremely low-permeance [<0.3 ng/(s·m2 ·Pa)] vapor retarder materials exist, it is extremely difficult to achieve a perfect barrier in a system that is field installed and includes numerous joints and penetrations. It follows that adequate system design, proper insulation and jacketing material selection, and careful workmanship are all equally important. For pipes operating at below-ambient temperatures, it is recommended that every 4.5 to 6 lineal metres, or at every fitting, a vapor stop (also called vapor dam) be installed. Should a leak occur in the vapor retarder, a vapor stop isolates vapor intrusion to that pipe insulation section and thereby prevents vapor and condensed water intrusion into the adjacent section(s) of pipe insulation or adjacent fitting insulation. A vapor stop is made by applying a vapor retarder mastic liberally to the pipe surface, for 75 mm along its length, adjacent to the end of the pipe insulation section. After installing that pipe insulation section, the mastic is then applied liberally to the end of the pipe insulation. Using a glass fiber or polyester scrim allows visual confirmation that the mastic is thick enough. For illustrations of vapor stops, see MICA’s (2011) National Commercial and Industrial Insulation Standards. Air-Handling Systems. Vapor retarders for equipment and duct insulation take various forms. Because of the relatively less severe and demanding conditions in air-handling systems located in conditioned spaces (because of their higher operating temperatures and lower indoor ambient humidity), current vapor retarder materials have been shown to adequately meet these performance requirements. In general, moisture problems are not often encountered if insulation design is adequate for the application, and some lowpermeability insulation materials are used without separate vapor retarders. For fiberglass duct wrap and duct board, a lamination of aluminum foil, scrim, and kraft paper (FSK) has long been the material of choice, although flexible vinyl and other white or black facings are occasionally used. All of these facings can be procured separately in roll form, and used on any type of insulation. ASTM Standard C1136, type II, is a typical specification for factoryapplied vapor retarder on duct insulation (except flexible duct). Flexible (flex) ducting typically incorporates a plastic film or film lamination that contains a metallized substrate as a vapor-retarding component. For outdoor ducts, laminate jacketing, manufacturerrated to have a low vapor permeance [<0.3 ng/(s·m2 ·Pa)] and for outdoor use, can be installed over the previously mentioned types of duct insulation using a compatible tape for closures, to provide protection from both weather exposure and vapor intrusion to the duct insulation. Application-specific pressure-sensitive tapes or mastics are typically used to seal joints. As in any cold system where a vapor retarder is required, design, selection of materials, and workmanship must be properly addressed. The insulation manufacturer’s recommendations should be followed.
3.
INSTALLATION
Pipe Insulation Small pipes can be insulated with cylindrical half-sections of rigid insulation or with preformed flexible material. Larger pipes can be insulated with flexible material or with curved, flat segmented, or
23.13 cylindrical half, third, or quarter sections of rigid insulation. Fittings (valves, tees, crosses, and elbows) use preformed fitting insulation, fabricated fitting insulation, individual pieces cut from sectional straight-pipe insulation, or insulating cements. Fitting insulations should always be equal in thermal performance to the pipe insulation. Securing Methods. The method of securing varies with the type of insulation, size of pipe, form and weight of insulation, and type of jacketing (i.e., field- or factory-applied). Insulation with factoryapplied jacketing can be secured on small piping by securing the overlapping jacket, which usually includes an integral sealing tape. Additional tape around the circumference may be necessary. Large piping may require supplemental wiring or banding. Insulation on large piping requiring separate jacketing is wired or banded in place, and the jacket is cemented, wired, or banded, depending on the type. Flexible closed-cell materials require no jacket for most applications and are applied using specially formulated contact adhesives. Insulating Pipe Hangers. All piping is held in place by hangers and supports. Selection and treatment of pipe hangers and supports can significantly affect thermal performance of an insulation system. Thus, it is important that the piping engineer and insulation specifier coordinate during project design to ensure that correct hangers are used and sufficient physical space is maintained to allow for the required thickness of insulation. Table 10
Minimum Saddle Lengths for Use with Fibrous Glass Pipe Insulation* Minimum Saddle Length, mm
Insulation Pipe Size, Thickness, NPS 1.2 mm
Hanger Spacing, m 1.5
1.8
2.1
25
13 25 38 50 75
125 125 150 200 75 75 125 125 75 75 125 125 75 75 75 75 75 75 75 75
50
13 25 38 50 75
150 125 125 125 125
80
13 25 38 50 75
200 125 125 125 125
2.4
2.7
3.0
200 150 150 125 125
275 200 200 150 150
275 225 200 150 150
300 275 225 200 150
350 275 225 200 200
300 275 225 225 225
375 300 275 225 225
425 375 300 275 275
500 425 350 300 300
525 450 375 350 350
3.3
3.7
600 500 425 350 350
650 575 450 375 375
*For pipe sizes above 80 mm, use high-density inserts to support the pipe.
Table 11 Minimum Saddle Lengths for Use with 32 kg/m3 Polyisocyanurate Foam Insulation (13 to 75 mm thick) Minimum Saddle Length,* mm Hanger Spacing, m
Pipe Size, NPS
1.2
1.5
1.8
2.1
2.4
2.7
3.0
3.3
3.7
80 100 150 200 250 300 400
100 100 150 200 200 200 300
100 100 150 200 200 200 300
100 100 150 200 200 300 450
100 150 150 200 300 300 450
100 150 200 200 300 300 450
100 150 200 200 300 300 450
100 200 200 200 300 300 450
100 200 200 200 300 300 450
150 200 200 300 300 300 450
*22 gage hung; 20 gage setting.
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23.14
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 6
Insulating Pipe Hangers
A typical ring or line size hanger is illustrated in Figure 6A. This type of hanger is commonly used on above-ambient lines at moderate temperature. However, it provides a thermal short circuit through the insulation, and the penetration is difficult to seal effectively against water vapor, so it is not recommended for belowambient applications. Pipe shoes (Figure 6B) are used for hot piping of large diameter (high mass) and where significant pipe movement is expected. The design allows for pipe movement without damage to the insulation or the finish. The design is not recommended for below-ambient applications because of the thermal short circuit and difficulty in vapor sealing. A better solution is to use clevis hangers (Figure 6C), which are sized to allow clearance for the specified thickness of insulation, and avoid the short circuit associated with ring hangers and pipe shoes. Shields (or saddles) spread the load from the pipe, its contents, and the insulation material over an area sufficient to support the system without significantly compressing the insulation material. Table 10 provides guidance on sheet metal saddle lengths for glass fiber pipe insulation. For pipe sizes above 80 mm, it is recommended that highcompressive-strength inserts (e.g., foam, high-density fiberglass, calcium silicate) be used. Table 11 gives recommended saddle lengths for 32 kg/m3 polyisocyanurate foam insulation. Preinsulated saddles are available. Note that wood blocks have poor thermal conductivity and are not recommended, especially for cold pipe systems. When the goal is avoiding compression of low-compressivestrength insulation products, it is recommended to use high-strength insulation inserts, made of a product that offers the desired compressive strength and other necessary performance properties. Other higher-strength materials that are not thermal insulation material and interrupt the insulation envelope, or do not allow complete sealing of an insulation system against water vapor ingress, are not recommended for supporting insulated piping on pipe hangers. Insulation Finish for Above-Ambient Temperatures. Requirements for pipe insulation finishes for above-ambient applications are usually governed by location. Appearance and durability are the primary design considerations for indoor applications. For outdoor applications, finishes are provided primarily for weather protection. The finishes may be factory-applied jackets or field-applied metal or polymeric jackets. On indoor steam and hot-water distribution piping, it is common for flange pairs and flanged fittings such as gate valves, butterfly valves, strainers, pressure-relief valves, and other pipe fittings to be left bare (i.e., uninsulated). Whether this is done intentionally or by neglect, the end result is the same: enormous quantities of wasted thermal energy and, in unventilated spaces, very high air temperatures (Hart 2011). It is recommended that all these fittings be insulated with conventional pipe insulation or with removable/reusable insulation blankets, because a bare 175°C steam pipe indoors loses about eight times more heat than does the same pipe with only
25 mm of conventional pipe insulation. If maintenance personnel need ready access to these flanged fittings, removable/reusable insulation blankets are preferable. ASTM Standard C1695-10 includes different requirements for indoor and outdoor applications of removable/reusable insulation blankets. Many blankets used indoors are secured in place using hookand-loop fastening tape, which allows personnel to remove the insulation, perform maintenance, then reinstall the insulation blankets without tools. Removable/reusable insulation blankets are available either as custom-made components or as kits. If specifying custom-made blankets, time must be allowed in the schedule to have a technical support person measure for the insulation, design and fabricate the blankets, deliver them, and install them. Although ASHRAE Standard 90.1-2010 specifies 113 to 125 mm thick insulation on pipes operating above 175°C, it is recommended that permission be sought to use removable/reusable insulation blankets that are 50 mm thick or less for easier removability and reinstallation. Insulation Finish for Below-Ambient Temperatures. Piping at temperatures below ambient is insulated to limit heat gain and prevent condensation of moisture from the ambient air. Because metallic piping is an absolute barrier to water vapor, it becomes the condensing surface. Therefore, for high-permeability insulation materials [i.e., permeability/thickness combination > 1.15 ng/ (s·m2 ·Pa)], the outer surface of the insulation should be covered by a low-permeance membrane. However, some flexible closed-cell insulation materials have been used successfully without a separate vapor retarder material. Vapor retarders for straight pipe insulation are generally designed to meet permeance, operating temperature, fire safety, and appearance requirements. Sheet-type vapor retarders used on belowambient pipe insulation should have a maximum permeance of 1.15 ng/(s·m2 ·Pa), when tested per ASTM Standard E96, procedure A (desiccant method) or B (water method). Insulation materials that meet the permeance requirements of an application can be installed without separate vapor retarders, relying on the low permeability and thickness of the insulation material to resist vapor flow, but must be carefully sealed or cemented at all joints to avoid gaps in the insulation. Jacketing used as a vapor retarder may use various materials, alone or in combination, such as paper, aluminum foil, vacuummetallized or low-permeance plastic films, and reinforcing. The most commonly used products have been ASJ (white), foil-scrimkraft (FSK), metallized polyester film, or plastic sheeting. An important feature of such jacketing is very low permeance in a relatively thin layer, which provides flexibility for ease of cementing and sealing laps and end-joint strips. This type of jacketing is commonly used indoors without additional treatment. In some cases of operating temperatures below –18°C, multilayer insulation and jacketing may be used. With long lines of piping, insulation should be sealed off every 5 or 6 m with vapor stops to limit water penetration if vapor retarder damage occurs (see Figure 7). Insulation fittings are usually vapor-sealed by applying suitable materials in the field, and may vary with the type of insulation and operating temperature. The vapor seal can be a lapped spiral wrap of plastic film adhesive tape or a relatively thin coat of vapor-seal mastic. If these do not provide the required permeance, a common practice is to double-wrap a very-low-permeance plastic film adhesive tape or apply two coats of vapor-seal mastic reinforced with openweave glass or other fabric. Vapor retarder mastics should be applied in a greater thickness to achieve a lower permeance. Consult the mastic manufacturer for recommendations on the thickness required to achieve a particular permeance. Insulated cold piping should receive special attention when exposed to ambient or unconditioned air. Because cold piping frequently operates year-round, a unidirectional vapor drive may exist.
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Insulation for Mechanical Systems Even with vapor-retarding insulation, jackets, and vapor sealing of joints and fittings, moisture inevitably accumulates in permeable insulations. This not only reduces the thermal resistance of the insulation, it also accelerates condensation on the jacket surface with consequent dripping of water and possible growth of mold and mildew. Depending on local conditions, these problems can arise in less than 3 years, or as many as 30 years. Periodic insulation replacement should be considered, and the piping installation should be accessible for such replacement. Very-low-permeability insulating materials, sometimes in combination with vapor retarders, can be used to extend system life and reduce replacement frequency. This is normally done by using vapor retarders, insulation materials (e.g., cellular glass insulation), or both, with a permeance no more than 1.15 ng/(s·m·Pa). However, some flexible closed-cell insulation materials have been used successfully without a separate vapor retarder material. The lower the insulation system’s permeance, the longer its life, given proper installation. An alternative approach is to accept the inevitable water vapor ingress, and to provide a means of removing condensed water from the system. One means of accomplishing this is the use of a hydrophilic wicking material to remove condensed water from the surface of cold piping for transport (via the combination of capillary forces and gravity) outside the system where it can be evaporated to the ambient air (Brower 2000; Crall 2002; Korsgaard 1993). The wick keeps the hydrophobic insulation dry, allowing the thermal insulation to perform effectively. Dripping is avoided if ambient conditions allow evaporation. The concept is limited to pipe temperatures above freezing. For dual-temperature service, where pipe operating temperatures cycle, the vapor-seal finish, including mastics, must withstand pipe movement and exposure temperatures without deterioration. When flexible closed-cell insulation is used, it should be applied slightly compressed to prevent it from being strained when the piping expands. Outdoor pipe insulation may be vapor-sealed in the same manner as indoor piping, by applying added weather protection jacketing without damage to the vapor retarder and sealing it to keep out water. In some instances, heavy-duty weather and vapor-seal finish may be used. It is recommended that weather protection jackets installed over vapor retarders use bands or some other closure method that does not penetrate the weather protection jacketing, because other types have a high likelihood of also penetrating the underlying vapor retarder. Underground Pipe Insulation. Both heated and cooled underground piping systems are insulated. Protecting underground insulated piping is more difficult than protecting aboveground piping. Groundwater conditions, including chemical or electrolytic contributions by the soil and the existence of water pressure, require special design to protect insulated pipes from corrosion and maintain insulation thermal integrity. For optimal performance, walk-through tunnels, conduits, or integral protective coverings are generally provided to protect the pipe and insulation from water. Examples and general design features of conduits, tunnels, and direct-burial systems can be found in Chapter 12 of the 2016 ASHRAE Handbook— HVAC Systems and Equipment.
Tanks, Vessels, and Equipment Flat, curved, and irregular surfaces, such as tanks, vessels, boilers, and chimney connectors, are normally insulated with flexible or semirigid sheets or boards or rigid insulation blocks fabricated to fit the specific application. Tank and vessel head segments must be curved or flat cut to fit in single piece or segments per ASTM Standard C450. Head segments must be cut to eliminate voids at the head section, and in a minimum number of pieces to minimize joints. Prefabricated flat head sections should be installed in the same number of layers and thickness as the vessel walls, and void
23.15 areas behind the flat head should be filled with packable insulation. Typically, the curved segments are fabricated to fit the contour of the vessel surface in equal pieces to go around the vessel with a minimum number of joints. Because no general procedure can apply to all materials and conditions, manufacturers’ specifications and instructions must be followed for specific applications. Securing Methods. Insulations are secured in various ways, depending on the form of insulation and contour of the surface to be insulated. Flexible insulations are adhered to tanks, vessels, and equipment using contact adhesive, pressure-sensitive adhesives, or other systems recommended by the manufacturers. The insulation’s flexibility lends itself to curved surfaces. Rigid or semirigid insulations on small-diameter, cylindrical vessels can be prefabricated and adhered or mechanically attached, as appropriate. On larger cylindrical vessels, angle iron ledges to support the insulation against slippage can supplement banding. Where diameters exceed 3 to 4 m, slotted angle iron may be run lengthwise on the cylinder, at intervals around the circumference to secure and avoid an excessive length of banding. Rigid and semirigid insulations can be secured on large flat and cylindrical surfaces by banding or wiring and can be supplemented by fastening with various welded studs at frequent intervals. Springs may be necessary on banding to allow for expansion/contraction of the tank and insulation. On large flat, cylindrical, and spherical surfaces, it is often advantageous to secure the insulation by impaling it on welded studs or pins and fastening it with speed washers. Flexible closed-cell insulations are adhered directly to these surfaces, using a suitable contact adhesive. Insulation Finish. Insulation finish is often required to protect insulation against mechanical damage and weather, consistent with acceptable appearance. On smaller indoor equipment, insulation is commonly covered with tightly stretched and secured hexagonal wire mesh. Then, a base and hard finish coat of cement is applied, and sometimes painted. For the same equipment outdoors, insulation can be finished with a coat of hard cement, properly secured hexagonal mesh, and a coat of weather-resistant mastic. A variation is to apply only two coats of weather-resistant mastic reinforced with open-mesh glass or other fabric; however, this finish is limited to an operating temperature of about 150°C, because metal expansion can rupture the finish at insulation joints. Larger equipment may be finished indoors and out with suitable sheet metal. Outdoor finish is generally metal jacketing with a 0.076 mm thick multilayer moisture barrier factory heat laminated to the interior surface, properly flashed around penetrations (e.g., access openings, pipe connections, and structural supports) to maintain weathertightness. Various outdoor finishes are available for different types of insulations. For below-ambient operating temperatures, insulation is finished, as required, to prevent condensation and protect against mechanical damage and weather, consistent with acceptable appearance. In accordance with the operating temperature, the finish must retard vapor to avoid moisture entry from surrounding air, which can increase the insulation’s thermal conductivity or deterioration, or corrode the metal equipment surface. Whenever a vapor retarder is required, all penetrations such as access openings, pipe connections, and structural supports must be properly treated with an appropriate vapor-retarder film, mastic, or other sealant. Equipment must be insulated from structural steel by isolating supports of high compressive strength and reasonably low thermal conductivity, such as a rigid insulation material. The vapor retarder must carry over this insulation from the equipment to the supporting steel to ensure proper sealing. If the equipment rests directly on steel supports, the supports must be insulated for some distance from the points of contact.
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23.16 Commonly, insulation and vapor retarder are extended four times the thickness of the insulation applied to the equipment. For dual-temperature service, where vessels are alternately cold and hot, vapor retarder finish materials and design must withstand movement caused by temperature change.
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Ducts Ducts are used to convey air or process gases for several purposes. In general, their uses can be divided into process ducts and HVAC ducts. Process ducts can range from industrial hot exhaust to subambient process gases, and can be outdoors or indoors. They can need insulation for various reasons, including thermal energy conservation, personnel protection, process control, condensation control and noise attenuation. Because of the wide range of possible operating temperatures and environmental conditions, selection of duct insulation and jacketing materials for industrial processes requires careful consideration of all operational, environmental, and human safety factors. Issues of energy conservation, personnel protection, condensation control, and process control can be solved by careful analysis of heat flow. Computer programs are available for calculating heat transfer, surface temperatures for personnel protection, condensation control, and economic thickness. HVAC ducts carry air to conditioned spaces inhabited by people, animals, sensitive equipment, or a combination thereof. Thermal and acoustical duct insulation is one of the keys to a well-designed system that provides both occupant comfort and acceptable indoor air quality (IAQ). These insulation products help maintain a consistent air temperature throughout the system, reduce condensation, absorb system operation noise, and conserve energy. Typical air temperatures for HVAC applications are 4 to 50°C. Because of the more moderate temperature ranges associated with HVAC applications, there is a wide range of insulation materials available. Where acoustical and thermal considerations are significant, sheet metal ducts are often internally insulated, or the ducts themselves are constructed of materials that form both the airconveying duct and the insulation. Where acoustical concerns are not significant, sheet metal ducts can be externally insulated with rigid, semirigid, or flexible insulation materials. Again, another alternative is to use ducts that incorporate insulation as part of their construction. The determining factor in duct construction and insulation materials selection is often a combination of performance criteria and budgetary limitations. The need for duct insulation is influenced by the • Duct location (e.g., indoors or outdoors; conditioned, semiconditioned, or unconditioned space) • Effect of heat loss or gain on equipment size and operating cost • Need to prevent condensation on low-temperature ducts • Need to control temperature change in long duct lengths • Need to control noise transmitted within the duct or through the duct wall All HVAC ducts exposed to outdoor conditions, as well as those passing through unconditioned or semiconditioned spaces, should be insulated. Analyses of temperature change, heat loss or gain, and other factors affecting the economics of thermal insulation are essential for large commercial and industrial projects. ASHRAE Standard 90.1 and building codes set minimum standards for thermal efficiency, but economic thickness is often greater than the minimum. Additionally, the standards and codes do not address surface condensation issues. These considerations are often the primary driver of minimum thickness in unconditioned or semiconditioned locations subject to moderate or greater relative humidity. Duct thermal efficiencies are generally regulated by local or national codes by specifying minimum thermal resistances, or R-values. These R-values are most often determined by testing per ASTM Standards C518 or C177, as required by the Federal Trade Commission (FTC) for
2017 ASHRAE Handbook—Fundamentals (SI) reporting R-values of duct wrap insulations. Neither method allows for increased thermal resistance caused by convective or radiative surface effects. To comply with current code language, it is recommended that R-value requirements in specifications for duct insulation be based on Standards C177 or C518 testing at 24°C mean temperature and at the installed thickness of the insulation. Insulation products for ducts are available in a range of R-values, dependent predominantly on insulation thickness, but also somewhat on insulation density. Temperature Control Calculations for Air Ducts. Duct heat gains or losses must be known for the calculation of supply air quantities, supply air temperatures, and coil loads (see Chapter 17 of this volume and Chapter 4 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment). Heat loss programs based on ASTM Standard C680 may be used to calculate thermal energy transfer through the duct walls. Duct air exit temperatures can then be estimated using the following equations: qPL tdrop or tgain = ------------------ VCp A then, for warm air ducts,
(3)
texit = tenter – tdrop
(4)
texit = tenter – tgain
(5)
and for cold air ducts,
where tdrop tenter tgain texit q P L V Cp A
= = = = = = = = = = =
temperature loss for warm air ducts, K entering air temperature, K temperature rise for cool air ducts, K exit temperature for either warm or cool air ducts, °C heat loss through duct wall, W/m2 duct perimeter, m length of duct run, m air velocity in duct, m/s specific heat of air, 1.001 J/(kg·K) density of air, 1.20 kg/m3 area of duct, m2
Example 1. A 20 m length of 0.6 by 0.9 m uninsulated sheet metal duct, freely suspended, conveys heated air through a space maintained at 4°C. The ASTM Standard C680 heat loss calculation gives a heat transfer rate of 444 W/m2. Based on heat loss calculations for the heated zone, 8.1 m3/s of standard air [cp = 1006 J/(kg·K)] at a supply air temperature of 50°C is required. The duct is connected directly to the heated zone. Determine the temperature of the air entering the duct. Solution: The area of the duct is 0.6 × 0.9 = 0.54 m2. Air velocity V is calculated as Volumetric flow 8.1 --------------------------------------- = ---------- = 15 m/s Area 0.54 Duct perimeter P is (0.6 + 0.9) 2 = 3 m Temperature drop tdrop is 444 3 20 --------------------------------------------------------- = 2.7 K 15 1006 1.20 0.54 Temperature of air entering the duct is thus 50 + 2.7 = 52.7°C Example 2. Repeat Example 1, except the duct is insulated externally with 25 mm thick insulation material having a heat transfer rate of 44.7 W/m2. Solution: All values except q remain the same as in the previous example. Therefore, tdrop is
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Fig. 7 R-Value Required to Prevent Condensation on Surface with Emittance = 0.1
Fig. 8 R-Value Required to Prevent Condensation on Surface with Emittance = 0.9
44.7 3 20 --------------------------------------------------------- = 0.27 K 15 1006 1.20 0.54 Temperature of air entering the duct is thus 50 + 0.3 = 50.3°C
Preventing Surface Condensation on Cool Air Ducts. Insulation also can prevent surface condensation on cool air ducts operating in warm and humid environments. This reduces the opportunity for microbial growth as well as other moisture-related building damage. Condensation forms on cold air-conditioning ducts anywhere the exterior surface temperature reaches the dew point. The moisture may remain in place or drip, causing moisture damage and creating a potential for microbial contamination. Preventing surface condensation requires that sufficient thermal resistance be installed for the condensation design conditions. Figures 7 and 8 give installed R-value requirements to prevent surface condensation on insulated air ducts. The first chart (for emittance = 0.1) should be used for foil-faced insulation products. The second chart is based on materials with a surface emittance of 0.9. The designer must choose the appropriate environmental conditions for the location. It may be financially imprudent to design for the most extreme condition that could occur during a system’s life, but it is also important to be aware that the worst-case condition for condensation control is not the maximum design load of coincident dry bulb and wet bulb. Rather, the worst case for condensation occurs when relative humidity is high, such as when the dry bulb is only very slightly above the wet bulb. This condition often occurs in the early morning in many climates. For cold-duct applications, it is critical that water vapor not be allowed to enter the insulation system and condense on the cold duct
23.17 surface. Once water begins to condense, loss of thermal efficiency is inevitable. This leads to further surface condensation on the surface of the insulation. To prevent this, exterior duct insulations must have a low vapor permeance. Permeance values of 5.75 ng/(s·m2 ·Pa) or less are generally recommended for cold-duct applications. It is equally important that all exterior duct insulation joints are well sealed. Areas of special concern are often found around duct flanges, hangers, and other fittings. Insulation Materials for HVAC Ducts. Insulated ducts in buildings can consist of insulated sheet metal, fibrous glass, or insulated flexible ducts, all of which provide combined air barrier, thermal insulation, and some degree of sound absorption. Ducts embedded in or below floor slabs may be of compressed fiber, ceramic tile, or other rigid materials. Depending on the insulation material, there are a number of standards that specify the material requirements. Duct insulations include semirigid boards and flexible blanket types, composed of organic and inorganic materials in fibrous, cellular, or bonded particle forms. Insulations for exterior surfaces may have attached vapor retarders or facings, or vapor retarders may be applied separately. When applied to the duct interior as a liner, insulation both insulates thermally and absorbs sound. Duct liner insulations have sound-permeable surfaces on the side facing the airstream capable of withstanding duct design air velocities or duct cleaning without deterioration. Abuse Resistance. One important consideration in the choice of external insulations for air ducts is abuse resistance. Some insulation materials have more abuse resistance than others, but most insulation materials will not withstand high abuse such as foot traffic. In high-traffic locations, a combination of insulation and protective jacketing materials is required. In areas where the insulation is generally inaccessible to human contact, less rigid insulations are acceptable. One important consideration for internal insulation abuse resistance is that, in large commercial units, it is common for maintenance personnel to enter and move around. In these instances, it is critical that structural elements be provided to keep foot traffic off the insulation. Duct Airstream Surface Durability. One of the most important considerations in choosing internal duct insulations is resistance to air erosion. Each material has a maximum airflow velocity rating, which should be determined using an erosion test methodology in accordance with either UL Standard 181 or ASTM Standard C1071. Under these methods, the liner material is subjected to velocities that are two and one-half times the maximum rated air velocity. Air erosion testing should include evaluation of the insulation for evidence of erosion, cracking, or delamination. Additionally, internal insulation must be resistant to aging effects in the air duct environment. Insulation materials have maximum temperatures for prolonged exposure, and some codes impose temperature requirements. Ensure that the material selected will have no aging effects at either the anticipated maximum duct temperature or the temperature specified by the code bodies for the local jurisdiction. Another durability concern is ultraviolet (UV) resistance. Ultraviolet-generating equipment is used to mitigate microbial activity. Determine, both from the UV equipment manufacturer as well as the insulation manufacturer, whether the anticipated UV exposure poses a threat to the insulation. Duct Airflow Characteristics. Internal duct insulations and ducts that have insulation as part of their construction have increased frictional pressure loss characteristics compared to bare sheet metal. Generally, duct dimensions are oversized to compensate for the increased frictional pressure losses and the decrease in internal cross section caused by the insulation thickness. However, frictional losses are only part of the total static pressure loss in a duct; dynamic fitting losses should also be considered in any required resizing. Generally,
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23.18
2017 ASHRAE Handbook—Fundamentals (SI)
internal linings conform to the shape of the fitting in which they are installed. For this reason, the insulated fitting is assumed to have the same dynamic pressure loss as the uninsulated fitting of the same dimension. See Chapter 21 for further details on frictional pressure loss characteristics of internal linings and how they affect pressure drop and duct-sizing requirements. Securing Methods. Exterior rigid or flexible duct insulation can be attached with adhesive, with supplemental preattached pins and clips, or with wiring or banding. Individual manufacturers of these materials should be consulted for their installation requirements. Flexible duct wraps do not require attachment except on bottom duct panels more than 600 mm wide. For larger ducts, pins placed at a maximum spacing of 600 mm or less are sufficient. Internal liners are attached with adhesive and pins, in accordance with industry standards. Leakage Considerations. To achieve the full thermal benefits of insulation, air ducts should be substantially sealed against leakage under operating pressure. The insulation material should not be counted on to provide leakage resistance, unless the insulation is part of the actual duct. Using the case in Example 1, 10% air leakage from an unsealed duct represents an energy loss of 1.66 times the energy lost through heat transfer through the entire 20 m of uninsulated duct. When that same 10% leakage is compared against an insulated duct, the energy lost through leakage is 15.3 times the losses through heat transfer through the insulation over the entire duct length. Outdoor Applications. Insulated air ducts located outdoors generally require specific protection against the elements, including water, ice, hail, wind, ultraviolet exposure, vermin, birds, other animals, and mechanical abuse. Strategies for protecting externally insulated ducts located outdoors include protective metal jackets and glass fabric with weather barrier mastic. Note that most of these protective weather treatments do not replace the need for a vapor retarder for cold-duct applications. Special Considerations. Cooling-only ducts in cold climates. These applications generally occur in northern climates with ducts that run through unconditioned spaces such as attics. Warm air from the conditioned space enters through registers and into the unused ducts. This relatively moist air then condenses in the cold portions of the duct. Condensation encourages odor problems, mold growth, and degradation of the insulation. In the worst cases, water build-up becomes so excessive that water-soaked or frozen ducts can collapse and break through ceilings. The solution is to completely seal all entry points into the ducts, generally by sealing behind all registers, using a very good vapor retarder such as 0.2 mm polyethylene sheet. Covering ducts with insulation in attics in hot and humid climates. In an effort to conserve energy, attic insulation levels have been increasing. Often, these attics are insulated with pneumatically applied loose fill insulation. If this insulation comes into contact with duct insulation, it could lower surface temperatures on the facing of the duct insulation below dew point during humid conditions. For this reason, it is important that the ducts be supported so that they are above the attic insulation. Many building codes in humid areas (e.g., Florida) require this, and it should be considered good practice in all humid climates.
4.
Q = –kA dT/dx where Q A k dT/dx
Fundamentals of heat transfer are covered in Chapter 4, and the concepts are extended to insulation systems in Chapter 25. Steadystate, one-dimensional heat flow through insulation systems is governed by Fourier’s law:
= = = =
rate of heat flow, W cross-sectional area normal to heat flow, m2 thermal conductivity of insulation material, W/(m·K) temperature gradient, °C/m
For flat geometry of finite thickness, the equation reduces to Q = kA(T1 – T2)/L
(7)
where L is insulation thickness, in m. For radial geometry, the equation becomes Q = kA2(T1 – T2)/[r2 ln(r2/r1)]
(8)
where r2 = outer radius, m r1 = inner radius, m A2 = area of outer surface, m2
The term r2 ln(r2 /r1 ) is sometimes called the equivalent thickness of the insulation layer. Equivalent thickness is the thickness of insulation that, if installed on a flat surface, would equal the heat flux at the outer surface of the cylindrical geometry. Heat transfer from surfaces is a combination of convection and radiation. Usually, these modes are assumed to be additive, and therefore a combined surface coefficient can be used to estimate the heat flow to and from a surface: hs = hc + hr
(9)
where hs = combined surface coefficient, W/(m2 ·K) hc = convection coefficient, W/(m2 ·K) hr = radiation coefficient, W/(m2 ·K)
Assuming the radiant environment is equal to the ambient air temperature, the heat loss/gain at a surface can be calculated as Q = hs A(Tsurf – Tamb)
(10)
The radiation coefficient is usually estimated as 4
4
hr = T surf – T amb /(Tsurf – Tamb)
(11)
where = surface emittance = Stephen-Boltzmann constant, 5.67 × 10–8 W/(m2 ·K4)
Table 12 gives the approximate emittance of commonly used materials.
Controlling Surface Temperatures A common calculation associated with mechanical insulation systems involves determining the thickness of insulation required to control the surface temperature to a certain value given the operating temperature of the process and the ambient temperature. For example, it may be desired to calculate the thickness of tank insulation required to keep the outer surface temperature at or below 60°C when the fluid in the tank is 232°C and the ambient temperature is 27°C. At steady state, the heat flow through the insulation to the outer surface equals heat flow from the surface to the ambient air:
DESIGN DATA
Estimating Heat Loss and Gain
(6)
Qins = Qsurf
(12)
(k/X)A(Thot – Tsurf ) = hA(Tsurf – Tamb)
(13)
or Rearranging this equation yields X = (k/h)[(Thot – Tsurf )/(Tsurf – Tamb)]
(14)
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Because the ratio of temperature differences is known, the required thickness can be calculated by multiplying the temperature difference and the ratio of the insulation material conductivity to the surface coefficient. For this example, assume the surface coefficient can be estimated as 5.7 W/(m2 ·K), and the conductivity of the insulation to be used is 0.036 W/(m·K). The required thickness can then be estimated as
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0.036 230 – 60 X = ------------- ------------------------- = 0.032 m or 32 mm 5.7 60 – 27
(15)
This estimated thickness would be rounded up to the next available size, probably 38 mm. For radial heat flow, the thickness calculated represents the equivalent thickness; the actual thickness (r2 – r1) is less, per Equation (8). This simple procedure can be used as a first-order estimate. In reality, the surface coefficient is not constant, but varies as a function of surface temperature, air velocity, orientation, and surface emittance. When performing these calculations, it is important to use the actual dimensions for the pipe and tubing insulation. Many (but not all) pipe and tubing insulation products conform to dimensional standards originally published by the U.S. Navy in Military Standard MIL-I-2781 and since adopted by other organizations, including ASTM. Standard pipe and insulation dimensions are given for reference in Table 13, and standard tubing and insulation dimensions in Table 14. Corresponding dimensional data for flexible closed-cell insulations are given in Tables 15 and 16. For mechanical insulation systems, it is also important to realize that the thermal conductivity k of most insulation products varies significantly with temperature. Manufacturer’s literature usually provides curves or tabulations of conductivity versus temperature. When performing heat transfer calculations, it is important to use the effective thermal conductivity, which can be obtained by integration of the conductivity versus temperature curve, or (as an approximation) using the conductivity evaluated at the mean temperature across the insulation layer. ASTM Standard C680 provides the algorithms and calculation methodologies for incorporating these equations in computer programs. These complications are readily handled for a variety of boundary conditions using available computer programs, such as the NAIMA 3E Plus® program [available as a download from the website of the North American Insulation Manufacturers Association (NAIMA), www.naima.org]. Estimates of the heat loss from bare pipe and tubing are given in Tables 17 and 18. These are useful for quickly estimating the cost of lost energy from uninsulated piping.
5.
PROJECT SPECIFICATIONS
Table 12 Emittance Data of Commonly Used Materials All-service jacket (ASJ) Aluminum paint Aluminum Anodized Commercial sheet Embossed Oxidized Polished Aluminum-zinc coated steel Canvas Colored mastic Copper Highly polished Oxidized Elastomeric or polyisobutylene Galvanized steel Dipped or dull New, bright Iron or steel Painted metal Plastic pipe or jacket (PVC, PVDC, or PET) Roofing felt and black mastic Rubber Silicon-impregnated fiberglass fabric Stainless steel New, cleaned Oxidized in service
Although each of these steps is important, identifying systems that must be insulated is critical. When defining the materials to be used, it is important to specify them exactly, while allowing for submission of generic equivalents or value-engineered materials. Overspecifying materials limits potentially good options, and underspecifying may allow underperforming materials to be used.
0.9 0.5 0.8 0.1 0.2 0.1 to 0.2 0.04 0.06 0.7 to 0.9 0.9 0.03 0.8 0.9 0.3 0.1 0.8 0.8 0.9 0.9 0.9 0.9 0.2 0.32
A proper balance is needed, and specifying key properties is required. Specifying installation requirements is as important as specifying the correct materials. Specifying procedures for submittals and quality control at the job ensures correctness. Reference standards from ASHRAE, ASTM, MICA, and others should be incorporated into the specifications; this practice saves time with respect to specification development, and shortens the specification considerably. Specifications that are not reviewed and updated periodically can perpetuate old technologies and obsolete materials, and fall out of code compliance. Essentially all manufacturers of mechanical insulation products offer guide specifications for their products. These documents are insightful and offer credible information about specific products and the accessories commonly used with them. These documents are widely available on manufacturers’ websites.
The importance of a well-prepared specification to meet energy conservation objectives is paramount. Specifications should • Identify systems and equipment that must be insulated • Identify precisely the materials selected, including thickness and jacketing, etc. • Define the procedure for submitting alternative materials and systems • Specify installation, inspection, and repair requirements • Describe procedures to ensure the job is done correctly • Comply with regional and national building codes
Emittance at ~27°C
Material
STANDARDS ANSI/ASHRAE Standard 90.2
Energy-Efficient Design of Low-Rise Residential Buildings
ANSI/ASHRAE/IES Standard 90.1 Energy Standard for Buildings Except LowRise Residential Buildings ASTM Standard 165 C177
Test Method for Measuring Compressive Properties of Thermal Insulations Test Method for Steady-State Heat Flux Measurements and Thermal Transmission Properties by Means of the Guarded-Hot-Plate Apparatus
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Table 13
Inner and Outer Diameters of Standard Pipe Insulation Insulation Nominal Thickness, mm
Pipe Size, mm
Pipe OD, mm.
Insulation ID, mm
25
38
50
63
75
88
100
113
125
15 20 25 32 40 50 65 80 90 100 115 125 150 175 200 225 250 275 300 350
21.3 26.7 33.4 42.2 48.3 60.3 73.0 88.9 102 114 127 141 168 194 219 244 273 298 324 356
22 27 34 43 49 61 74 90 102 115 128 143 170 196 221 246 275 300 326 358
73 73 89 89 102 114 127 141 168 168 194 194 219 — — — — — — —
102 102 114 127 127 141 168 168 194 194 219 219 244 273 298 324 356 381 406 432
127 127 141 141 168 168 194 194 219 219 244 244 273 298 324 356 381 406 432 457
168 168 168 168 194 194 219 219 244 244 273 273 298 324 356 381 406 432 457 483
194 194 194 194 219 219 244 244 273 273 298 298 324 356 381 406 432 457 483 508
219 219 219 219 244 244 273 273 298 298 324 324 356 381 406 432 457 483 508 533
244 244 244 244 273 273 298 298 324 324 356 356 381 406 432 457 483 508 533 559
273 273 273 273 298 298 324 324 356 356 381 381 406 432 457 483 508 533 559 584
298 298 298 298 324 324 356 356 356 381 381 406 432 457 483 508 533 559 584 610
Table 14 Inner and Outer Diameters of Standard Tubing Insulation Tube Size, Tube OD, Insulation mm mm ID, mm 10 15 20 25 32 40 50 65 80 90 100 125 150
12.7 15.9 22.2 28.6 34.9 41.3 54.0 66.7 79.4 92.1 105 130 156
13 16 23 29 35 42 55 68 80 93 106 131 157
Insulation Nominal Thickness, mm 25
38
50
63
75
88
100
113
125
60 73 73 73 89 89 102 114 127 141 168 194 219
89 89 102 102 114 114 127 141 168 168 194 219 244
114 114 127 127 141 141 168 168 194 194 219 244 273
141 141 168 168 168 168 194 194 219 219 244 273 298
168 168 194 194 194 194 219 219 244 244 273 298 324
— — 219 219 219 219 244 244 273 273 298 324 356
— — 244 244 244 244 273 273 298 298 324 356 381
— — 273 273 273 273 298 298 324 324 356 381 406
— — 298 298 298 298 324 324 356 356 381 406 432
Table 15 Inner and Outer Diameters of Standard Flexible Closed-Cell Pipe Insulation Insulation OD, mm Pipe Size, mm 15 20 25 32 40 50 65 80 90 100 125 150 200
Pipe OD, Insulation mm ID, mm 21.3 26.7 33.4 42.2 48.3 60.3 73.0 88.9 102 114 141 168 219
24.6 28.7 36.6 45.2 51.6 63.5 76.2 94.0 107 119 146 173 224
Insulation Nominal Thickness, mm 13
19
25
47.5 51.6 62.0 70.6 77.0 88.9 102 118 135 149 174 201 252
62.7 66.8 74.7 85.9 92.2 104 117 134 150 163 189 217 267
75.4 79.5 87.4 96.0 102 114 127 146 163 175 202 229 —
Table 16 Inner and Outer Diameters of Standard Flexible Closed-Cell Tubing Insulation Insulation OD, mm Tube Insulation Nominal Thickness, mm Nominal Tube OD, Insulation Size, mm mm ID, mm 13 19 25 10
12.7
15.2
38.1
49.5
—
15
15.9
19.1
41.9
54.0
69.9
20
22.2
25.4
49.5
63.5
76.2
25
28.6
31.8
56.4
72.4
82.6
32
34.9
38.1
63.5
78.7
88.9
40
41.3
44.5
69.9
85.1
95.3
50
54.0
57.2
82.6
97.8
108
65
66.7
69.9
95.3
110
121
80
79.4
82.6
108
123
133
90
92.1
95.3
123
138
151
100
105
108
136
151
164
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Table 17 Heat Loss from Bare Steel Pipe to Still Air at 27°C, W/m Nominal Pipe Size, mm
82
138
193
249
304
15 20 25 32 40 50 65 80 90 100 115 125 150 175 200 225 250 300 350 400 450 500 600
54.1 65.5 79.3 98.1 110 135 161 193 219 244 270 301 354 405 455 505 560 660 718 817 916 1020 1210
133 160 195 241 272 336 400 480 543 607 670 747 880 1000 1130 1260 1390 1640 1790 2040 2290 2530 3030
234 285 346 429 484 599 714 856 971 1090 1200 1340 1580 1810 2030 2250 2510 2950 3210 3660 4100 4550 5440
362 441 538 668 756 936 1120 1350 1520 1700 1880 2100 2480 2840 3190 3540 3940 4640 5060 5770 6470 7170 8570
524 640 781 971 1100 1360 1630 1960 2220 2490 2750 3070 3620 4140 4670 5190 5770 6820 7420 8450 9490 10 500 12 600
10 15 20 25 32 40 50 65 80 90 100 125 150 200 250 300
C355 C411 C423
C450
C518
C534
Pipe Inside Temperature, °C
Table 18 Heat Loss from Bare Copper Tube to Still Air at 27°C, W/m Nominal Pipe Size, mm
C533
C547 C552 C553 C578 C585 C591 C612 C680
795 C921 C1055 C1071
Pipe Inside Temperature, °C 49
66
82
99
116
10.2 12.2 16.1 19.9 23.6 27.4 34.7 42.0 49.2 56.4 63.5 77.8 91.9 120 148 176
19.8 23.7 31.4 38.9 46.4 53.7 68.3 82.7 97.1 112 125 153 181 237 291 346
30.7 36.7 48.7 60.5 72.0 83.5 106 129 151 173 195 238 283 368 455 540
42.5 51.0 67.7 84.1 100 116 148 180 210 241 272 334 394 515 635 756
55.3 66.5 88.3 110 131 152 193 235 276 316 357 436 517 676 833 991
Test Method for Steady-State Heat Transfer Properties of Horizontal Pipe Insulation Test Method for Hot-Surface Performance of High-Temperature Thermal Insulation Test Method for Sound Absorption and Sound Absorption Coefficients by the Reverberation Room Method Practice for Fabrication of Thermal Insulating Fitting Covers for NPS Piping, and Vessel Lagging Test Method for Steady-State Thermal Transmission Properties by Means of the Heat Flow Meter Apparatus
C1126 C1136 C1427 C1695 D1621 E84 E96 E119 E136 E795 E1222 E1529 E2231
Specification for Calcium Silicate Block and Pipe Thermal Insulation Specification for Preformed Flexible Elastomeric Cellular Thermal Insulation in Sheet and Tubular Form Specification for Mineral Fiber Pipe Insulation Specification for Cellular Glass Thermal Insulation Specification for Mineral Fiber Blanket Thermal Insulation for Commercial and Industrial Applications Specification for Rigid, Cellular Polystyrene Thermal Insulation Practice for Inner and Outer Diameters of Rigid Thermal Insulation for Nominal Sizes of Pipe and Tubing (NPS System) Specification for Unfaced Preformed Rigid Cellular Polyisocyanurate Thermal Insulation Specification for Mineral Fiber Block and Board Thermal Insulation Practice for Estimate of the Heat Gain or Loss and the Surface Temperatures of Insulated Flat, Cylindrical, and Spherical Systems by Use of Computer Programs Specification for Thermal Insulation for Use in Contact with Austentic Stainless Steel Practice for Determining the Properties of Jacketing Materials for Thermal Insulation Guide for Heated System Surface Conditions That Produce Contact Burn Injuries Specification for Fibrous Glass Duct Lining Insulation (Thermal and Sound Absorbing Material) Specification for Faced or Unfaced Rigid Cellular Phenolic Thermal Insulation Specification for Flexible Low Permeance Vapor Retarders for Thermal Insulation Specification for Preformed Flexible Cellular Polyolefin Thermal Insulation in Sheet and Tubular Form Standard Specification for Fabrication of Flexible Removable and Reusable Blanket Insulation for Hot Service Test Method for Compressive Properties of Rigid Cellular Plastics Test Method for Surface Burning Characteristics of Building Materials Test Methods for Water Vapor Transmission of Materials Test Methods for Fire Tests of Building Construction and Materials Test Method for Behavior of Materials in a Vertical Tube Furnace at 750°C Practice for Mounting Test Specimens during Sound Absorption Tests Test Method for Laboratory Measurement of the Insertion Loss of Pipe Lagging Systems Test Methods for Determining Effects of Large Hydrocarbon Pool Fires on Structural Members and Assemblies Practice for Specimen Preparation and Mounting of Pipe and Duct Insulation Materials to Assess Surface Burning Characteristics
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23.22 MICA 2011 NACE SP0198
NFPA Standard 90A 90B 255 259
Licensed for single user. © 2017 ASHRAE, Inc.
UL/UL Canada Standard 181 723 CAN/ULC-S101 CAN/ULC-S102 CAN4-S114
2017 ASHRAE Handbook—Fundamentals (SI)
National Commercial and Industrial Insulation Standards, 7th ed. The Control of Corrosion under Thermal Insulation and Fireproofing Materials—A Systems Approach Installation of Air-Conditioning and Ventilating Systems Installation of Warm Air Heating and AirConditioning Systems Method of Test of Surface Burning Characteristics of Building Materials Test Method for Potential Heat of Building Materials Factory-Made Air Ducts and Air Connectors Standard for Test for Surface Burning Characteristics of Building Materials Methods of Fire Endurance Tests of Building Construction and Materials Method of Test for Surface Burning Characteristics of Building Materials and Assemblies Method of Test for Determination of NonCombustibility in Building Materials (Rev 1997)
U.S. Navy Standard MIL-1-2781 Insulation, Pipe, Thermal
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by
nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. ASTM. 2001. Moisture analysis and condensation control in building envelopes. MNL 40. American Society for Testing and Materials, West Conshohocken, PA. Brower, G. 2000. A new solution for controlling water vapor problems in low temperature insulation systems. Insulation Outlook (September). Crall, G.C.P. 2002. The use of wicking technology to manage moisture in below ambient insulation systems. Insulation Materials, Testing and Applications, vol. 4, pp. 326-334, A.O. Desjarlais and R.R. Zarr, eds. ASTM STP 1426. American Society for Testing and Materials, West Conshohocken, PA. Cremaschi, L., A. Ghajar, S. Cai, and K. Worthington. 2012. Methodology to measure thermal performance of pipe insulation at below-ambient temperatures (RP-1356). ASHRAE Research Project RP-1356, Final Report. Gordon, J. 1996. An investigation into freezing and bursting water pipes in residential construction. Research Report 90-1. University of Illinois Building Research Council. Hart, G. 2006. Thermal insulation coatings (TICs): How effective are they as insulation? Insulation Outlook (July). Hart, G. 2011. Saving energy by insulating pipe components on steam and hot water distribution systems. ASHRAE Journal (October). Kalis, J. 1999. Water and insulation: A corrosive mix. Insulation Outlook (April). Korsgaard, V. 1993. Innovative concept to prevent moisture formation and icing of cold pipe insulation. ASHRAE Transactions 99(1):270-273. Kuntz, H.L., and R.M. Hoover. 1987. The interrelationship between the physical properties of fibrous duct lining materials and lined duct sound attenuation (RP-478). ASHRAE Transactions 93(2). Marion, W., and K. Urban. 1995. User’s manual for TMY2’s typical meteorological years. National Renewable Energy Laboratory, Golden, CO. Miller, W.S. 2001. Acoustical lagging systems. Insulation Outlook (April): 41-46. Mumaw, J.R. 2001. Below ambient piping insulation systems. Insulation Outlook (September). Turner, W.C., and J.F. Malloy. 1981. Thermal insulation handbook. Robert E. Kreiger Publishing, McGraw Hill Book Company, New York. WDBG. 2012. Mechanical insulation design guide—Design objectives. www.wbdg.org/design/midg_design.php. Whole Building Design Guide, National Institute of Building Sciences, Washington, D.C. Young, J. 2011. Preventing corrosion on the interior surface of metal jacketing. Insulation Outlook (November).
Related Commercial Resources
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Related Commercial Resources CHAPTER 24
AIRFLOW AROUND BUILDINGS Flow Patterns........................................................................... Wind Pressure on Buildings .................................................... Sources of Wind Data .............................................................. Wind Effects on System Operation...........................................
24.1 24.4 24.7 24.8
IRFLOW around buildings affects worker safety, process and building equipment operation, pollution infiltration at building inlets, and the ability to control indoor environmental parameters such as temperature, humidity, air motion, and contaminants. Wind causes variable surface pressures on buildings that can change intake and exhaust system flow rates, natural ventilation, infiltration and exfiltration, and interior pressures. The mean flow patterns and turbulence of wind passing over a building can also lead to recirculation of exhaust gases into air intakes. This chapter provides basic information for evaluating wind flow patterns, estimating wind pressures, and identifying problems caused by the effects of wind on pedestrians and buildings, including ventilation intakes, exhausts, and equipment. In most cases, detailed solutions are addressed in other chapters. Related information can be found in Chapters 11, 14, 16, and 37 of this volume; in Chapters 31, 32, 45, 47, and 53 of the 2015 ASHRAE Handbook—HVAC Applications; and in Chapters 30, 35, and 40 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment.
Building Pressure Balance and Internal Flow Control ......... Environmental Impacts of Building External Flow ............... Physical and Computational Modeling.................................. Symbols ..................................................................................
24.10 24.11 24.12 24.14
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A
1.
FLOW PATTERNS
Flow Patterns Around Isolated, Rectangular BlockType Buildings Buildings with even moderately complex shapes, such as L- or Ushaped structures, can generate flow patterns too complex to generalize for design. To determine flow conditions for such buildings, wind tunnel or water channel tests of physical scale models, full-scale tests of existing buildings, or appropriate computational modeling efforts are required (see the section on Physical and Computational Modeling). Thus, only isolated, rectangular block-type buildings are discussed here. Figure 1 shows the wind flow pattern around a single, wide, highrise building slab, with the main flow features indicated by numbers. The following description of the flow pattern is adapted from Blocken et al. (2011). As wind impinges on the building, part of the flow is deviated over the building (point 1) and part flows around it (2, 9). A stagnation point is present at the windward façade at about 70% of the building height. From this point, part of the flow is deviated upward (upwash) (3), part is deviated sideways (4), and a large part is directed downwards (downwash) (5). This downflow develops into a ground-level vortex (6) called standing vortex, frontal vortex, or horseshoe vortex. The main flow direction of this vortex near ground level is opposite to the direction of the approach flow. Both flows collide at the stagnation point at ground level in front of the building (7). The standing vortex subsequently wraps around the building corners, yielding the concentrated corner streams, characterized by very high wind speed amplification (8). These corner streams are further amplified by the general ground-level flow around the building (9). At the building’s leeward side, the underpressure zone results in recirculation flow (10, 13). A stagnation zone is also present downstream of the building at ground level, The preparation of this chapter is assigned to TC 4.3, Ventilation Requirements and Infiltration.
24.1 Copyright © 2017, ASHRAE
Fig. 1
Wind Flow Pattern Around High-Rise Building Slab
[Adapted from Blocken et al. (2016) and Beranek and Van Koten (1979)]
where the flow directions are opposite and wind speeds are low (11). Further downstream, the wind speed remains low for a considerable distance behind the building (i.e., the far wake) (12). Backflow is also responsible for creating slow-rotating vortices behind the building (13). Between these vortices and the corner streams (9) is a zone with a high velocity gradient (shear layer) that comprises small, fast-rotating vortices (16). Figure 2 provides a more detailed illustration of the wind flow pattern around an isolated building. It more clearly shows the vortical nature of the corner streams, and it indicates the areas of flow separation and reattachment and the flow in the near wake. It is important to note that Figures 1 and 2 only show the mean wind flow pattern, and that the actual flow pattern exhibits pronounced transient features, such as the build-up and collapse of the separation/ recirculation bubbles and periodic vortex shedding in the wake (Murakami 1993; Tominaga et al. 2008a). For a building with height H that is three or more times the width W of the upwind face, an intermediate stagnation zone can exist between the upwash and downwash regions, where surface streamlines pass horizontally around the building (Figure 3A). (In Figure 3, the upwind building surface is “folded out” to illustrate upwash, downwash, and stagnation zones.) Downwash on the lower surface of the upwind face separates from the building before it reaches ground level and moves upwind to form the standing vortex. Figure 3B shows the near-surface flow patterns for oblique approach flow. Strong vortices develop from the upwind edges of the roof, causing strong downwash onto the roof. High speeds in these vortices (vorticity) cause large negative pressures near roof corners that can be a hazard to roof-mounted equipment during high winds. In some extreme cases, the negative pressures can be strong enough to lift heavy objects such as roof pavers, which can result in a projectile hazard. When the angle between the wind direction and the upwind face of the building is less than about 70°, the upwash/downwash patterns on the upwind face of the building are less pronounced, as is the ground-level vortex shown in Figure 1 and 2. Figure 3B shows that,
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24.2
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 2 Wind Flow Pattern Around Isolated Building (Hunt et al. 1978)
Fig. 3 Surface Flow Patterns for Normal and Oblique Winds (Wilson 1979)
for an approach flow angle of 45°, streamlines remain close to the horizontal in their passage around the sides of the building, except near roof level, where the flow is drawn upwards into the roof edge vortices (Cochran 1992). Both the upwind velocity profile shape and its turbulence intensity strongly influence flow patterns and surface pressures (Melbourne 1979). The downwind wall of a building exhibits a region of low average velocity and high turbulence (i.e., a flow recirculation region) extending a distance Lr downwind. If the building has sufficient length L in the windward direction, the flow reattaches to the building and
may generate two distinct regions of separated recirculation flow, on the roof of the building and in its wake, as shown in Figure 4. Figure 4 also shows a rooftop recirculation cavity of length Lc at the upwind roof edge and a recirculation zone of length Lr downwind of the rooftop penthouse. Velocities near the downwind wall are typically onequarter of those at the corresponding upwind wall location. Figures 2 and 3 show that an upward flow exists over most of the downwind walls. Streamline patterns and the size of the wake(s) are generally independent of wind speed and depend mainly on building shape and upwind conditions. Because of the three-dimensional flow around a
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Airflow Around Buildings
24.3
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Fig. 4
Flow Recirculation Regions
building, the shape and size of the recirculation airflow are not constant over the surface. Airflow reattaches closer to the upwind building face along the edges of the building than it does near the middle of the roof and sidewalls (Figure 3). Recirculation cavity height Hc (Figure 4) also decreases near roof edges. Calculating characteristic dimensions for recirculation zones Hc, Lc, and Lr is discussed in Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications. For wind perpendicular to a building wall, the height H and width W of the upwind building face determine the scaling length R that characterizes the building’s influence on wind flow. According to Wilson (1979), 0.67 0.33 BL
R = Bs
(1)
where Bs = smaller of upwind building face dimensions H and W BL = larger of upwind building face dimensions H and W
Fig. 5
Buildings in (A) Converging and (B) Diverging Configuration
When BL is larger than 8Bs, use BL = 8Bs in Equation (1). For buildings with varying roof levels or with wings separated by at least a distance of Bs, only the height and width of the building face below the portion of the roof in question should be used to calculate R. Flow accelerates as the streamlines compress over the roof and decelerates as they spread downward over the wake on the downwind side of the building. The height above roof level where a building influences the flow is approximately 1.5R. In addition, roof pitch also begins to affect flow when it exceeds about 15° (1:4). When roof pitch reaches 20° (1:3), flow remains attached to the upwind pitched roof and produces a recirculation region downwind of the roof ridge that is larger than that for a flat roof.
Flow Patterns Around Building Groups In building groups, the flow patterns can interact, yielding a higher complexity. To determine flow conditions around building groups, wind tunnel or water channel tests of physical scale models, full-scale tests of existing buildings, or careful computational modeling efforts are required (see the section on Physical and Computational Modeling). English and Fricke (1997), Hosker (1984, 1985), Khanduri et al. (1998), Saunders and Melbourne (1979), and Walker et al. (1996) review the effects of nearby buildings, whereas Blocken et al. (2007a, 2008), Stathopoulos and Storms (1986), Yoshie et al. (2007), and others assess the effects of nearby buildings by wind tunnel testing and computational modeling. To illustrate the complexity of wind flow patterns induced by nearby buildings, Figure 5 provides a top view of two high rise buildings in V shape. Depending on the wind direction, this configuration is labeled as converging or diverging. Although the highest wind speed in the passage between both buildings might be expected to occur for the converging arrangement, wind tunnel tests and numerical simulations indicate that the diverging arrangement actually has higher wind speed. This is shown by the
Fig. 6 Amplification Factor K in Horizontal Plane at y = 2 m above Ground for Converging and Diverging Arrangement with H = 30 m and w = 75 m and 20 m (Blocken et al. 2008)
amplification factors in Figure 6, which are defined as the ratio of the local wind speed to the wind speed that would occur at the same height in absence of the buildings. The higher amplification factor in the diverging arrangement is caused by the lesser flow resistance in this configuration. The Venturi effect does not apply in this case: the Venturi effect refers to confined flows, whereas wind flow in the atmosphere is unconfined. Further information on this study can be found in Blocken et al. (2008). Figure 7 shows the flow over building arrays with increasing H/W. The description of these flow patterns is adopted from Oke (1988). If the buildings are sufficiently apart (H/W > 0.05), their flow fields do
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24.4
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Terrain Category 1
2
3
Fig. 7 Flow Regimes Associated with Airflow over Building Arrays of Increasing H/W
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(Oke 1988)
4
not interact, and the flow is called isolated roughness flow. When the buildings are positioned closer together, their flow patterns show some degree of interaction, mainly manifested as a disturbance of the wake structure (wake interference flow). When the ratio H/W increases further, a stable circulatory vortex is formed in the canyon and the bulk of the flow does not enter the canyon (skimming flow).
2.
WIND PRESSURE ON BUILDINGS
In addition to flow patterns described previously, the turbulence or gustiness of approaching wind and the unsteady character of separated flows cause surface pressures to fluctuate. Pressures discussed here are time-averaged values, with a full-scale averaging period of about 600 s. This is approximately the shortest time period considered to be a “steady-state” condition when considering atmospheric winds; the longest is typically 3600 s. Instantaneous pressures may vary significantly above and below these averages, and peak pressures two or three times the mean values are possible. Peak pressures are important with regard to structural loads, and mean values are more appropriate for computing infiltration and ventilation rates. Time-averaged surface pressures are proportional to wind velocity pressure pv given by Bernoulli’s equation: 2
a UH pv = -------------2
(2)
where pv = wind velocity pressure at roof level, Pa UH = approach wind speed at upwind wall height H, m/s [see Equation (4)] a = ambient (outdoor) air density, kg/m3
The proportional relationship is shown in the following equation, in which the difference ps between the pressure on the building surface and the local outdoor atmospheric pressure at the same level in an undisturbed wind approaching the building is ps = Cp pv
(3)
where Cp is the local wind pressure coefficient at a point on the building surface.
Atmospheric Boundary Layer Parameters Layer Exponent Thickness a , m
Description Large city centers, in which at least 50% of buildings are higher than 25 m, over a distance of at least 0.8 km or 10 times the height of the structure upwind, whichever is greater Urban and suburban areas, wooded areas, or other terrain with numerous closely spaced obstructions having the size of single-family dwellings or larger, over a distance of at least 460 m or 10 times the height of the structure upwind, whichever is greater Open terrain with scattered obstructions having heights generally less than 9 m, including flat open country typical of meteorological station surroundings Flat, unobstructed areas exposed to wind flowing over water for at least 1.6 km, over a distance of 460 m or 10 times the height of the structure inland, whichever is greater
0.33
460
0.22
370
0.14
270
0.10
210
Hmet , usually 10 m above ground level. The hourly average wind speed UH in the undisturbed wind approaching a building in its local terrain can be calculated from Umet as follows: met UH = Umet ----------- H met
a met
H ----
a
(4)
The atmospheric boundary layer thickness and exponent a for the local building terrain and amet and met for the meteorological station are determined from Table 1. Typical values for meteorological stations (category 3 in Table 1) are amet = 0.14 and met = 270 m. The values and terrain categories in Table 1 are consistent with those adopted in other engineering applications (e.g., ASCE Standard 7). Equation (4) gives the wind speed that occurs at a certain height H above the average height of local obstacles, such as buildings and vegetation, weighted by the plan area. At heights at or below this average obstacle height (e.g., at roof height in densely built-up suburbs), speed depends on the geometrical arrangement of the buildings, and Equation (4) is less reliable. An alternative mathematical description of the atmospheric boundary layer, which uses a logarithmic function, is given by Deaves and Harris (1978). Although their model is more complicated than the power law used in Equation (4), it more closely models the real physics of the atmosphere and has been adopted by several codes around the world (e.g., SA/SNZ 2002). Example 1. Assuming a 10 m/s anemometer wind speed for a height Hmet of 10 m at a nearby airport, determine the wind speed UH at roof level H = 15 m above grade for a building located in a city suburb. Solution: From Table 1, the atmospheric boundary layer properties for the anemometer are amet = 0.14 and met = 270 m. The atmospheric boundary layer properties at the building site are a = 0.22 and = 370 m. Using Equation (4), wind speed UH at 15 m is 15 0.22 270 0.14 -------= 7.8 m/s UH = 10 ------- 370- 10
Approach Wind Speed The local wind speed UH at the top of the wall required for Equation (2) is estimated by applying terrain and height corrections to the hourly wind speed Umet from a nearby meteorological station. Umet is generally measured in flat, open terrain (i.e., category 3 in Table 1). The anemometer that records Umet is located at height
Local Wind Pressure Coefficients Values of the mean local wind pressure coefficient Cp used in Equation (3) depend on building shape, wind direction, and influence of nearby buildings, vegetation, and terrain features. Accurate determination of Cp can be obtained only from wind tunnel model
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Airflow Around Buildings
24.5
tests of the specific site and building or full-scale tests. Ventilation rate calculations for single, unshielded rectangular buildings can be reasonably estimated using existing wind tunnel data. Many wind load codes (e.g., ASCE Standard 7; ASCE 1999; SA/SNZ Standard AS/NZS 1170.2) give mean pressure coefficients for common building shapes. Figure 8 shows pressure coefficients for walls of a tall rectangular cross section high-rise building sited in urban terrain (Davenport and Hui 1982). Figure 9 shows pressure coefficients for walls of a low-rise building (Holmes 1986). Generally, for high-rise buildings, height H is more than three times the crosswind width W. For H > 3W, use Figure 8; for H < 3W, use Figure 9. At a wind angle = 0° (e.g., wind perpendicular to the face in question), pressure coefficients are positive, and their magnitudes decrease near the sides and the top as flow velocities increase. As shown in Figure 8, Cp generally increases with height, which reflects increasing velocity pressure in the approach flow as wind speed increases with height. As wind direction moves off normal ( = 0°), the region of maximum pressure occurs closer to the upwind edge (B in Figure 8) of the building. At a wind angle of = 45°, pressures become negative at the downwind edge (A in Figure 8) of the front face. At some angle between 60° and 75°, pressures become negative over the whole front face. For = 90°, maximum suction (negative) pressure occurs near the upwind edge (B in Figure 8) of the building side and then recovers towards a lower-magnitude negative coefficient as the downwind edge (A in Figure 8) is approached. The degree of this recovery depends on the length of the side in relation to the width W of the structure. For wind angles larger than = 100°, the side is completely within the
separated flow of the wake and spatial variations in pressure over the face are not as great. In summary, the average pressure on a face is positive for wind angles from = 0° to almost 60° and negative (suction) for = 60° to 180°. A similar pattern of behavior in wall pressure coefficients for a low-rise building is shown in Figure 9. Here, recovery from strong suction with distance from the upwind edge is more rapid.
Surface-Averaged Wall Pressures Surface-averaged pressure coefficients may be used to determine ventilation and/or infiltration rates, as discussed in Chapter 16. Figure 10 shows the surface pressure coefficient Cs averaged over a complete wall of a high-rise building (Akins et al. 1979). Similar results for a low-rise building are shown in Figure 11 (based on methodology of Swami and Chandra 1987). This figure also includes values calculated from pressure distributions in Figure 9.
Roof Pressures Figure 12 shows the average pressure coefficient over the roof of a tall building (Akins et al. 1979). Surface pressures on the roof of a low-rise building depend strongly on roof slope. Figure 13 shows typical distributions for a wind direction normal to a side of the building. Note that the direction and magnitude of pressure coefficients are indicated by the direction and length of the arrows. For very low slopes (less than about 10°), pressures are negative over the whole roof surface. The magnitude is greatest within the separated flow zone near the leading edge and recovers toward the free-stream pressure downwind of
Fig. 8 Local Pressure Coefficients (Cp 100) for High-Rise Building with Varying Wind Direction (Davenport and Hui 1982)
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2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 11 Surface-Averaged Wall Pressure Coefficients for Low-Rise Buildings
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Courtesy of Florida Solar Energy Center (based on methodology of Swami and Chandra 1987)
Fig. 9 Local Pressure Coefficients for Low-Rise Building with Varying Wind Direction (Holmes 1986)
Fig. 12 Surface-Averaged Roof Pressure Coefficients for Tall Buildings (Akins et al. 1979)
Fig. 10 Surface-Averaged Wall Pressure Coefficients for High-Rise Buildings (Akins et al. 1979)
the edge. For intermediate slopes (about 10 to 20°), two largemagnitude low-pressure regions a re formed, one at the leading roof edge and another one beginning at the roof peak. For steeper slopes (greater than about 20°), pressures are weakly positive on the upwind slope and negative within the separated flow over the downwind slope. With a wind angle of about 45°, the vortices originating at the leading corner of a roof with a low slope can induce very large,
Fig. 13 Local Roof Pressure Coefficients for Roof of Low-Rise Buildings (Holmes 1983)
localized negative pressures (see Figure 3B). A similar vortex forms on the downwind side of a leading ridge end on a steep roof, as discussed in Cochran et al. (1999). Roof corner vortices and how to disrupt their influence are discussed in Cochran and Cermak (1992) and Cochran and English (1997).
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Airflow Around Buildings
24.7
Interference and Shielding Effects on Pressures Nearby structures strongly influence surface pressures on both high- and low-rise buildings, particularly for spacing-to-height ratios less than five, where distributions of pressure shown in Figures 8 to 13 do not apply. Although the effect of shielding for low-rise buildings is still significant at larger spacing, it is largely accounted for by the reduction in pv with increased terrain roughness. Bailey and Kwok (1985), Khanduri et al. (1998), Saunders and Melbourne (1979), Sherman and Grimsrud (1980), and Walker et al. (1996) discuss interference. English and Fricke (1997) discuss shielding through use of an interference index, and Walker et al. (1996) present a wind shadow model for predicting shelter factors. Chapter 16 gives shielding classes for air infiltration and ventilation applications.
3.
SOURCES OF WIND DATA
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Wind at Recording Stations To design buildings taking into account wind effects, wind speed and direction frequency data are necessary. The simplest forms of wind data are tables or charts of climatic normals, which give hourly average wind speeds, prevailing wind directions, and peak gust wind speeds for each month. This information can be found in sources such as The Weather Almanac (Bair 1992) and the Climatic Atlas of the United States (DOC 1968). Climatic design information, including wind speed at various frequencies of occurrence, is included in Chapter 14. Information on wind speed and direction frequencies is available from the National Climatic Data Center (NCDC) in Asheville, NC. Where more detailed information is required, digital records of hourly winds and other meteorological parameters are available from the NCDC for stations throughout the world. Most countries also have weather services that provide data. For example, in Canada, the Meteorological Service of Canada provides hourly meteorological data and summaries. When an hourly wind speed Umet at a specified probability level (e.g., the wind speed that is exceeded 1% of the time) is desired, but only the average annual wind speed Uannual is available for a given meteorological station, Umet may be estimated using Table 2. The ratios Umet /Uannual are based on long-term data from 24 weather stations widely distributed over North America. At these stations, Uannual ranges from 3.1 to 6.3 m/s The uncertainty ranges listed in Table 2 are one standard deviation of the wind speed ratios. Example 2 demonstrates the use of Table 2.
it is often called the prevailing wind, even though winds from other directions may be almost as frequent. When using long-term meteorological records, check the anemometer location history, because the instrument may have been relocated and its height varied. This can affect its directional exposure and the recorded wind speeds. Equation (4) can be used to correct wind data collected at different mounting heights. Poor anemometer exposure caused by obstructions or mounting on top of a building cannot be easily corrected, and records for that period should be deleted. If an estimate of the probability of an extreme wind speed outside the range of the recorded values at a site is required, the observations may be fitted to an appropriate probability distribution (e.g., a Weibull distribution) and the particular probabilities calculated from the resulting function (Figure 14). This process is usually repeated for each of 16 wind directions (e.g., 22.5° intervals). Note that most recent wind data records are provided in 10° intervals, for which the same method may be used, except that the process is repeated for each of 36 wind directions. If both types of data are to be used, one data set must be transformed to match the other. Where estimates at extremely low probability (high wind speed) are required, curve fitting at the tail of the probability distribution is very important and may require special statistical techniques applicable to extreme values (see Chapter 14). Building codes for wind loading on structures contain information on estimating extreme wind conditions. For ventilation applications, extreme winds are usually not required, and the 99th percentile limit can be accurately estimated from meteorological station data averaged over less than 10 years.
Example 2. The wind speed Umet that is exceeded 1% of the time (88 hours per year) is needed for a building pressure or exhaust dilution calculation. If Uannual = 4 m/s, find Umet . Solution: From Table 2, the wind speed Umet exceeded 1% of the time is 2.5 ± 0.4 times Uannual. For Uannual = 4 m/s, Umet is 10 m/s with an uncertainty range of 8.4 to 11.6 m/s at one standard deviation.
Using a single prevailing wind direction for design can cause serious errors. For any set of wind direction frequencies, one direction always has a somewhat higher frequency of occurrence. Thus, Table 2 Typical Relationship of Hourly Wind Speed Umet to Annual Average Wind Speed Uannual Percentage of Hourly Values That Exceed Umet 90% 75% 50% 25% 10% 5% 1%
Wind Speed Ratio Umet /Uannual 0.2 ± 0.1 0.5 ± 0.1 0.8 ± 0.1 1.2 ± 0.15 1.6 ± 0.2 1.9 ± 0.3 2.5 ± 0.4
Fig. 14
Frequency Distribution of Wind Speed and Direction
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Estimating Wind at Sites Remote from Recording Stations Many building sites are located far from the nearest long-term wind recording site, which is usually an airport meteorological station. To estimate wind conditions at such sites, the terrain surrounding both the anemometer site and the building site should be checked. In the simplest case of flat or slightly undulating terrain with few obstructions extending for large distances around and between the anemometer site and building site, recorded wind data can be assumed to be representative of that at the building site. Wind direction occurrence frequency at a building site should be inferred from airport data only if the two locations are on the same terrain, with no terrain features that could alter wind direction between them. In cases where the only significant difference between the anemometer site terrain and the building site terrain is surface roughness, the mean wind speed can be adjusted using Equation (4) and Table 1, to yield approximate wind velocities at the building site. Wind direction frequencies at the site are assumed to be the same as at the recording station. In using Equation (4), there may be cases where, for a given wind direction, the terrain upwind of either the building or recording site does not fall into just one of the categories in Table 1. The terrain immediately upwind of the site may fall into one category, and that farther upwind fall into a different category. For example, at a downtown airport the terrain may be flat and open (category 3) immediately around the recording instrument, but urban or suburban (category 2) a relatively short distance away. This difference in terrains also occurs when a building or recording site is in an urban area near open water or at the edge of town. In these cases, the suggested approach is to use the terrain category most representative of the average condition within approximately 1.6 km upwind of the site (Deaves 1981). If the average condition is somewhere between two categories described in Table 1, the values of a and can be interpolated from those given in the table. Several other factors are important in causing wind speed and direction at a building site to differ from values recorded at a nearby meteorological station. Wind speeds for buildings on hillcrests or in valleys where the wind is accelerated or channeled can be 1.5 times higher than meteorological station data. Wind speeds for buildings sheltered in the lee of hills and escarpments can be reduced to 0.5 times the values at nearby flat meteorological station terrain. In more complex terrain, both wind speed and direction may be significantly different from those at the distant recording site. In these cases, building site wind conditions should not be estimated from airport data. Options are either to establish an on-site wind recording station or to commission a detailed wind tunnel correlation study between the building site and long-term meteorological station wind observations. When wind is calm or light in the rural area surrounding a city, urban air tends to rise in a buoyant plume over the city center. This rising air, heated by anthropogenic sources and higher solar absorption in the city, is replaced by air pushed toward the city center from the edges. In this way, the urban heat island can produce light wind speeds and direction frequencies significantly different than those at a rural meteorological station.
4.
WIND EFFECTS ON SYSTEM OPERATION
A building with only upwind openings is positively pressurized because of the wind (Figure 15A). Building pressures are negative when there are only downwind openings (Figure 15B). A building with internal partitions and openings (Figure 15C) is under various pressures, depending on the relative sizes of openings and wind direction. With larger openings on the upwind face, the building interior tends toward positive pressure; the reverse is also true (see Figures 8 to 13, and Chapter 16).
Fig. 15 Sensitivity of System Volume to Locations of Building Openings, Intakes, and Exhausts With few exceptions, building intakes and exhausts cannot be located or oriented such that a prevailing wind ensures effective ventilation and air-conditioning system operation. Wind can assist or hinder inlet and exhaust fans, depending on their positions on the building, but even in locations with a predominant wind direction, the ventilating system must perform adequately for all other directions. To avoid variable system flow rates, use Figures 8, 9, and 12 as a guide to placing inlets and exhausts in locations where surface pressure coefficients do not vary greatly with wind direction. Also consider the potential for cross contamination; see Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications for details. Airflow through a wall opening results from differential pressures, which may exceed 125 Pa during high winds. Supply and exhaust systems, openings, dampers, louvers, doors, and windows make building flow conditions too complex for direct calculation. Iterative calculations are required because of the nonlinear dependence of volume flow rate on the differential pressure across an opening. Several multizone airflow models are available for these iterative calculations (Feustel and Dieris 1992; Walton and Dols 2005). Opening and closing of doors and windows by building occupants add further complications. In determining Cp , wind direction is more important than the position of an opening on a wall, as shown in Figures 8 and 9. Refer to Chapter 16 for details on wind effects on building ventilation, including natural and mechanical systems. Cooling towers and similar equipment should be oriented to take advantage of prevailing wind directions, if possible, based on careful study of meteorological data and flow patterns on the building for the area and time of year.
Natural and Mechanical Ventilation With natural ventilation, wind may augment, impede, or sometimes reverse the airflow through a building. For flat roof areas with large along-wind sides, wind can reattach to the roof downwind of the leading edge (see Figures 3 and 4). For peaked roofs, the upwind slope may be positively pressurized while the downwind slope may be negatively pressurized, as shown in Figure 12. Thus, any natural ventilation openings could see either a positive or negative pressure, dependent on wind speed and direction. Positive pressure existing where negative pressures were expected could reverse expected natural ventilation. These reversals can be avoided by using stacks,
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continuous roof ventilators, or other exhaust devices in which flow is augmented by wind. Mechanical ventilation is also affected by wind conditions. A low-pressure wall exhaust fan (12 to 25 Pa) can suffer drastic reduction in capacity. Flow can be reduced or reversed by wind pressure on upwind walls, or increased substantially when subjected to negative pressure on the downwind wall. Side walls may be subjected to either positive or negative pressure, depending on wind direction. Clarke (1967), measuring medium-pressure airconditioning systems (250 to 370 Pa), found flow rate changes of 25% for wind blowing into intakes on an L-shaped building compared to wind blowing away from intakes. Such changes in flow rate can cause noise at supply outlets and drafts in the space served. For mechanical systems, wind can be thought of as an additional pressure source in series with a system fan, either assisting or opposing it (Houlihan 1965). Where system stability is essential, supply and exhaust systems must be designed for high pressures (about 750 to 1000 Pa) or use devices to actively minimize unacceptable variations in flow rate. To conserve energy, the selected system pressure should be the minimum consistent with system needs. Quantitative estimates of wind effects on a mechanical ventilation system can be made by using the pressure coefficients in Figures 8 to 13 to calculate wind pressure on air intakes and exhausts. A simple worst-case estimate is to assume a system with 100% makeup air supplied by a single intake and exhausted from a single outlet. The building is treated as a single zone, with an exhaust-only fan as shown in Figure 16. This overestimates the effect of wind on system volume flow. Combining Equations (2) and (3), surface wind pressures at air intake and exhaust locations are a U H2 ps intake = Cp intake -------------2
(5) 2
a UH ps exhaust = Cp exhaust -------------2
(6)
For the single-zone building shown in Figure 16, a worst-case estimate of wind effect neglects any flow resistance in the intake grill and duct, making interior building pressure pinterior equal to outdoor wind pressure on the intake (pinterior = ps intake). Then, with all system flow resistance assigned to the exhaust duct in Figure 16, and a pressure rise pfan across the fan, pressure drop from outdoor intake to outdoor exhaust yields
Fig. 16
Intake and Exhaust Pressures on Exhaust Fan in Single-Zone Building
Q 2 (ps intake – ps exhaust) + pfan = Fsys --------2 AL
(7)
where Fsys is system flow resistance, AL is flow leakage area, and Q is system volume flow rate. This result shows that, for the worstcase estimate, the wind-induced pressure difference simply adds to or subtracts from the fan pressure rise. With inlet and exhaust pressures from Equations (5) and (6), the effective fan pressure rise pfan eff is pfan eff = pfan + pwind
(8)
where 2
a UH pwind = (Cp intake – Cp exhaust) -------------2
(9)
The fan is wind assisted when Cp intake > Cp exhaust and wind opposed when the wind direction changes, causing Cp intake < Cp exhaust. The effect of wind-assisted and wind-opposed pressure differences is shown in Figure 17. Example 3. Make a worst-case estimate for the effect of wind on the supply fan for a low-rise building with height H = 15 m, located in a city suburb. Use the hourly average wind speed that will be exceeded only 1% of the time and assume an annual hourly average speed of Uannual = 4 m/s measured on a meteorological tower at height Hmet = 10 m at a nearby airport. Outdoor air density is a = 1.2 kg/m3. Solution: From Table 2, the wind speed exceeded only 1% of the hours each year is a factor of 2.5 ± 0.4 higher than the annual average of 4 m/s, so the 1% maximum speed at the airport meteorological station is Umet = 2.5 4 = 10 m/s From Example 1, building wind speed UH is 7.8 m/s. A worst-case estimate of wind effect must assume intake and exhaust locations on the building that produce the largest difference (Cp intake – Cp exhaust) in Equations (8) and (9). From Figure 9, the largest difference occurs for the intake on the upwind wall AB and the exhaust on the downwind wall CD, with a wind angle AB = 0°. For this worst case, Cp intake = +0.8 on the upwind wall and Cp exhaust = –0.43 on the downwind wall. Using these coefficients in Equations (8) and (9) to evaluate effective fan pressure pfan eff ,
Fig. 17 Effect of Wind-Assisted and Wind-Opposed Flow
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1.2 7.8 pfan eff = pfan + [0.8 – (–0.43)] ---------------------2 = pfan + 44.9 Pa
This wind-assisted hourly averaged pressure is exceeded only 1% of the time (88 hours per year). When wind direction reverses, the outlet will be on the upwind wall and the inlet on the downwind wall, producing wind-opposed flow, changing the sign from +44.9 to –44.9 Pa. The importance of these pressures depends on their size relative to the fan pressure rise pfan, as shown in Figure 17.
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Minimizing Wind Effect on System Volume Flow Rate Wind effect can be reduced by careful selection of inlet and exhaust locations. Because wall surfaces are subject to a wide variety of positive and negative pressures, wall openings should be avoided when possible. When they are required, wall openings should be away from corners formed by building wings (see Figure 15). Mechanical ventilation systems should operate at a pressure high enough to minimize wind effect. Low-pressure systems and propeller exhaust fans should not be used with wall openings unless their ventilation rates are small or they are used in noncritical services (e.g., storage areas). Although roof air intakes in flow recirculation zones best minimize wind effect on system flow rates, current and future air quality in these zones must be considered. These locations should be avoided if a contamination source exists or may be added in the future. The best area is near the middle of the roof, because the negative pressure there is small and least affected by changes in wind direction (see Figure 12). Consider avoiding edges of the roof and walls, where large pressure fluctuations occur. Either vertical or horizontal (mushroom) openings can be used. On roofs with large areas, where intake may be outside the roof recirculation zone, mushroom or 180° gooseneck designs minimize impact pressure from wind flow. Vertical louvered openings or 135° goosenecks are undesirable for this purpose or for rain protection. Heated air or contaminants should be exhausted vertically through stacks, above the roof recirculation zone. Horizontal, louvered (45° down), and 135° gooseneck discharges are undesirable, even for heat removal systems, because of their sensitivity to wind effects. A 180° gooseneck for hot-air systems may be undesirable because of air impingement on tar and felt roofs. Vertically discharging stacks in a recirculation region (except near a wall) have the advantage of being subjected only to negative pressure created by wind flow over the tip of the stack. See Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications for information on stack design.
Chemical Hood Operation Wind effects can interfere with safe chemical hood operation. Supply volume flow rate variations can cause both disturbances at hood faces and a lack of adequate hood makeup air. Volume flow rate surges, caused by fluctuating wind pressures acting on the exhaust system, can cause momentary inadequate hood exhaust. If highly toxic contaminants are involved, surging is unacceptable. The system should be designed to eliminate this condition. On low-pressure exhaust systems, it is impossible to test the hoods under wind-induced, surging conditions. These systems should be tested during calm conditions for safe flow into the hood faces, and rechecked by smoke tests during high wind conditions. For more information on chemical hoods, see Chapter 16 of the 2015 ASHRAE Handbook—HVAC Applications. For more information on stack and intake design, see Chapter 45 of that volume.
5.
BUILDING PRESSURE BALANCE AND INTERNAL FLOW CONTROL
Proper building pressure balance avoids flow conditions that make doors hard to open and cause drafts. In some cases (e.g., office buildings), pressure balance may be used to prevent confinement of contaminants to specific areas. In other cases (e.g., laboratories), the correct internal airflow is towards the contaminated area.
Pressure Balance Although supply and exhaust systems in an indoor area may be in nominal balance, wind can upset this balance, not only because of its effects on fan capacity but also by superimposing infiltrated or exfiltrated air (or both). These effects can make it challenging to control environmental conditions. Where building balance and infiltration are important, consider the following: • Design HVAC system with pressure adequate to minimize wind effects • Include controls to regulate flow rate, pressure, or both • Separate supply and exhaust systems to serve each building area requiring control or balance • Use revolving or other self-closing doors or double-door air locks to noncontrolled adjacent areas, particularly exterior doors • Seal windows and other leakage sources • Close natural ventilation openings
Internal Flow Control Airflow direction is maintained by controlling pressure differentials between spaces. In a laboratory building, for example, peripheral rooms such as offices and conference rooms are kept at positive pressure, and laboratories at negative pressure, both with reference to corridor pressure. Pressure differentials between spaces are normally obtained by balancing supply system airflows in the spaces in conjunction with exhaust systems in the laboratories. Differential pressure instrumentation is normally used to control airflow. The pressure differential for a room adjacent to a corridor can be controlled using the corridor pressure as the reference. Outdoor pressure cannot usually control pressure differentials within internal spaces, even during periods of relatively constant wind velocity (wind-induced pressure). A single pressure sensor can measure the outdoor pressure at one point only and may not be representative of pressures elsewhere. Airflow (or pressure) in corridors is sometimes controlled by an outdoor reference probe that senses static pressure at doorways and air intakes. The differential pressure measured between the corridor and the outdoors may then signal a controller to increase or decrease airflow to (or pressure in) the corridor. Unfortunately, it is difficult to locate an external probe where it will sense the proper external static pressure. High wind velocity and resulting pressure changes around entrances can cause great variations in pressure. To measure ambient static pressure, the probe should be located where airflow streamlines are not affected by the building or nearby buildings. One possibility is at a height of 1.5R, as shown in Figure 18. However, this is usually not feasible. If an internal space is to be pressurized relative to ambient conditions, the pressure must be known on each exterior surface in contact with the space. For example, a room at the northeast corner of the building should be pressurized with respect to pressure on both the north and east building faces, and possibly the roof. In some cases, multiple probes on a single building face may be required. Figures 8 to 12 may be used as guides in locating external pressure probes. System volume and pressure control is described in Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications.
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Fig. 18 Flow Patterns Around Rectangular Block Building (modified from Hosker 1984)
6.
ENVIRONMENTAL IMPACTS OF BUILDING EXTERNAL FLOW
Pollutant Dispersion and Exhaust Reentrainment Pollutant dispersion around buildings is highly affected by the complex flow field. Contaminants are not always transported along the approaching flow direction and can be advected windward by reverse flows and retained in wake flows (Huber and Snyder 1982; Li and Meroney 1983; Stathopoulos et al. 2002). Intakes and exhausts should be installed carefully, considering wind direction and roof geometry (e.g., stack height, rooftop structures) to avoid air intake contamination and exhaust reentrainment (Gupta et al. 2012; Lazure et al. 2002; Stathopoulos et al. 2004). Empirical guidelines that can be used are given in Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications. More detailed prediction can be done by physical and numerical modeling, as explained in the section on Physical and Computational Modeling. State-of-the-art reviews on modeling of pollutant dispersion were performed by Canepa (2004), Di Sabatino et al. (2013), Lateb et al. (2016), Meroney (2004), and Tominaga and Stathopoulos (2013).
Pedestrian Wind Comfort and Safety Although thermal comfort is also important [e.g., Metje et al. (2008); Stathopoulos (2006)], wind comfort and safety generally only refer to the mechanical effects of wind on people [e.g., Lawson and Penwarden (1975); Willemsen and Wisse (2007)]. Particularly near high-rise buildings, high wind velocities can occur at pedestrian level that can be uncomfortable or even dangerous. For an isolated high-rise building or one amid low rise buildings, high wind speed at pedestrian level can be caused by the downflow that creates the standing vortex and the corner streams (see Figure 1). For building groups, amplified wind speed can occur in passages through and between buildings. Uncomfortable wind conditions can be detrimental to the success of new buildings. Wise (1970) reported shops that were left untenanted because of the windy environment that discouraged shoppers. Lawson and Penwarden (1975) reported dangerous wind conditions to be responsible for the death of two elderly women who were blown over by sudden wind gusts near a high-rise building. Many current urban authorities recognize the importance of pedestrian wind comfort and wind safety, and require studies before granting building permits for new buildings or new urban
areas. ASCE (2004) documented the state of the art of outdoor human comfort and its assessment. The first standard on wind comfort and wind safety was developed in the Netherlands and published in 2006 (NEN Standard 8100; Willemsen and Wisse 2007) and applied in several published case studies [e.g., Blocken et al. (2012)]. Reviews on studies of pedestrian wind comfort and safety were provided by Blocken (2014), Blocken and Stathopoulos (2013), Blocken et al. (2016), Mochida and Lun (2008), and Stathopoulos (2006). Although laser Doppler anemometry, particle image velocimetry, and large-eddy simulation (LES) are inherently more accurate techniques, Blocken et al. (2016) recommended faster and less expensive techniques for pedestrian-level wind (PLW) studies, such as hot-wire anemometry, Irwin probes, or steady Reynoldsaveraged Navier-Stokes computational fluid dynamics (RANS CFD) simulations. The reason is that their lower accuracy at lower amplification factors does not necessarily compromise the accuracy of PLW comfort assessment, because the higher amplification factors provide the largest contribution to the discomfort exceedance probability in the comfort criterion.
Wind-Driven Rain on Buildings Wind-driven rain (WDR), also called driving rain, is one of the most important moisture sources for building facades. It is an essential boundary condition for the analysis of the hygrothermal behavior and durability of historical and contemporary building facade components (Blocken and Carmeliet 2004; Dalgliesh and Surry 2003; Masters et al. 2008; Sanders 1996; Tang et al. 2004). Wind-driven rain can be assessed by full-scale measurements, wind tunnel measurements, semiempirical formulas, or numerical simulation with CFD. The experimental methods consist of measuring WDR with WDR gages. However, for practical purposes, measurements are generally time consuming, expensive, and often impractical. Blocken and Carmeliet (2006) and Högberg et al. (1999) found that WDR measurements are very prone to error. In addition, measurements made on facades of a particular building at a particular site have limited applicability to facades of other buildings at other sites. This awareness has led researchers to develop calculation models, which have been progressively improved throughout the years. Today, the most advanced and most frequently used models are the semiempirical model in ISO Standard 15927-3 (ISO model), the semiempirical model by Straube (1998) and
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Straube and Burnett (2000) (SB model), and the CFD model by Choi (1991, 1993, 1994) extended into the time domain by Blocken and Carmeliet (2002). State-of-the-art reviews on the assessment of WDR on building facades were provided by ASCE (2014) and Blocken and Carmeliet (2004, 2010).
7.
PHYSICAL AND COMPUTATIONAL MODELING
For many routine design applications, flow patterns and wind pressures can be estimated using the data and equations presented in the previous sections. Exhaust dilution for simple building geometries in homogeneous terrain environments (e.g., no larger buildings or terrain features nearby) can be estimated using the data and equations presented in the previous sections and in Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications. However, in critical applications, such as where health and safety are of concern, more accurate estimates may be required.
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Physical Modeling Measurements on small-scale models in wind tunnels or water channels can provide information for design before construction. These measurements can also be used as an economical method of performance evaluation for existing facilities. Full-scale testing is not generally useful in the initial design phase because of the time and expense required to obtain meaningful information, but it is useful for verifying data derived from physical modeling and for planning remedial changes to improve existing facilities (Cochran 2006). Detailed accounts of physical modeling, field measurements and applications, and engineering problems resulting from atmospheric flow around buildings are available in international journals, proceedings of conferences, and research reports on wind engineering (see the Bibliography). The wind tunnel is the main tool used to assess and understand airflow around buildings. Water channels or tanks can also be used, but are more difficult to implement and give only qualitative results for some cases. Models of buildings, complexes, and the local surrounding topography are constructed and tested in a simulated turbulent atmospheric boundary layer. Airflow, wind pressures, snow loads, structural response, or pollutant concentrations can then be measured directly by properly scaling wind, building geometry, and exhaust flow characteristics. Wind tunnel studies of natural ventilation are particularly suitable for buildings with large openings that provide a strong coupling between outdoor wind flow and indoor airflow (Karava et al. 2011; Kato et al. 1992). Dalgliesh (1975) and Petersen (1987a) found generally good agreement between the results of wind tunnel simulations and corresponding full-scale data. Cochran (1992) and Cochran and Cermak (1992) found good agreement between model and full-scale measurements of low-rise architectural aerodynamics and cladding pressures, respectively. Stathopoulos et al. (1999, 2002, 2004) obtained good agreement between model and full-scale measurements of the dispersion of gaseous pollutants from rooftop stacks on two different buildings in an urban environment.
Similarity Requirements Physical modeling is most appropriate for applications involving small-scale atmospheric motions, such as recirculation of exhaust downwind of a laboratory, wind loads on structures, wind speeds around building clusters, snow loads on roofs, and airflow over hills or other terrain features. Winds associated with tornadoes, thunderstorms, and large-scale atmospheric motion cannot currently be physically modeled accurately, although the physical modeling of tornadoes and thunderstorm downbursts is a current topic of significant research.
Snyder (1981) gives guidelines for fluid modeling of atmospheric diffusion. This report contains explicit directions and should be used whenever designing wind tunnel studies to assess concentration levels of air pollutants. ASCE Standard 7, ASCE Manual of Practice 67 (ASCE 1999), and AWES Quality Assurance Manual (AWES 2001) also provide guidance when wind tunnels are used for evaluating wind effects on structures. A complete and exact simulation of airflow over buildings and the resulting concentration or pressure distributions cannot be achieved in a physical model. However, this is not a serious limitation. Cermak (1971, 1975, 1976a, 1976b), Petersen (1987a, 1987b), and Snyder (1981) found that transport and dispersion of laboratory exhaust can be modeled accurately if the following criteria are met in the model and full scale: 1. Match exhaust velocity to wind speed ratios, Ve /UH. 2. Match exhaust to ambient air density ratios, e /a. 3. Match exhaust Froude numbers. Fr 2 = a V e2 /[(e – a)gd], where d is effective exhaust stack diameter. 4. Ensure fully turbulent stack gas flow by ensuring stack flow Reynolds number (Res = Ve d/) is greater than 2000 [where is the kinematic viscosity of ambient (outdoor) air], or by placing an obstruction inside the stack to enhance turbulence. 5. Ensure fully turbulent wind flow. 6. Scale all dimensions and roughness by a common factor. 7. Match atmospheric stability by the bulk Richardson number (Cermak 1975). For most applications related to airflow around buildings, neutral stratification is assumed, and no Richardson number matching is required. 8. Match mean velocity and turbulence distributions in the wind. 9. Ensure building wind Reynolds number (Reb = UH R/) is greater than 11 000 for sharp-edged structures, or greater than 90 000 for round-edged structures. 10. Ensure less than 5% blockage of wind tunnel cross section. For wind speeds, flow patterns, or pressure distributions around buildings, only conditions 5 to 10 are necessary. Usually, each wind tunnel study requires a detailed assessment to determine the appropriate parameters to match in the model and full scale. In wind tunnel simulations of exhaust gas recirculation, buoyancy of the exhaust gas (condition 3) is often not modeled. This allows using a high wind tunnel speed or a smaller model to achieve high enough Reynolds numbers (conditions 4, 5, and 9). Neglecting buoyancy is justified if density of building exhaust air is within 10% of the ambient (outdoor) air. Also, critical minimum dilution Dcrit occurs generally at wind speeds high enough to produce a wellmixed, neutrally stable atmosphere, allowing stability matching (condition 7) to be neglected (see Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications for discussion of Dcrit). However, in some cases and depending on emission sources, calm conditions may produce critical dilution. Nevertheless, omission of conditions 3 and 7 simplifies the test procedure considerably, reducing both testing time and cost. Buoyancy should be properly simulated for high-temperature exhausts such as boilers and diesel generators. Equality of model and prototype Froude numbers (condition 3) requires tunnel speeds of less than 0.5 m/s for testing. However, greater tunnel speeds may be needed to meet the minimum building Reynolds number requirement (condition 4).
Wind Simulation Facilities Boundary-layer wind tunnels are required for conducting most wind studies. The wind tunnel test section should be long enough to establish, upwind of the model building, a deep boundary layer that slowly changes with downwind distance. Other important wind tunnel characteristics include width and height of the test section, range of wind speeds, roof adjustability,
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and temperature control. Larger models can be used in tunnels that are wider and taller, which, in turn, give better measurement resolution. Model blockage effects can be minimized by an adjustable roof height. Temperature control of the tunnel surface and airflow is required when atmospheric conditions other than neutral stability are to be simulated. Boundary-layer characteristics appropriate for the site are established by using roughness elements on the tunnel floor that produce mean velocity, turbulence intensity profiles, and spectra characteristic of full scale. Water can also be used for the modeling fluid if an appropriate flow facility is available. Flow facilities may be in the form of a tunnel, tank, or open channel. Water tanks with a free surface ranging in size up to that of a wind tunnel test section have been used by towing a model (upside down) through the nonflowing fluid. Stable stratification can be obtained by adding a salt solution. This technique does not allow development of a boundary layer and therefore yields only approximate, qualitative information on flow around buildings. Water channels can be designed to develop thick, turbulent boundary layers similar to those developed in the wind tunnel. One advantage of such a flow system is ease of flow visualization, but this is offset by a greater difficulty in developing the correct turbulence structure and the measurement of flow variables and concentrations.
Designing Model Test Programs The first step in planning a test program is selecting the model length scale. This choice depends on cross-sectional dimensions of the test section, dimensions of the buildings to be modeled, and/or topographic features and thickness of the simulated atmospheric boundary layer. Typical geometric scales range from about 120:1 to 1000:1. Because a large model is desirable to meet minimum Reynolds and Froude number requirements, a wide test section is advantageous. In general, the model at any section should be small compared to the test section area so that blockage is less than 5% (Melbourne 1982). The test program must include specifications of the meteorological variables to be considered (e.g., wind direction, wind speed, thermal stability). Data taken at the nearest meteorological station should be reviewed to obtain a realistic assessment of wind climate for a particular site. Ordinarily, local winds around a building, pressures, and/or concentrations are measured for 16 wind directions (e.g., 22.5° intervals). This is easily accomplished by mounting the building model and its nearby surroundings on a turntable. More than 16 wind directions are required for highly toxic exhausts or for finding peak fluctuating pressures on a building. If only local wind information and pressures are of interest, testing at one wind speed with neutral stability is sufficient.
Computational Modeling Computational fluid dynamics (CFD) models attempt to resolve airflow around buildings by solving the Navier-Stokes equations or an approximate form of these equations in discretized form. The potential for computational wind engineering (CWE) has increased tremendously in the past decades. Nevertheless, there are some topics for which CWE in its current stage remains inappropriate. According to Stathopoulos (2000, 2002), there is great potential for CWE, but the numerical wind tunnel “is still virtual rather than real.” According to Murakami (2000), CWE has become a more popular tool, but results usually include numerical errors and prediction inaccuracies. According to a review by Blocken (2014), since 1960, computational wind engineering (CWE) has undergone a successful transition from an emerging field into an increasingly established field in wind engineering research, practice, and education, and its application has continued to spread to a very large range of topics, from pedestrian wind conditions to natural ventilation to
24.13 pollutant dispersion around buildings. However, in line with the reviews by Stathopoulos (1997, 2000, 2002), it is clear that the largest potential of CWE is still in the field of environmental wind engineering rather than structural wind engineering. Indeed, although CWE can adequately provide mean values of variables for pedestrian wind comfort and wind safety, natural ventilation, or wind-driven rain studies, correct prediction of peak values for wind loading studies, for example, remains very difficult. Methods for predicting turbulent flow around buildings include the following. Direct numerical simulation (DNS) directly resolves all the spatial and temporal scales in the flow based on the exact NavierStokes equations. This requires very extensive computational resources (runs lasting from several hours to days, depending on computer characteristics, power, and capacity) and can at present only be applied for flow in simple geometries and at low Reynolds numbers (Re). For complex, high-Re flows in wind engineering, application of DNS will not be possible in the foreseeable future. Large eddy simulation (LES) is a simplified method in which the spatially filtered Navier-Stokes equations are solved. Turbulent structures larger than the filter (sometimes taken equal to the grid size) are explicitly solved, whereas those smaller than the filter are modeled (i.e., approximated) by a subfilter model. Information on filtering and subfilter models can be found in Ferziger and Peric (2002), Geurts (2003), and Meyers et al. (2008). In Reynolds-averaged Navier-Stokes (RANS) simulation, the equations are obtained by averaging the Navier-Stokes equations (time-averaging if the flow is statistically steady, or ensembleaveraging for time-dependent flows). With RANS, only the mean flow is solved, whereas all scales of turbulence must be modeled. Averaging generates additional unknowns for which turbulence models are required. Many turbulence models are available, but no single turbulence model is universally accepted as being the best for all types of applications. In addition, hybrid RANS/LES approaches are available, in which unsteady RANS (URANS) is used near the wall, and LES in the rest of the flow field. This avoids the excessively high near-wall grid resolution required for application of LES near walls in highRe flow problems. An example of a hybrid RANS/LES approach is detached eddy simulation (DES), as proposed by Spalart et al. (1997). The statistically steady RANS method is the most widely applied and validated in CWE. It has been used for a wide range of building applications, including estimating pressure coefficients (Meroney et al. 2002; Murakami et al. 1992; Oliveira and Younis 2000; Richards and Hoxey 1992; Stathopoulos 1997; Stathopoulos and Zhou 1993), natural ventilation (Chen 2009; Evola and Popov 2006; Kato et al. 1997; Norton et al. 2009; Ramponi and Blocken 2012; van Hooff and Blocken 2010), wind-driven rain (Blocken and Carmeliet 2004; 2010; Choi 1993, 1994; Tang and Davidson 2004), pollutant dispersion (Cowan et al. 1997; Dawson et al. 1991; Gousseau et al. 2011; Leitl et al. 1997; Li and Stathopoulos 1997; Meroney 2004; Meroney et al. 1999; Tominaga and Stathopoulos 2010, 2011, 2013), pedestrian wind conditions (Blocken et al. 2008, 2012; Richards et al. 2002; Stathopoulos and Baskaran 1996; Yoshie et al. 2007), snow drift (Sundsbo 1998; Thiis 2000; Tominaga and Mochida 1999), and cooling tower drift (Meroney 2006, 2008). Although many past applications of RANS have been limited to isolated buildings or relatively simple building arrangements, large and sometimes very large discrepancies have been found in comparisons with wind tunnel and full-scale measurements. These are at least partly attributed to turbulence model limitations and to the statistically steady solution of flows that exhibit pronounced transient features, such as intermittent separation, recirculation zones, and vortex shedding [e.g., Murakami (1993), Tominaga et al. (2008a)]. In addition, a wide range of other computational aspects can contribute to uncer-
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24.14
2017 ASHRAE Handbook—Fundamentals (SI)
tainties and errors, divided by COST 732 (Franke et al. 2007) into two broad categories: physical and numerical. Physical modeling errors and uncertainties result from assumptions and approximations made in the mathematical description of the physical process. Examples are simplifications of the actual physical complexity (e.g., using RANS instead of DNS) and uncertainties and/or simplifications of the geometric and physical boundary conditions. Numerical errors and uncertainties are the result of the numerical solution of the mathematical model. Examples are computer programming errors, computer round-off errors, spatial and temporal discretization errors, and iterative convergence errors. LES is a time-dependent approach in which more of the turbulence is resolved. It therefore has a larger potential to provide accurate results than statistically steady RANS simulations (Murakami et al. 1992; Tominaga et al. 1997). LES also provides more information about the flow, such as instantaneous and peak wind speeds, pressures, and pollutant concentrations. However, it requires considerably higher CPU times and memory than RANS. It also requires time- and space-resolved data as boundary conditions to properly simulate inflow. Such experimental data are rarely available in practice (Franke et al. 2007). LES is also considered to require more experience for users to apply effectively than does RANS. Several studies have compared RANS and LES modeling for atmospheric dispersion for generic configurations such as isolated buildings (Tominaga and Stathopoulos 2010) and street canyons (Salim et al. 2011; Tominaga and Stathopoulos 2011) and for actual urban environments (Gousseau et al. 2011; Hanna et al. 2006), where LES is shown to consistently outperform steady RANS modeling. However, the drawbacks of LES imply that the practical application of CWE will continue to be mainly based on statistically steady RANS for a considerable while. Guidelines for using CFD have been developed and assembled to help users avoid, reduce, and estimate errors and uncertainties in applying CFD. ERCOFTAC (2000) provides extensive guidelines for industrial CFD applications, many of which are also applicable to CWE. Franke et al. (2007) assembled a comprehensive bestpractice guideline document for CFD simulation of flows in the urban environment. Important guidelines for application of CFD to pedestrian wind conditions around buildings and for predicting wind loads on buildings have been developed by the Architectural Institute of Japan and reported by Mochida et al. (2002), Tamura et al. (2008), Tominaga et al. (2008b), and Yoshie et al. (2007). A set of ten tips and tricks for CFD simulations in urban physics has been provided by Blocken (2015). Other efforts have focused on specific problems, such as those encountered in simulating equilibrium atmospheric boundary layers in computational domains [e.g., Balogh and Parente (2015); Blocken et al. (2007a, 2007b); Gorlé et al. (2009); Hargreaves and Wright (2007); Parente et al. (2011); Richards and Hoxey (1993); Richards and Norris (2011, 2015); Yang et al. (2008)]. Most of these guidelines apply to statistically steady RANS simulations. Regardless of whether RANS or LES is used, evaluating the accuracy of CFD results by comparing them with wind tunnel or field experiments is very important because turbulence models are based on assumptions; no turbulence model is universally valid for all applications. Physical modeling therefore remains an indispensable tool in (computational) wind engineering.
8.
SYMBOLS
a = exponent in power law wind speed profile for local building terrain, Equation (4) and Table 1, dimensionless AL = flow leakage area, Equation (7), m2 amet = exponent a for the meteorological station, Equation (4) and Table 1, dimensionless BL = larger of two upwind building face dimensions H and W, Equation (1), m
Bs = smaller of two upwind building face dimensions H and W, Equation (1), m Cp = local wind pressure coefficient for building surface, Equation (3), dimensionless Cs = surface-averaged pressure coefficient, Figure 6, dimensionless d = effective stack diameter, m Dcrit = critical dilution factor at roof level for uncapped vertical exhaust at critical wind speed (see Chapter 45 of the 2015 ASHRAE Handbook—HVAC Applications), dimensionless Fr = Froude number, dimensionless Fsys = system flow resistance, Equation (7), dimensionless g = acceleration of gravity, 9.8 m/s2 H = wall height above ground on upwind building face, Equation (4) and Figure 3, m Hc = maximum height above roof level of upwind roof edge flow recirculation zone, Figure 4, m Hmet = height of anemometer at meteorological station, Equation (4), m hs = exhaust stack height (typically above roof unless otherwise specified, m (see Chapter 45 in the 2015 ASHRAE Handbook— HVAC Applications) L = length of building in wind direction, Figure 3, m Lc = length of upwind roof edge recirculation zone, Figure 4, m Lr = length of flow recirculation zone behind rooftop obstacle or building, Figure 4, m ps = wind pressure difference between exterior building surface and local ambient (outdoor) atmospheric pressure at same elevation in undisturbed approach wind, Equation (3), Pa pv = wind velocity pressure at roof level, Equation (2), Pa Q = volumetric flow rate, Equation (7), m3/s R = scaling length for roof flow patterns, Equation (1), m Reb = building Reynolds number, dimensionless Res = stack flow Reynolds number, dimensionless S = stretched-string distance; shortest distance from exhaust to intake over obstacles and along building surface, m (see Chapter 45 in the 2015 ASHRAE Handbook—HVAC Applications) Uannual = annual average of hourly wind speeds Umet, Table 2, m/s UH = mean wind speed at height H of upwind wall in undisturbed flow approaching building, Equation (2) and Figure 4, m/s Umet = meteorological station hourly wind speed, measured at height Hmet above ground in smooth terrain, Equation (4) and Table 2, m/s Ve = exhaust face velocity, m/s W = width of upwind building face, Figure 3, m
Greek = fully developed atmospheric boundary layer thickness, Equation (4) and Table 1, m met = atmospheric boundary layer thickness at meteorological station, Equation (4) and Table 1, m pfan = pressure rise across fan, Equation (7), Pa pfan eff = effective pressure rise across fan, Equation (8), Pa pwind = wind-induced pressure, Equations (8) and (9), Pa = angle between perpendicular line from upwind building face and wind direction, Figures 8 to 12, degrees = kinematic viscosity of ambient (outdoor) air, m2/s a = ambient (outdoor) air density, Equation (2), kg/m3 e = density of exhaust gas mixture, kg/m3
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. Akins, R.E., J.A. Peterka, and J.E. Cermak. 1979. Averaged pressure coefficients for rectangular buildings. Wind Engineering: Proceedings of the Fifth International Conference, vol. 7, pp. 369-380. ASCE. 2010. Minimum design loads for buildings and other structures. Standard ASCE/SEI 7-10. American Society of Civil Engineers, Reston, VA.
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Airflow Around Buildings Norton, T., J. Grant, R. Fallon, and D.W. Sun. 2009. Assessing the ventilation effectiveness of naturally ventilated livestock buildings under wind dominated conditions using computational fluid dynamics. Biosystems Engineering 103(1):78-99. Oke, T.R. 1988. Street design and urban canopy layer climate. Energy and Buildings 11:103-113. Oliveira, P.J., and B.A. Younis. 2000. On the prediction of turbulent flows around full-scale buildings. Journal of Wind Engineering and Industrial Aerodynamics 86(2-3):203-220. Parente, A., C. Gorlé, J. van Beeck, and C. Benocci. 2011. Improved k- model and wall function formulation for the RANS simulation of ABL flows. Journal of Wind Engineering and Industrial Aerodynamics 99(4):267-278. Petersen, R.L. 1987a. Wind tunnel investigation of the effect of platformtype structures on dispersion of effluents from short stacks. Journal of Air Pollution Control Association 36:1347-1352. Petersen, R.L. 1987b. Designing building exhausts to achieve acceptable concentrations of toxic effluent. ASHRAE Transactions 93(2):2165-2185. Ramponi, R., and B. Blocken. 2012. CFD simulation of cross-ventilation for a generic isolated building: Impact of computational parameters. Building and Environment 53:34-48. Richards, P.J., and R.P. Hoxey. 1992. Computational and wind tunnel modelling of mean wind loads on the Silsoe structures building. Journal of Wind Engineering and Industrial Aerodynamics 43(1-3):1641-1652. Richards, P.J., and R.P. Hoxey. 1993. Appropriate boundary conditions for computational wind engineering models using the k- turbulence model. Journal of Wind Engineering and Industrial Aerodynamics 46/47:145153. Richards, P.J., and S.E. Norris. 2011. Appropriate boundary conditions for computational wind engineering models revisited. Journal of Wind Engineering and Industrial Aerodynamics 99(4):257-266. Richards, P.J., and S.E. Norris. 2015. Appropriate boundary conditions for a pressure-driven boundary layer. Journal of Wind Engineering and Industrial Aerodynamics 142(7):43-52. Richards, P.J., G.D. Mallison, D. McMillan, and Y.F. Li. 2002. Pedestrian level wind speeds in downtown Auckland. Wind and Structures 5(2-4): 151-164. Salim, M.S., R. Buccolieri, A. Chan, and S. Di Sabatino. 2011. Numerical simulation of atmospheric pollutant dispersion in an urban street canyon: Comparison between RANS and LES. Journal of Wind Engineering and Industrial Aerodynamics 99(2-3):103-113. Sanders, C. 1996. Heat, air and moisture transfer in insulated envelope parts: Environmental conditions. International Energy Agency, Annex 24. Final report, vol. 2. Acco, Leuven. SA/SNZ. 2002. Structural design actions—Part 2: Wind actions. Standard AS/NZS 1170.2:2002. Standards Australia International Ltd., Sydney. Saunders, J.W., and W.H. Melbourne. 1979. Buffeting effect of upwind buildings. Fifth International Conference on Wind Engineering. Pergamon Press, pp. 593-606. Sherman, M.H., and D.T. Grimsrud. 1980. The measurement of infiltration using fan pressurization and weather data. Report LBL-10852. Lawrence Berkeley Laboratory, University of California. Snyder, W.H. 1981. Guideline for fluid modeling of atmospheric diffusion. Environmental Protection Agency Report EPA-600/881-009. Spalart, P., W.-H. Jou, M. Strelets, and S. Allmaras. 1997. Comments on the feasibility of LES for wings and on the hybrid RANS/LES approach. Advances in DNS/LES, 1st AFOSR International Conference on DNS/ LES, Greden Press. Stathopoulos, T. 1997. Computational wind engineering: Past achievements and future challenges. Journal of Wind Engineering and Industrial Aerodynamics 67/68:509-532. Stathopoulos, T. 2000. The numerical wind tunnel for industrial aerodynamics: Real or virtual in the new millennium? Third International Symposium on Computational Wind Engineering. PF Consultants. Stathopoulos, T. 2002. The numerical wind tunnel for industrial aerodynamics: Real or virtual in the new millennium? Wind and Structures 5(2-4): 193-208. Stathopoulos T. 2006. Pedestrian level winds and outdoor human comfort. Journal of Wind Engineering and Industrial Aerodynamics 94(11):769780. Stathopoulos, T., and B.A. Baskaran. 1996. Computer simulation of wind environmental conditions around buildings. Engineering Structures 18(11):876-885.
24.17 Stathopoulos, T., and R. Storms. 1986. Wind environmental conditions in passages between buildings. Journal of Wind Engineering and Industrial Aerodynamics 24(1):19-31. Stathopoulos, T., and Y.S. Zhou. 1993. Numerical simulation of windinduced pressures on buildings of various geometries. Journal of Wind Engineering and Industrial Aerodynamics 46/47:419-430. Stathopoulos, T., L. Lazure, and P. Saathoff. 1999. Tracer gas investigation of reingestion of building exhaust in an urban environment. IRSST Report R-213, Robert-Sauvé Institute of Occupational Health and Safety Research (IRSST), Montreal, Canada. Stathopoulos, T., L. Lazure, P. Saathoff, and X. Wei. 2002. Dilution of exhaust from a rooftop stack on a cubical building in an urban environment. Atmospheric Environment 36:4577-4591. Stathopoulos, T., L. Lazure, P. Saathoff, and A. Gupta. 2004. The effect of stack height, stack location and rooftop structures on air intake contamination. A laboratory and full-scale study. IRSST Report R-392. RobertSauvé Institute of Occupational Health and Safety Research (IRSST), Montreal, Canada. Straube, J.F. 1998. Moisture control and enclosure wall systems. Ph.D. dissertation, Civil Engineering, University of Waterloo, Ontario. Straube, J.F., and E.F.P. Burnett. 2000. Simplified prediction of driving rain on buildings. Proceedings of the International Building Physics Conference, Eindhoven, The Netherlands, pp. 375-382. Sundsbo, P.A. 1998. Numerical simulations of wind deflection fins to control snow accumulation in building steps. Journal of Wind Engineering and Industrial Aerodynamics 74-76:543-552. Swami, M.V., and S. Chandra. 1987. Procedures for calculating natural ventilation airflow rates in buildings. Final Report FSEC-CR-163-86. Florida Solar Energy Center, Cape Canaveral. Tamura, T., K. Nozawa, and K. Kondo. 2008 AIJ guide for numerical prediction of wind loads on buildings. Journal of Wind Engineering and Industrial Aerodynamics 96(10-11):1974-1984. Tang, W., and C.I. Davidson. 2004. Erosion of limestone building surfaces caused by wind-driven rain: 2. Numerical modeling. Atmospheric Environment 38(33):5601-5609. Tang, W., C.I. Davidson, S. Finger, and K. Vance. 2004. Erosion of limestone building surfaces caused by wind-driven rain: 1. Field measurements. Atmospheric Environment 38(33):5589-5599. Thiis, T.K. 2000. A comparison of numerical simulations and full-scale measurements of snowdrifts around buildings. Wind and Structures 3(2): 73-81. Tominaga, Y., and A. Mochida. 1999. CFD prediction of flowfield and snowdrift around a building complex in a snowy region. Journal of Wind Engineering and Industrial Aerodynamics 81(1-3):273-282. Tominaga, Y., and T. Stathopoulos. 2010. Numerical simulation of dispersion around an isolated cubic building: Model evaluation of RANS and LES. Building and Environment 45(10):2231-2239. Tominaga, Y., and T. Stathopoulos. 2011. CFD modeling of pollution dispersion in a street canyon: Comparison between LES and RANS. Journal of Wind Engineering and Industrial Aerodynamics 99(4):340-348. Tominaga, Y., and T. Stathopoulos. 2013. CFD simulation of near-field pollutant dispersion in the urban environment: A review of current modelling techniques. Atmospheric Environment 79:716-730. Tominaga, Y., S. Murakami, and A. Mochida. 1997. CFD prediction of gaseous diffusion around a cubic model using a dynamic mixed SGS model based on composite grid technique. Journal of Wind Engineering and Industrial Aerodynamics 67/68:827-841. Tominaga, Y., A. Mochida, S. Murakami, and S. Sawaki. 2008a. Comparison of various revised k- models and LES applied to flow around a highrise building model with 1:1:2 shape placed within the surface boundary layer. Journal of Wind Engineering and Industrial Aerodynamics 96(4): 389-411. Tominaga, Y., A. Mochida, R. Yoshie, H. Kataoka, T. Nozu, M. Yoshikawa, and T. Shirasawa. 2008b. AIJ guidelines for practical applications of CFD to pedestrian wind environment around buildings. Journal of Wind Engineering and Industrial Aerodynamics 96(10-11):1749-1761. van Hooff, T., and B. Blocken. 2010. Coupled urban wind flow and indoor natural ventilation modelling on a high-resolution grid: A case study for the Amsterdam Arena stadium. Environmental Modelling and Software 25(1):51-65.
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Walker, I.S., D.J. Wilson, and T.W. Forest. 1996. Wind shadow model for air infiltration sheltering by upwind obstacles. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 2(4):265-283. Walton, G.N., and W.S. Dols. 2005. CONTAM 2.4 user guide and program documentation. NISTIR 7251. National Institute of Standards and Technology, Gaithersburg, Maryland. Willemsen, E., and J.A. Wisse. 2007. Design for wind comfort in The Netherlands: Procedures, criteria and open research issues. Journal of Wind Engineering and Industrial Aerodynamics 95(9-11):1541-1550. Wilson, D.J. 1979. Flow patterns over flat roofed buildings and application to exhaust stack design. ASHRAE Transactions 85(2):284-295. Wise, A.F.E. 1970. Wind effects due to groups of buildings. Proceedings of the Royal Society Symposium Architectural Aerodynamics, Session 3, Effect of Buildings on the Local Wind, London. pp. 26-27. Yang, W., Y. Quan, X. Jin, Y. Tamura, and M. Gu. 2008. Influences of equilibrium atmosphere boundary layer and turbulence parameters on wind load distributions of low-rise buildings. Journal of Wind Engineering and Industrial Aerodynamics 96(10-11):2080-2092. Yoshie, R., A. Mochida, Y. Tominaga, H. Kataoka, K. Harimoto, T. Nozu, and T. Shirasawa. 2007. Cooperative project for CFD prediction of pedestrian wind environment in the architectural institute of Japan. Journal of Wind Engineering and Industrial Aerodynamics 95(9-11):15511578.
BIBLIOGRAPHY AIHA. 2012. Laboratory ventilation. ANSI/AIHA Standard Z9.5-2012. American Industrial Hygiene Association, Fairfax, VA. ASCE. 1999. Wind tunnel studies of buildings and structures. Manual of Practice 67. N. Isyumov, ed. American Society of Civil Engineers, New York. ASCE. 2012. Wind tunnel testing for buildings and other structures. ASCE/ SEI Standard 49-12. Structural Engineering Institute, American Society of Civil Engineers, New York. Cermak, J.E. 1977. Wind-tunnel testing of structures. Journal of the Engineering Mechanics Division, ASCE 103, EM6:1125. Cermak, J.E., ed. 1979. Wind engineering. Wind Engineering: Proceedings of the Fifth International Conference, Colorado State University, Fort Collins, CO. Pergamon Press, New York.
Clarke, J.H. 1965. The design and location of building inlets and outlets to minimize wind effect and building reentry of exhaust fumes. Journal of American Industrial Hygiene Association 26:242. CWE. 1993. Proceedings of the 1st International Symposium on Computational Wind Engineering, Tokyo, Japan. Elsevier. CWE. 1997. Proceedings of the 2nd International Symposium on Computational Wind Engineering, Colorado State University, Fort Collins. Elsevier. CWE. 2000. Proceedings of the 3rd International Symposium on Computational Wind Engineering. PF Consultants. CWE. 2006. Proceedings of the 4th International Symposium on Computational Wind Engineering, Yokohama, Japan. Elsevier. Defant, F. 1951. Local winds. In Compendium of meteorology, pp. 655-672. American Meteorology Society, Boston. Elliot, W.P. 1958. The growth of the atmospheric internal boundary layer. Transactions of the American Geophysical Union 39:1048-1054. ESDU. 1990. Strong winds in the atmospheric boundary layer. Part 1: Mean hourly wind speeds, pp. 15-17. Engineering Science Data Unit, Item 8226, London. Geiger, R. 1966. The climate near the ground. Harvard University, Cambridge. Houghton, E.L., and N.B. Carruthers. 1976. Wind forces on buildings and structures: An introduction. Edward Arnold, London. Landsberg, H. 1981. The urban climate. Academic Press, New York. Meroney, R.N., and B. Bienkiewicz, eds. 1997. Computational wind engineering 2. Elsevier, Amsterdam. Moonen, P., T. Defraeye, V. Dorer, B. Blocken, and J. Carmeliet. 2012. Urban physics: Effect of the microclimate on comfort, health and energy demand. Frontiers of Architectural Research 1(3):197-228. Panofsky, H.A., and J.A. Dutton. 1984. Atmospheric turbulence: Models and methods for engineering applications. John Wiley & Sons, New York. Simiu, V., and R. Scanlan. 1986. Wind effects on structures: An introduction to wind engineering, 2nd ed. Wiley Interscience, New York. Svendsen, S.D. 1955. Driving rain. Experimental research on the resistance of external walls against rain penetration. Report 20, Norwegian Building Research Institute, Oslo. WERC. 1985. Proceedings of the 5th U.S. National Conference on Wind Engineering, 6-8 November, Texas Tech University, Lubbock. K.C. Mehta and R.A. Dillingham, eds. Wind Engineering Research Center, Lubbock.
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HEAT, AIR, AND MOISTURE CONTROL IN BUILDING ASSEMBLIES—FUNDAMENTALS FUNDAMENTALS................................................................... 25.1 Terminology and Symbols........................................................ 25.1 Hygrothermal Loads and Driving Forces................................................................................... 25.2 HEAT TRANSFER ................................................................... 25.5 Steady-State Thermal Response............................................... 25.5 Transient Thermal Response ................................................... 25.8 AIRFLOW ................................................................................ 25.9 MOISTURE TRANSFER........................................................ 25.10
Moisture Storage in Building Materials ................................ Moisture Flow Mechanisms ................................................... COMBINED HEAT, AIR , AND MOISTURE TRANSFER..... SIMPLIFIED HYGROTHERMAL DESIGN CALCULATIONS AND ANALYSES ................................... Surface Humidity and Condensation ..................................... Interstitial Condensation and Drying .................................... TRANSIENT COMPUTATIONAL ANALYSIS ....................... Criteria to Evaluate Hygrothermal Simulation Results .........
ROPER design of space heating, cooling, and air-conditioning systems requires detailed knowledge of the building envelope’s overall heat, air, and moisture performance. This chapter discusses the fundamentals of combined heat, air, and moisture movement as it relates to the analysis and design of envelope assemblies. Guidance for designing mechanical systems is found in other chapters of the ASHRAE Handbook. Because heat, air, and moisture transfer are coupled and interact closely with each other, they should not be treated separately. For example, improving a building envelope’s energy performance may cause moisture-related problems. Conversely, evaporation of water and removal of moisture by other means are processes that require energy. Only a sophisticated moisture control strategy can ensure hygienic conditions and adequate durability for modern, energyefficient building assemblies. Effective moisture control design must deal with all hygrothermal (heat and humidity) loads acting on the building envelope.
homogeneity.) Units are W/(m·K). For anisotropic materials, the direction of heat flux in the material must be noted. Thermal conductivity must be evaluated for a specific mean temperature, thickness, age, and moisture content. Thermal conductivity is normally considered an intrinsic property of a homogenous material. In porous materials, heat flow occurs by a combination of conduction, convection, and radiation, and may depend on orientation, direction, or both. When nonconductive modes of heat transfer occur within the specimen or the test specimen is nonhomogeneous, the measured property of such materials is called apparent thermal conductivity. The specific test conditions (e.g., sample thickness, orientation, environment, environmental pressure, surface temperature, mean temperature, temperature difference, moisture distribution) should be reported with the values of apparent thermal conductivity. The symbol kapp (or app) is used to denote the absence of pure conduction or to indicate that all values reported are apparent. Materials with a low apparent thermal conductivity are called insulation materials (see Chapter 26 for more detail). Thermal resistivity ru is the reciprocal of thermal conductivity. Units are (m·K)/W. Thermal resistance R is an extrinsic property that describes the resistance of a material layer or assembly to heat transfer. It is determined by the steady-state or time-averaged temperature difference (between two defined surfaces of a material layer within a building assembly) that induces a unit heat flux, in (m2 ·K)/W. When the two defined surfaces have unequal areas, as with heat flux through material layers of nonuniform thickness, an appropriate mean area and mean thickness must be given. Thermal resistance formulas involving materials that are not uniform slabs must contain shape factors to account for the area variation involved. When heat flux occurs by conduction alone, the thermal resistance of a layer of constant thickness may be obtained by dividing the material’s thickness by its thermal conductivity. When several modes of heat transfer are involved, the apparent thermal resistance may be obtained by dividing the material’s thickness by its apparent thermal conductivity. When air circulates within or passes through insulation, as may happen in lowdensity fibrous materials, the apparent thermal resistance is affected. Thermal resistances of common building and insulation materials are listed in Chapter 26. Thermal conductance C is the reciprocal of thermal resistance. Units are W/(m2 ·K). Heat transfer or surface film coefficient h is the value that describes the total heat flux by both convection and radiation between a surface and the surrounding environment. It is defined as the heat transfer per unit time and unit area induced by a unit temperature difference between the surface and the reference temperature in the surrounding environment. Units are W/(m2 ·K). For convection to
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P
1. 1.1
FUNDAMENTALS
TERMINOLOGY AND SYMBOLS
The following heat, air, and moisture definitions, properties, and symbols are commonly used. A building envelope or building enclosure provides physical separation between the indoor space and the outdoor environment. A building assembly is any part of the building envelope, such as a wall, window, or roof assembly, that faces the interior and exterior of the building. A building component is any element, layer, or material within a building assembly.
Heat Specific heat capacity c is the change in heat (energy) of a unit mass of material for a unit change of temperature in J/(kg·K). Volumetric heat capacity c is the change in heat stored in a unit volume of material for a unit change of temperature, in J/(m3 ·K). Heat flux q, a vector, is the time rate of heat transfer through a unit area, in the direction perpendicular to that area, in W/m2. Thermal conductivity k [in Europe, the Greek letter (lambda) is used] is a material property describing ability to conduct heat, and is defined by Fourier’s law of heat conduction. Thermal conductivity is the property that describes heat flux through a unit thickness of a material in a direction perpendicular to the isothermal planes, induced by a unit temperature difference. (ASTM Standard C168 defines The preparation of this chapter is assigned to TC 4.4, Building Materials and Building Envelope Performance.
25.1 Copyright © 2017, ASHRAE
25.10 25.11 25.14 25.14 25.14 25.14 25.15 25.16
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25.2 occur, the surrounding space must be filled with a fluid, usually air. If the space is evacuated, heat flow occurs by radiation only. In the context of this discussion, indoor or outdoor heat transfer or surface film coefficient hi or ho relates to an interior or exterior surface of a building envelope assembly. The heat transfer film coefficient is also commonly known as the surface film conductance. Thermal transmittance U is the quantity equal to the steadystate or time-averaged heat flux from the environment on the one side of a body to the environment on the other side, per unit temperature difference between the two environments, in W/(m2 ·K). Thermal transmittance is sometimes called the overall coefficient of heat transfer or U-factor. Average thermal transmittance differs from clear-wall transmittance, in that the former considers all thermal bridge effects in the assembly. Thermal emissivity is the ratio of radiant flux emitted by a surface to that emitted by a black surface at the same temperature. Emissivity refers to intrinsic properties of a material’s surface and is defined only for a specimen of the material that is thick enough to be completely opaque and has an optically smooth surface. Effective emittance E refers to the properties of a particular object. It depends on surface layer thickness, oxidation, roughness, etc.
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Air Air transfer Ma is the time rate of mass transfer by airflow induced by an air pressure difference, caused by wind, stack effect, or mechanical systems, in kg/s. Air flux ma, a vector, is the air transfer through a unit area in the direction perpendicular to that unit area, in kg/(s·m2). Air permeability ka is an intrinsic property of a porous material defined by Darcy’s Law (the equation for laminar flow through porous materials). Air permeability is the quantity of air flux induced by a unit air pressure difference through a unit thickness of homogeneous porous material in the direction perpendicular to the isobaric planes. Units are in kg/(Pa·s·m) or s. Air permeance Ka is the extrinsic quantity equivalent to the time rate of steady-state air transfer through a unit surface of a porous membrane or layer, a unit length of joint or crack, or a local leak induced by a unit air pressure difference over that layer, joint and crack, or local leak. Units are kg/(Pa·s·m2) for a layer, kg/(Pa·s·m) for a joint or crack, or kg/(Pa·s) for a local leak.
Moisture Moisture content w is the amount of moisture per unit volume of porous material, in kg/m3. Moisture ratio X (in mass) or (in volume) is the amount of moisture per unit mass of dry porous material or the volume of moisture per unit volume of dry material, in percent. Specific moisture content is the ratio between a change in moisture content and the corresponding change in driving potential (i.e., relative humidity or suction). Specific moisture ratio is the ratio between a change in moisture ratio and the corresponding change in driving potential (i.e., relative humidity or suction). Water vapor flux mv , a vector, is the time rate of water vapor transfer through a unit area in the direction perpendicular to that unit area, in kg/(s·m2 ). Moisture transfer Mm is the moisture flow induced by a difference in suction or in relative humidity, in kg/s. Moisture flux mm, a vector, is the time rate of moisture transfer through a unit area in the direction perpendicular to that unit area, in kg/(s·m2 ). Water vapor permeability p is the steady-state water vapor flux through a unit thickness of homogeneous material in a direction perpendicular to the isobaric planes, induced by a unit partial water vapor pressure difference, under specified conditions of temperature and relative humidity. Units are kg/(Pa·s·m). When
2017 ASHRAE Handbook—Fundamentals (SI) permeability varies with psychrometric conditions, the specific permeability defines the property at a specific condition. Water vapor permeance M is the steady-state water vapor flux by diffusion through a unit area of a flat layer, induced by a unit partial water vapor pressure difference across that layer, in kg/(Pa·s·m2). Water vapor resistance Z is the reciprocal of water vapor permeance, in ( m2 · s·Pa)/kg. Moisture permeability km is the steady-state moisture flux through a unit thickness of a homogeneous material in a direction perpendicular to the isosuction planes, induced by a unit difference in suction. Units are kg/(Pa·s·m) (suction). Moisture diffusivity Dm is the ratio between the moisture permeability and the specific moisture content, in m2/s.
1.2
HYGROTHERMAL LOADS AND DRIVING FORCES
This section describes the hygrothermal loads acting on the building envelope. That description is used to predict the influence on the hygrothermal behavior of building assemblies, as a basis for design recommendations and moisture control measures (Künzel and Karagiozis 2004). Cooling and heating load estimations for sizing mechanical systems can be found in Chapters 17 and 18. In Figure 1, the loads relevant for building envelope design are presented schematically for an external wall. Generally, they show diurnal and seasonal variations at the exterior surface and mainly seasonal variations at the interior surface. In sunny weather, the exterior wall surface heats by solar radiation, leading to evaporation of moisture from the surface layer. Around sunset, when solar radiation decreases, long-wave (infrared) emission to the clear sky may lead to cooling of the exterior surface below the ambient air temperature, even below the dew-point temperature, so surface condensation may occur. This phenomenon is called undercooling. The exterior surfaces are also exposed to moisture from precipitation and wind-driven rain. Usually, several load cycles overlap (e.g., summer/winter, day/ night, rain/sun). Therefore, a precise analysis of the expected loads should be done before designing any building envelope component. However, the magnitude of the loads is not independent of building geometry and the component’s properties. Analysis of the transient hygrothermal loads is generally based on hourly meteorological
Fig. 1 Hygrothermal Loads and Alternating Diurnal or Seasonal Directions Acting on Building Envelope
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Heat, Air, and Moisture Control in Building Assemblies—Fundamentals
25.3
data, although a shorter time step may be needed. However, determination of local conditions at the envelope’s surface is complicated and requires specific experience. In some cases, computer simulations are necessary to assess the microclimate acting on differently oriented, overhang-protected, or inclined building assemblies.
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Ambient Temperature and Humidity Ambient temperature and humidity, represented by the partial water vapor pressure, are the boundary conditions always affecting both sides of the building envelope. The climate-dependent exterior conditions may show large diurnal and seasonal variations. Therefore, at least hourly data are required for detailed building simulations, though monthly data may suffice in case simple calculation methods are applicable. Chapter 14 provides such meteorological data sets, including temperature and relative humidity, for many locations worldwide. These data sets usually represent average meteorological years based on long-term observations at specific locations. However, data of more extreme climate conditions may be important to assess the risks of moisture damage. Therefore, Sanders (1996) proposed using data of the coldest or warmest year in 10 years for hygrothermal analysis instead of data from an average year. Another method to obtain a severe annual data set concerning the moisture-related damage risk starting from several decades of hourly data has been developed by Salonvaara (2011). This method analyzes the data with respect to their effect on moisture behavior of typical building assemblies. The more severe data sets increase the safety of risk prediction for the service life of building envelope components, but they are less suitable for analyzing the long-term behavior (performance over several years) of constructions because the probability of a sequence of severe years is very low. Also, note that the temperature at the building site may differ from the meteorological reference data when the site’s altitude differs from that of the station recording the data. On average, there is a temperature shift of 0.65 K for every ±100 m. The microclimate around the building may result in an additional temperature shift that depends on the season. For example, the proximity of a lake can moderate seasonal temperature variations, with higher temperatures in winter and lower temperatures in summer compared to sites without water nearby. A low-lying site experiences lower temperatures in winter, whereas city temperatures are higher year round (METEOTEST 2007).
Indoor Temperature and Humidity Indoor conditions depend on the purpose and occupation of the building. For most commercial constructions, temperature and humidity are controlled by HVAC systems with usually well-defined set points. Indoor humidity conditions in residential buildings, however, are often influenced by the outdoor climate and by occupant behavior. (For details on this highly variable vapor release, see Chapter 36.) That water vapor must be removed by ventilation or air conditioning. The resulting relative humidity may be determined by a hygrothermal whole-building simulation or by simple estimation methods using information on moisture production, air change rates, and climate-dependent HVAC operation (TenWolde and Walker 2001). The presence of spas or swimming pools increases the load substantially. Less obvious but sometimes of equal importance are loads from the ground, from penetrating precipitation, or from construction moisture in the building materials. Moisture loads from occupant behavior show an especially transient pattern: they are characterized by peaks (e.g., cooking, showering). Humidity-buffering envelope materials, partition wall materials, and furniture (e.g., carpets, curtains, paper) help to dampen indoor humidity peaks, but they also reduce the moisture removal efficiency of intermittent ventilation (e.g., periodically opening windows, operating ventilation fans). Information on typical indoor climate conditions of specialpurpose constructions such as swimming pools, spas, ice rinks, or
Fig. 2
Solar Vapor Drive and Interstitial Condensation
agricultural buildings and production plants may be found in the 2015 ASHRAE Handbook—HVAC Applications.
Solar Radiation Incident solar radiation is the major thermal load at the building exterior. For direct solar radiation, the resultant irradiation depends on the angle between the sun and the normal on the exposed surface and on its color (short-wave absorptivity). For calculation of incident solar heat flux and spectra, see Chapter 15. For moisture control, solar radiation is usually considered beneficial. However, in some cases solar radiation combined with water from precipitation or other sources (e.g., construction moisture) can lead to severe moisture problems by solar-driven vapor flow. For example, as shown in Figure 2, if the water-absorbing exterior layer of an assembly (e.g., brick veneer, a typical example of “reservoir” cladding) has been wetted by wind-driven rain, solar irradiation creates such high vapor pressure that, in addition to vapor diffusion toward the outdoors, part of the evaporating water diffuses inwards, leading to condensation on and in material layers within the assembly (e.g., sheathing boards, insulation layers, vapor retarders). Adapting the permeance of vapor retarders and weather-resistive barriers (WRB) to the potential loads may improve the situation. ASHRAE research project RP-1091 (Burnett et al. 2004) showed that cladding ventilation is also an effective remedy within specified exterior air humidity limits.
Exterior Condensation Long-Wave Radiant Effects. Long-wave radiation exchange of the envelope surface with the cold layers of the lower atmosphere is a major heat transfer process. At night or with the sun at a low angle, it results in a net heat flux to the sky (i.e., heat energy sink) (see Chapter 15). Depending on the building assembly’s thermal properties, this may lead to a drop in the envelope’s outdoor surface temperature below the ambient air temperature (undercooling). If this surface temperature reaches the air’s dew point, condensation occurs on that exterior surface. Massive structures with a high thermal inertia do not usually lose enough heat to the nighttime radiation sink to bring the outdoor surface temperature below the air’s dew point for a significant period of time. However, many modern building assemblies, such as lightweight roofs or exterior insulation finish systems (EIFSs), have little thermal inertia in their exterior surface layers and are therefore subject to considerable amounts of exterior condensation (Künzel 2007). Interior Temperature Differential. Exterior condensation can also occur on poorly insulated assemblies in cooling climates because of the operation of air-conditioning systems. Repeated
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2017 ASHRAE Handbook—Fundamentals (SI)
exterior condensation or long-lasting, high relative humidity often provides the basis for soiling or microbial growth (fungi or algae), which may not be acceptable even though the durability of the assembly is unlikely to be affected. Effect on Other Layers. Under exterior condensation conditions, ventilated assemblies may also experience condensation within the ventilated air layer. This phenomenon was discovered by investigating pitched roofs with cathedral ceiling insulation (Hens 1992; Janssens 1998; Künzel and Grosskinski 1989). However, damage because of condensation in the ventilation plane is rare, except in metal roofs and ventilated low-mass, low-sloped roofs with moisture-sensitive decks (Hens et al. 2007a, 2007b; Zheng et al. 2004). Occasionally, soiling because of condensate runoff has been reported.
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Wind-Driven Rain The load from rain, especially wind-driven rain, is the main reason for moisture-related building failure. Because the requirements of sometimes costly rain-protection measures depend on the local climate, some countries have introduced regional drivingrain classifications. Generally, coastal regions and those on the windward side of mountains receive the highest precipitation load. Areas of low rainfall do not have the potential for severe winddriven rain. Regional precipitation and wind are significant factors in determining local wind-driven rain load, but local exposure conditions are of equal importance. A building in an open field receives a higher load than one sheltered by a forest or other buildings. A quantification of exposure conditions for walls depending on landscape, neighborhood, and building size and geometry can be found in the British Standard BS 8104 and in the European ISO/DIN Standard 15927-3:2006. The average wind-driven rain load RD in open ground was investigated by Lacy (1965). It is estimated from normal rain RN and the wind velocity component v parallel to the considered orientation: RD = fvRN
(1)
where RD f v RN
= = = =
wind-driven rain intensity, kg/(s·m2) empirical factor = approximately 0.2 s/m mean wind velocity, m/s rain intensity on a horizontal surface in open field, kg/(s·m2)
Figure 3 shows a “rain rose” of results from Equation (1) plotted in polar coordinates indicating the amount of wind-driven rain in mass per unit area hitting an unobstructed and isolated vertical surface in the open field. The driving rain load close to a façade is considerably less than in the open field (as shown in Figure 4), and it becomes irregular. Tops and edges of walls generally receive the highest amount. This
is caused by the airflow pattern around a building (see Chapter 24 for more information). At the windward side, high pressure gradients coincide with large changes in air velocity. The building acts as an obstacle for the wind, slowing down airflow and subsequently reducing the wind-driven rain load near the façade. Gravity and the rain droplets’ momentum prevent them from following the airflow around the building, causing them to strike the façade mainly at the edges of the flow obstacle (Straube and Burnett 2000). However, the irregular driving rain deposition is often evened out by water running off the hard-hit areas, especially when the façade surface has low water absorptivity or the wind-driven rain load is high enough to capillary-saturate the most exposed surface layers. Roof overhangs can reduce the driving rain load on low-rise buildings. Slightly inclined wall sections or protruding façade elements may receive a considerable amount of splash water from façade areas above them, in addition to the direct driving rain deposition. This is often a problem for buildings with walls slightly out of vertical (Künzel 2007). Rain penetrating the exterior cladding of exposed walls may cause severe damage if it cannot be drained and dried out quickly enough. Experience shows that it is almost impossible to seal joints and connections hermetically against wind-driven rain. Therefore, building envelope assemblies should be designed to tolerate a limited amount of water entry (see ASHRAE Standard 160-2009).
Construction Moisture Building damage as a result of migrating construction moisture has become more frequent because tight construction schedules leave little time for building materials to dry. Although often disregarded, construction moisture is either delivered with the building products or absorbed by the materials during storage or construction. Cast-in-place concrete, autoclaved aerated concrete (AAC), calcium silicate brick (CSB), and “green” wood are examples of materials that contain significant volumes of moisture when delivered. Stucco, mortar, clay brick, and concrete blocks are examples of materials that are either mixed or brought into contact with water at the construction site. All other porous building materials may take up considerable amounts of precipitation or groundwater when left unprotected during storage or construction before the enclosure of the building. A single-family house made of AAC may initially contain up to more than 13 Mg of water in its walls. Care is needed to safely remove that water, either by additional ventilation during the first years of operation or by using construction dryers while heating the building before putting it into service. Even “dry” materials have an initial water content of approximately the equilibrium moisture content at 80% rh (EMC80).When significant construction moisture is encountered, EMC80 can be exceeded by a factor of two or more.
Ground- and Surface Water A high groundwater table or surface water running toward the building and filling the loose fill triangle around the basement represents an important moisture load to the lower parts of the building envelope. These loads should be met by grading the
Fig. 4 Fig. 3
Typical Wind-Driven Rain Rose for Open Ground
Measured Reduction in Catch Ratio Close to Façade of One-Story Building at Height of 2 m
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ground away from the building, by perimeter drainage, and by waterproofing the basement and foundation. Instead of waterproofing with bituminous membranes or coatings, water-impermeable structural elements may also be used (e.g., reinforced concrete, which may, however, be vapor permeable). The resulting vapor flux also presents a load that must be addressed (e.g., by basement ventilation). Moisture loads in the ground may impair performance of exterior basement insulation applied on the outside of the waterproofing layer. Therefore, special care must be taken to protect insulation from moisture accumulation unless the insulation material is itself impermeable to water and vapor [e.g., extruded polystyrene (XPS), foam glass]. Wicking of ground- or surface water into porous walls by capillary action is called rising damp. This phenomenon may be a sign of poor drainage or waterproofing of the building’s basement or foundation. However, other phenomena show moisture patterns similar to rising damp. If the wall is contaminated with salts, which is often the case in historic buildings, the wall’s moisture content might stay elevated because of a hygroscopicity increase caused by water uptake by the salt crystals. Another reason for the appearance of rising damp is surface condensation in unheated buildings during summer.
Air Pressure Differentials Wind, mechanical systems, and stack effects (caused by differences between indoor and outdoor temperature) result in air pressure differentials over the building envelope. In contrast to wind, stack pressure is a constant load that may not be neglected. Worse, pressure differentials may drive airflow in the same direction as vapor pressure: from indoors to outdoors during the heating season, and in the opposite direction during the cooling season. Therefore, airflow through cracks, imperfect joints, or air-permeable assembly layers may cause interstitial condensation in a manner similar to vapor diffusion. However, condensation is likely to be more intense and concentrated around leaks in the building envelope. This can become a problem at the top of a building, which may be especially vulnerable because of discontinuities in the air barrier at the parapets. To avoid moisture damage, airflow through and within the building envelope should be prevented by a continuous air barrier. However, it is difficult to guarantee total airtightness, so the hygrothermal effect of airflow can be important, especially when high pressure differentials are expected (e.g., in multistory or mechanically pressurized buildings). For the practical determination of pressure differentials and airflow, see Chapter 16. Air pressures across the envelope may also drive liquid water inward or outward.
2.
HEAT TRANSFER
Heat flow through the building envelope is mainly associated with the building’s energy performance. However, other aspects are equally important. Interior surface temperatures not only serve as an indicator for hygienic conditions in the building (e.g., conditions preventing surface condensation or mold growth), but they can also be a major factor for thermal comfort. Temperature peaks and fluctuations within the building envelope or on its surfaces may further affect the envelope’s durability. At low temperatures, some building materials tend to become less elastic and sometimes brittle, making them vulnerable to strain or mechanical impact. At high temperatures, some materials degrade because of chemical reactions or irreversible deformation. Deformation and local mechanical failure can also occur under the influence of steep temperature gradients or transients. Whereas some of these aspects can be assessed by steady-state calculations (e.g., heating energy losses, energy end use), others require transient simulations for accurate evaluation.
25.5
As explained in Chapter 4, heat transfer by apparent conduction in a solid is governed by Fourier’s law: t t t + k y ----- + k z ----- q = –k grad(t) = – k x ----x y z
(2)
where q = heat flux, W/m2 t = temperature, °C kx, ky , kz = apparent thermal conductivity in direction of x, y, and z axes, W/(m·K) grad(t) = gradient of temperature (change in temperature per unit length, perpendicular to isothermal surfaces in solid), K/m t/x = gradient of temperature along x axis, K/m t/y = gradient of temperature along y axis, K/m t/z = gradient of temperature along z axis, K/m
In Equation (2), the thermal conductivity k of the material is assumed to be directionally dependent. In fact, many building materials (e.g., wood and wood-based materials, mineral fiber insulation, perforated bricks) show considerable anisotropy. Therefore, kx, ky, and kz are not equal in these materials; in isotropic materials, they are equal. Substituting Equation (2) into the relationship for conservation of energy yields h t ------ ----- = div k grad t + S t t t t = ----- k x ----- + ----- k y ----- + ----- k z ----- + S x x y y z z
(3)
where
with
h = enthalpy per unit volume, J/m3 S = heat sources and sinks [e.g., caused by latent heat of evaporation/ condensation in presence of moisture, by chemical reactions such as hydration in concrete, or by phase change from solid to liquid or vice versa of special additives consisting of paraffins or salt hydrates, known as phase-change materials (PCM)], W/m3
h ------ = scs + wcw
(4)
where s cs cw w
= = = =
density of solid (dry material), kg/m3 specific heat capacity of dry solid, J/(kg·K) specific heat capacity of liquid water, J/(kg·K) moisture content, kg/m3
2.1 STEADY-STATE THERMAL RESPONSE In steady state without sources or sinks, Equation (3) reduces to t t t ----- k x ----- + ----- k y ----- + ----- k z ----- = 0 x x y y z z
(5)
If the steady-state heat flux is only in one direction (e.g., perpendicular to the building envelope) and materials are assumed to be isotropic, Equation (2) can be rewritten for each material layer within the building envelope as t 1 q = – k m ------ = – C t = – --- t x R
(6)
where t = temperature difference between two interfaces of one material layer, K x = layer thickness, m km = mean thermal conductivity of material layer with thickness x, W/(m·K) C = thermal conductance of layer with thickness x, W/(m2 ·K) R = thermal resistance of layer with thickness x, (m2 ·K)/W
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25.6 Under steady-state conditions, the one-dimensional heat flux is the same through all material layers, but their individual thermal conductance or resistance is usually different.
Surface-to-Surface Thermal Resistance of a Flat Assembly A single layer’s thermal resistance to heat flow is given by the ratio of its thickness to its apparent thermal conductivity. Accordingly, the surface-to-surface thermal resistance of a flat building assembly composed of parallel layers (e.g., a ceiling, floor, or wall), or a slightly curved component, consists of the sum of the resistances of all layers in series: Rs = R1 + R2 + R3 + R4 + … + Rn (7) where R1, R2, . . ., Rn = resistances of individual layers, (m2 ·K)/W Rs = resistance of building assembly surface to surface (system resistance), (m2 ·K)/W
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For building components with nonuniform or irregular sections, such as hollow clay and concrete blocks, use the R-value of the unit as manufactured.
Combined Convective and Radiative Surface Heat Transfer The surface film resistances and their reciprocal, the surface film coefficients, specify heat transfer to or from a surface by the effects of convection and radiation. Although heat transfer by convection is affected by surface roughness and temperature difference between air and surface, the largest influence is that of air movement, turbulence, and velocity close to the surface. Because air movement at the envelope’s outer surface depends on wind speed and direction, as well as on flow patterns around the building, which are usually unknown, an average surface heat transfer film coefficient at the exterior is normally used. Correlations such as those of Schwarz (1971) link the convective film coefficient to wind speed recorded at a height of 10 m and to orientation of the surface (windward or leeward side). The same holds for the inside surface, where buoyancy plays a prime role. However, because the surface-to-surface thermal resistance of a wall is usually high compared with the surface film resistances, an exact value is of minor importance for most applications. Because air is rather permeable to long-wave radiation, heat transfer by radiation takes place between the surface of the building and the surfaces of objects in the environment, not the surrounding air. Heat transfer by radiation between two surfaces is controlled by the character of the surfaces (emittance and reflectance), the temperature difference between them, and the angle factor through which they see each other. Indoors, the external wall surface exchanges radiation with partition walls, floor, and ceiling, furniture, and other external walls. In winter, most of the other surfaces have a higher temperature than the external wall surface; therefore, radiative exchange gives a net heat flux to the external wall. Outdoors, the external wall surface sees the ground, neighboring buildings, and the sky. Without sun, thermal radiation from the sky and the environment is normally lower than radiation from the wall. This means the wall is losing energy. Especially during clear nights, the temperature of the exterior wall surface may drop below the ambient air temperature. In this case, convective and radiative heat transfer at the surface are opposed to each other. For simplicity, convective and radiative surface heat transfer are often combined, leading to an apparent surface heat transfer film coefficient h: q = h(ten – ts) (8) with
2017 ASHRAE Handbook—Fundamentals (SI) h = hc + hr
(9)
where q = total surface heat transfer, W/m2 h = apparent surface film transfer coefficient, W/(m2 ·K) hr = radiant surface film coefficient to account for long-wave radiation exchange, W/(m2 ·K) hc = convective surface film coefficient, W/(m2 ·K) ten = environmental reference temperature, °C ts = surface temperature, °C
For indoor surface heat transfer, this approach is acceptable when only heat transport through the building envelope is considered. Environmental temperature ten also includes the air temperature as the mean temperature of all surfaces in the field of view of the considered envelope assembly. When all these surfaces are of partition walls and floors that have the same temperature as the indoor air, ten may be replaced by the indoor air temperature. This approach becomes questionable when heat transfer at the outdoor surface is concerned. Because radiation to the sky can lead to surface temperatures below ambient air temperature, Equation (8) underestimates the real heat flux when the environmental temperature is replaced by the outdoor air temperature. Therefore, ten must include all short- and long-wave radiation contributions perpendicular to the assembly’s exterior surface. However, ten cannot be used for moisture transfer calculations. Therefore, a more convenient way may be to treat heat transfer by convection and radiation separately. In this case, hr is skipped in Equation (9), which now applies to convection only, and ten equals the outdoor air temperature. The heat exchange by radiation is calculated by balancing the solar and environmental radiation onto the assembly’s exterior surface with the long-wave emission from it. Steady-state calculation of thermal transport through the building envelope is generally done using surface film resistances based on combined surface heat transfer by radiation and convection, with R being the inverse of the combined surface film coefficient h. Because of greater air movement outdoors, the mean thermal surface film resistance at the exterior surface is lower than at the interior surface. Typical ranges for the combined exterior and interior surface film resistances with surface infrared reflectance 0.1 (nonmetallic) are Ro = 0.03 to 0.06 (m2 ·K)/W Ri = 0.12 to 0.20 (m2 ·K)/W
Heat Flow Across an Air Space Heat flow across an air space is affected by the nature of the boundary surfaces, slope of the air space, distance between boundary surfaces, direction of heat flow, mean temperature of air, and temperature difference between both boundary surfaces. Air space thermal conductance, the reciprocal of the air space thermal resistance, is the sum of a radiation component, a conduction component, and a convection component. For computational purposes, spaces are considered airtight, with neither air leakage nor air washing along the boundary surfaces. The radiation portion depends on the temperature of the two boundary surfaces and their respective surface properties. Assuming infinite parallel plates, radiation is not affected by thickness or slope of the air space, direction of heat flow, or which surface is hot or cold. For surfaces that can be considered ideally gray, the surface properties are emittance, absorptance, and reflectance. Chapter 4 explains all three in depth. For an opaque surface, reflectance is equal to one minus the emittance, which varies with surface type and condition and radiation wavelength. The combined effect of the emittances of the two boundary surfaces is expressed by the effective emittance E of the air space. Table 2 in Chapter 26 lists typical emittance values for reflective surfaces and building materials, and
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25.7 U = 1/RT
(11)
Calculating thermal transmittance requires knowing the (1) apparent thermal resistance of all homogeneous layers, (2) thermal resistance of the nonhomogeneous layers, (3) surface film resistances at both sides of the construction, and (4) thermal resistances of air spaces in the construction. The lower values of the surface film resistances given previously should be used. The steady-state heat flux Qn across the building envelope assembly is then defined by Qn = AnUn(ti – to)
(12)
where ti , to = indoor and outdoor reference temperatures, °C An = component area, m2 Un = U-factor of component, W/(m2 ·K)
Interface Temperatures in a Flat Building Component
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The temperature drop through any layer of an assembly is proportional to its thermal resistance. Thus, the temperature drop tj through layer j is Fig. 5 Heat Flux by Thermal Radiation and Combined Convection and Conduction Across Vertical or Horizontal Air Layer the corresponding effective emittance for air spaces. More exact surface emittance values should be obtained by tests. The convective portion is affected markedly by the slope of the air space, direction of heat flow, temperature difference across the space, and, in some cases, thickness of the space. It is also slightly affected by the mean temperatures of both surfaces. For air spaces in building components, radiation and convection together define total heat flow. An example of their magnitudes for total flow across a vertical or horizontal airspace (up and down) is given in Figure 5. Table 3 in Chapter 26 lists typical thermal resistance values of sealed air spaces of uniform thickness with moderately smooth, plane, parallel surfaces. These data are based on experimental measurements (Robinson et al. 1954). Resistance values for systems with air spaces can be estimated from these results if emittance values are corrected for field conditions. However, for some common composite building insulation systems involving mass-type insulation with a reflective surface in conjunction with an air space, the resistance value may be appreciably lower than the estimated value, particularly if the air space is not sealed or of uniform thickness (Palfey 1980). For critical applications, a particular design’s effectiveness should be confirmed by actual test data undertaken by using the ASTM hot-box method (ASTM Standard C1363). This test is especially necessary for constructions combining reflective and nonreflective thermal insulation.
Total Thermal Resistance of a Flat Building Assembly Total thermal resistance to heat flow through a flat building assembly composed of parallel layers between the environments at both sides is given by RT = Ri + Rs + Ro (10) where Ri = combined inner-surface film resistance, (m2 ·K)/W Ro = combined outer-surface film resistance, (m2 ·K)/W Rs = resistance of building assembly surface to surface, including thermal resistances of possible air layers in component (system resistance), (m2 ·K)/W
Thermal Transmittance of a Flat Building Assembly The thermal transmittance or U-factor of a flat building assembly composed of parallel layers is the reciprocal of RT :
Rj ti – to tj = ----------------------RT
(13)
The temperature in an interface j then becomes (to < ti) j
Ro tj = to + ------ (ti – to) RT
(14)
where R oj is the sum of thermal resistances between inside and interface j in the flat assembly, in (m2 ·K)/W. If the apparent thermal conductivity of materials in a building component is highly temperature dependent, the mean temperature must be known before assigning an appropriate thermal resistance. In such a case, apply successive calculation steps; some software can perform these iterative calculations. First, select the thermal resistances for the particular layers. Then calculate total resistance RT with Equation (9) and the temperature at each interface using Equation (13). The mean temperature in each layer (arithmetic mean of its surface temperatures) can then be used to obtain secondgeneration R-values. The procedure is repeated until the R-values are correctly selected for the resulting mean temperatures. Generally, this demands two or three steps. To calculate interior surface temperatures for risk assessment of surface condensation or mold growth, the higher interior and lower exterior surface film resistance values, given previously, should be used.
Series and Parallel Heat Flow Paths In many building assemblies (e.g., wood-frame construction), components are arranged so that heat flows in parallel paths of different conductances. If no heat flows through lateral paths, the thermal transmittance through each path may be calculated. The average transmittance of the enclosure is then Uav = aUa + bUb + … + nUn
(15)
where a, b, . . . , n are the surface-weighted path fractions for a typical basic area composed of several different paths with transmittances Ua, Ub, . . . , Un. If heat can flow laterally with little resistance in any continuous layer, so that transverse isothermal planes result, the flat construction performs as a series combination of layers, of which one or more provide parallel paths. Total average resistance RT(av) in that case is the sum of the resistance of the layers between the isothermal
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25.8 planes, each layer being calculated and the results weighted by the contributing surface area. For further information, see Chapter 27. The U-factor, assuming parallel heat flow only, is usually lower than that assuming combined series-parallel heat flow. The actual U-factor lies between the two. Without test results, a best choice must be selected. Generally, if the construction contains a layer in which lateral heat conduction is high compared to heat flux through the wall, a value closer to the series-parallel calculation should be used. If, however, there is no layer of high lateral thermal conductance, use a value closer to the parallel calculation. For assemblies with large differences in material thermal conductivities (e.g., assemblies using metal structural elements), the zone method is recommended (see Chapter 27) or the methods discussed in the following section.
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Thermal Bridging and Thermal Performance of Multidimensional Construction Passing highly conductive materials through insulation layers (thermal bridging) results in building envelopes with higher overall thermal transmittances and colder surface temperatures compared to an assembly with continuous, unbroken insulation. Not recognizing the effect of thermal bridging on the building envelope’s thermal performance can lead to inefficient design of HVAC systems, building operation inefficiencies, inadequate condensation resistance at component intersections, and compromised occupant comfort. Heat flow through building envelopes occurs in two and three dimensions when considering all components and their intersections (e.g., glazing, wall, roof, parapet, balconies, floor slabs). Multidimensional heat flow caused by highly conductive thermal bridges (e.g., steel and concrete sections) cannot be effectively evaluated using simplified hand calculations (see Chapter 27) and must be evaluated using a multidimensional computer model or guarded hot-box test measurement (ASTM Standard C1363). Construction details are often lumped into an overall heat flow of the entire opaque area or evaluated separately by defining an effective length or area (or zone of influence). Individual details with transmittances defined by an effective area are combined with other components to calculate an overall thermal transmittance using a weighted average method. However, effective areas often have no real significance or have a large variance that depends on many factors (location of insulation layers in relation to structural framing, insulation levels, orientation of structural framing, predominate heat flow path, etc.). Moreover, the effect of individual details is averaged over the adjacent assemblies, regardless of size of the effective area or length. Consequently, the absolute effect or thermal quality of a detail is difficult to assess using an effective area approach (Morrison Hershfield 2011). Contributions of heat flow for specific construction details (e.g., slab edges, parapets, glazing transitions) are best quantified by determining the extra heat loss caused by an individual detail (i.e., thermal bridge at an intersection of components) above the heat loss of the undisturbed assembly and ascribe that difference to a line or point through their linear or point thermal transmittance. This method can simplify calculation of overall heat loss and highlight the effect of the thermal bridge (Morrison Hershfield 2011).
Linear and Point Thermal Transmittances Using linear and point thermal transmittance requires dividing thermal transmittances into three categories: • Clear field: heat loss if no thermal bridges modified the heat flow through the assembly (area based) • Linear: additional heat loss along a considerable portion of a building perimeter or height in one dimension (e.g., slab edges, balconies, parapets, corner framing, window interfaces)
2017 ASHRAE Handbook—Fundamentals (SI) • Point: additional heat loss from thermal bridges at countable points on a building (e.g., three-way corners, beam penetrations) Calculating the overall heat flow is simply adding the contribution of each linear and point thermal transmittance to the clear-field assembly heat flow. The overall heat flow through the opaque elements of the building envelope (wall or roof) then is Q =
Qanomalies + Qo = L + + Qo
(16)
where Q = overall heat flow through building envelope, W/K Qanomalies = additional heat flow for linear and point transmittance details, W/K Qo = clear-field heat flow without linear and point transmittance details, W/K = linear transmittance, W/(m·K) = point transmittance, W/K L = characteristic length of linear transmittance detail, m
The overall heat flow per unit area, U-value, can be derived by dividing the previous equation by the total projected surface area of the assembly considered.
L + U = ------------------------------------- + Uo A Total
(17)
where U = overall thermal transmittance, including anomalies, W/(m2 ·K) Uo = clear field thermal transmittance (assembly), W/(m2 ·K) Atotal = total opaque projected surface area, m2
Thermal bridging and multidimensional heat flow also affect surface temperatures, concealed surfaces, and surfaces exposed to the indoor and outdoor environments. The temperature distribution from multidimensional heat flow is important to consider for controlling localized dirt pick-up on cold surfaces, mold growth, and condensation. A practical, convenient means to evaluate surface temperatures for multidimensional construction is to represent the coldest surface temperatures of interest relative to a temperature difference. This nondimensional ratio is sometimes referred to as a temperature index, factor, or ratio, with the following basic form but represented by many different symbols (CAN/CSA Standard A440; ISO Standard 13788; Morrison Hershfield 2011): T surface – T outdoor Tindex = --------------------------------------------T indoor – T outdoor
(18)
where Tindex Tsurface Toutdoor Tindoor
= = = =
temperature index coldest temperature of surface outdoor temperature indoor temperature
The temperature index for a critical surface can then be compared to a minimum or design temperature index based on numerous performance criteria (e.g., risk of condensation, mold growth, corrosion). More detailed discussion of using temperature ratios and hygrothermal analysis can be found in the section on Simplified Hygrothermal Design Calculations and Analyses.
2.2
TRANSIENT THERMAL RESPONSE
Steady-state calculations are used to estimate the net heating energy demand on a monthly basis in cold and cool climates. However, in climates where daily temperature swings oscillate around a comfortable mean temperature, transient analysis to define net energy demand for heating and cooling and judge overheating probability is more appropriate. In order of importance, the thermal response of a building to daily swings in temperature and solar
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Heat, Air, and Moisture Control in Building Assemblies—Fundamentals radiation depends on the thermal transmittance and solar heat gain coefficient (SHGC) of transparent components (fenestration) in the envelope, ventilation strategy, accessible thermal capacity of the internal walls and floors, and thermal transmittance/inertia of opaque components in the envelope. The effects of the mutual dependences of these four factors are complex. In cool climates, a simplified approach that accounts for these interactions combines a steady-state daily mean heat balance for a most probable hot day with a lower-limit value for the daily harmonic temperature damping at room level. Temperature damping at room level increases with higher admittance and higher harmonic thermal resistance of opaque envelope components; higher admittance and higher harmonic thermal resistance of all inside walls, floor, and ceiling; and higher thermal inertia of furniture and furnishings. A lower thermal transmittance of transparent components in the envelope and more outdoor air ventilation results in decreased daily harmonic temperature damping at room level. In general, however, and in any climate, whole-building simulations complying with ANSI/ASHRAE Standard 140 are recommended when a clear picture of overheating probability and net energy demand for heating and cooling is needed. High admittances presume the presence of thermal storage materials that are easily assessable for heat. Stored heat can be sensible or latent, as shown in Equation (19). The first requires the use of heavy materials with high and constant capacitance and sufficient thickness to store heat by increasing the temperature of the materials (e.g., bricks, stone, sand-lime stone, concrete).The second uses phase change materials (PCMs), which are materials that store heat by changing phase, typically between solid to liquid. Most models used in building energy simulation programs simulate PCMs by using a temperature-dependent specific heat or enthalpy formulation of Equation (19). dT T cV ------ = kA ----- ------ dt x x
(19)
where c V dT/dt k A T/x
= = = = = = =
density, kg/(m3) specific heat, kJ/(kg·K) volume, m3 gradient of temperature with respect to time, K/s thermal conductivity, W/(m·K) surface area, m3 gradient of temperature along x axis, K/m
There are multiple approaches to solve for transient heat transfer equation. Chapters 4 and 18 describe some approaches to solve for transient problem; Mitchel and Braun (2012) provide more detail. For information about PCM standards and properties, see Chapter 26.
3.
AIRFLOW
Airflow through and within building components is driven by stack pressure, wind pressure, and pressure differentials induced by mechanicals. These driving forces are all described in greater detail in Chapters 16 and 24. In calculating air flux in buildings, a distinction must be made between flow through open porous materials, and that through open orifices such as layers composed of small elements, cavities, cracks, leaks, and intentional vents. Air flux through an open porous material is given by ma = –ka grad(Pa)
(20)
where ma = air flux, kg/(s·m2) ka = air permeability of open porous material, kg/(Pa·s·m) grad(Pa) = gradient in total air pressure (stack, wind, and mechanical systems), (Pa/m
25.9
Fig. 6 Examples of Airflow Patterns The air flux or air transfer equation for flow through the various orifice types is ma or Ma = C(Pa)n
(21)
where the flow coefficient C and flow exponent n are determined experimentally. As shown in Figure 6, there are six simplified single airflow patterns characteristic of flow in buildings: • Exfiltration (air outflow): air passes across an envelope component moving from inside the building to the outdoors • Infiltration (air inflow): air passes across an envelope component from the outdoors to the indoors • Cavity ventilation: outdoor air flows along an air cavity at the exterior of the thermal insulation layer without washing or penetrating the insulation layer • Wind washing: outdoor air permeates the thermal insulation layer and/or flows along the air layer behind • Indoor air washing: indoor air permeates the thermal insulation layer and/or flows along the air layer in front • Air looping: buoyancy forces cause air to flow around and wash the thermal insulation layer filling a cavity In reality, these single patterns never act in isolation but in combination, creating complicated airflow networks along and through building components. These combined flows act to degrade the hygrothermal response of components, envelopes, and even whole building fabrics. For calculating airflow in such cases, Kronvall (1982) developed an equivalent hydraulic network methodology, which was adapted by Janssens (1998) to calculate airflow in lowmass sloped roofs. A single layer with low air permeability (an air barrier) can substantially minimize air inflow and outflow as long as it is both continuous and leak free. An air barrier must also be strong enough to withstand the air pressure difference imposed across the building envelope. This approach can also avoid moisture damage by preventing airflow through the building envelope.
Heat Flux with Airflow Air leakage through building components may undesirably contribute to the ventilation in a building beyond that needed for comfort and indoor air quality (see Chapter 16). Air also carries energy that may degrade a building’s thermal performance. A conditioned building also requires more energy to maintain internal comfort conditions when conditioned air is able to leak out of the building,
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25.10
2017 ASHRAE Handbook—Fundamentals (SI)
and unconditioned air is able to leak into the building through infiltration. Airflow changes the assumption implicit in Equation (1), that no mass flow develops in the solid. In general, the sensible heat (enthalpy) displaced by airflow equals = cMa (t – to)
(22)
where
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c Ma t to
= = = =
specific heat capacity of air, kJ/(kg·K) airflow, kg/s air temperature, °C reference temperature, °C
Only a few simple steady-state cases of combined heat conduction and air-carried enthalpy displacement can be solved analytically. In most cases, testing is the preferred way to get information about the impact. Note that enthalpy flow can increase heat exchange substantially, while reducing temperature damping and time shifting. For example, a full-scale straw bale wall was constructed according to the Tucson, Arizona, structural code with stucco on the exterior side and two layers of 13 mm gypsum board on the interior, with a straw bale thickness of 470 mm. The thermal resistance of the straw by itself was measured as 12.3 (m·K)/W. However, the measured heat flow (in a hot box, tested according to ASTM Standard C1363) was more than twice that expected for the level of thermal resistance. Subsequent dissection of the wall revealed small gaps between the facing surfaces and the straw bales, creating air looping, as shown in Figure 6. A computational fluid dynamics model, using the measured anisotropic air permeability of the straw bales, explored the increased heat transfer through the wall caused by circulation through these gaps. That model found that without the gaps, the wall would have performed as predicted, even considering the relatively high air permeance of the straw itself. However, even very small gaps increased the heat transfer to a value comparable to the experimental measurements. A second wall was built with special attention paid to eliminating these gaps, and the heat transfer fell by 60% (Christian et al. 1998).
4.
MOISTURE TRANSFER
Moisture may enter a building envelope by various paths, including construction moisture, water leaks, wind-driven rain, rising damp, and foundation leaks. Water vapor activates sorption in the
Fig. 7
envelope materials, and water vapor flow in and through the envelope may cause condensation on both nonporous and wet, porous surfaces. Visible and invisible degradation caused by moisture is an important factor limiting the service life of building components. Invisible degradation includes the decrease of thermal resistance of building and insulating materials and the decrease in strength and stiffness of load-bearing materials. Visible degradation includes (1) mold on surfaces, (2) decay of wood-based materials, (3) spalling of masonry and concrete caused by freeze/thaw cycles, (4) hydration of plastic materials, (5) corrosion of metals, (6) damage from expansion of materials (e.g., buckling of wood floors), and (7) decline in appearance. In addition, high moisture levels can lead to odors.
4.1
MOISTURE STORAGE IN BUILDING MATERIALS
Many building materials are porous. The pores provide a large internal surface, which generally has an affinity for water molecules. In some materials, such as wood, moisture may also be adsorbed in the cell wall itself. The amount of water in these hygroscopic (water-attracting) materials is related to the relative humidity of the surrounding air. When relative humidity rises, hygroscopic materials gain moisture (adsorption), and when relative humidity drops, they lose moisture (desorption). The relationship between relative humidity and moisture content at a particular temperature is represented in a graph called the sorption isotherm (Figure 7). Isotherms obtained by adsorption are not identical to those obtained by desorption; this difference is called hysteresis. At high relative humidity, small pores become entirely filled with water by capillary condensation. The maximum moisture content should be reached at 100% rh, when all pores are filled, but experimentally this can only be achieved in a vacuum, by boiling the material, or by keeping it in contact with water for an extremely long time. In practice, the maximum moisture content of a porous material is lower. That value is referred to as free water saturation wf or sometimes capillary moisture content. Figure 7 shows a typical sorption curve, giving the equilibrium moisture content as a function of relative humidity. The equilibrium moisture content increases with relative humidity, especially above 80% rh. It decreases slightly with increasing temperature. Moisture contents above w95 (the equilibrium water content at 95% rh) cannot be
Sorption Isotherms for Porous Building Materials
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Fig. 8 Sorption Isotherm and Suction Curve for Autoclaved Aerated Concrete (AAC) (Künzel and Holm 2001)
achieved solely by vapor adsorption, because this region is characterized by capillary (unbound) water. Chapter 32 describes hygroscopic substances and their use as dehumidifying agents. Chapter 26 has data on the moisture content of various materials in equilibrium with the atmosphere at various relative humidities. Wood and many other hygroscopic materials change dimensions with variations in moisture content. Porous materials also absorb liquid water when in contact with it. Liquid water may be present because of construction moisture, leaks, rain penetration, flooding, or surface and interstitial condensation. Wetting may be so complete that the material reaches free water saturation once the largest pores are filled with water. Up to this point there is still a distinct equilibrium between the moisture content of the material and its environment. This becomes evident when different porous materials are brought in direct (capillary) contact with each other. In that case, there is capillary flow from one material to the other until all pores at a certain size are filled with water in both materials; all pores with sizes above this limit remain empty because smaller capillaries have a higher suction force than larger ones. This phenomenon is used to determine the moisture storage function above 95% rh, which represents the limit of vapor sorption tests in climatic chambers. Dalehaug et al. (2005), Krus (1996), and Roels et al. (2003) described using a pressure plate apparatus, in which water-saturated material samples are placed on a porous membrane permeable to water but impermeable to air. Then pressure is applied in different steps until capillary equilibrium is achieved. The equilibrium moisture content at each pressure step is determined by weighing the samples. The moisture storage function from zero pressure (free water saturation at 100% rh) up to 10 MPa, which corresponds to approximately 93% rh, is defined by plotting the equilibrium water content over the applied pressure (Figure 8), which is assumed to be equal to the suction pressure of the largest still-water-filled capillaries. For a continuous moisture storage function from the dry state to 100% rh, the sorption isotherm and the resultant curve from the pressure plate test are combined, either by converting the suction pressure into relative humidity or vice versa, using Kelvin’s equation:
s = exp – ----------------- w R D T
(23)
where s w RD T
= = = = =
relative humidity of air in pores suction pressure, Pa density of water, kg/m3 gas constant for water vapor, J/(kg·K) absolute temperature, K
The hatched zones in Figure 8 represent the overhygroscopic range where the converted results from pressure plate tests are plotted to complete the sorption isotherm. This narrow range is less important if vapor diffusion is the dominant moisture transport mechanism, for which an approximative interpolation of the moisture storage function between the end of the sorption isotherm and the free water saturation suffices. However, if capillary water flow from one material to the other becomes dominant (e.g., water absorption by bricks from mortar or stucco), the influence of the pressure plate results on the calculation’s outcome may not be negligible (Krus 1996). In that case, the detailed suction curve (Figure 8, right) should be used for simulations.
4.2
MOISTURE FLOW MECHANISMS
Water vapor and liquid water migrate by a variety of transport mechanisms, including the following: • Water vapor diffusion by partial water vapor pressure gradients • Displacement of water vapor by air movement • Surface diffusion and capillary suction of liquid water in porous building materials • Liquid flow by gravity or water and air pressure gradients In the past, moisture control strategies focused on water vapor diffusion. Displacement of water vapor by air movement was treated superficially, and liquid water transport provoked by winddriven rain or soil moisture was overlooked almost completely. When present, however, these mechanisms can move far greater amounts of moisture than diffusion does. Therefore, air movement and liquid flow have a high priority in moisture control.
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2017 ASHRAE Handbook—Fundamentals (SI)
Liquid flow by gravity and by pressure gradients is not discussed here, but a short description of the other mechanisms follows. More comprehensive treatment of moisture transport and storage may be found in Hens (1996), Künzel (1995), and Pedersen (1990). For a discussion of water vapor in air, see Chapter 1.
Water Vapor Flow by Air Movement Air transports not only enthalpy but also the water vapor it contains. Related water vapor flux is represented by 0.62 mv = Wma ---------- ma p Pa
Water Vapor Flow by Diffusion Normally, diffusion moves water vapor through air and building materials, in small quantities. As a driver, it can still be important in industrial applications, such as cold-storage facilities and built-in refrigerators, or in buildings where a high indoor partial water vapor pressure is needed or present because of activities in the space (e.g., in natatoriums). Controlling diffusion also becomes more important with increasingly airtight construction. The equation used to calculate water vapor flux by diffusion through materials is based on Fick’s law for diffusion of a very dilute gas (water vapor) in a binary system (water vapor and dry air): mv = – p grad( p)
(24)
where
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grad( p) = gradient of partial water vapor pressure, Pa p = water vapor permeability of porous material, kg/(Pa·s·m)
According to Equation (24), water vapor flux by diffusion closely parallels Fourier’s equation for heat flux by conduction. However, actual diffusion of water vapor through a material is far more complex than the equation suggests. For hygroscopic materials, water vapor permeability may be a function of relative humidity or, more accurately, moisture content. Also, temperature has an impact. The permeability may even vary spatially or by orientation because of variations or anisotropy in the material’s porous system. Test methods for measuring water vapor permeability are described in ASTM Standard E96. Water vapor flux through a material is determined gravimetrically while maintaining constant temperature and partial water vapor pressure differential across the specimen. Tests are usually done in a climatic chamber at controlled temperature (20 or 23°C) and 50% rh. The material samples are sealed to the top of a cup that contains either a desiccant (dry-cup) or water or a saturated salt solution (wet-cup). Permeability is usually expressed in kg/(Pa·s·m) and permeance in kg/(Pa·s·m2). Whereas permeability refers to the water vapor flux per unit thickness, permeance is used in reference to a material of a specific thickness. For example, a material that is 50 mm thick generally is assumed to have half the permeance of a 25 mm thick material, even though permeances of many materials often are not strictly proportional to thickness. In many cases, the property ignores the effect of cracks or holes in the surface. It is inappropriate to refer to permeability with regard to inhomogeneous or composite materials, such as structural insulated panels (SIPs) or film-faced insulation batts. Methods have been developed that allow measurement of water vapor transport with temperature gradients across the specimen (Douglas et al. 1992; Galbraith et al. 1998; Krus 1996). These methods may give more accurate data on water vapor transfer through materials and eventually allow better distinction between the various transport modes. There are some plastic materials [e.g., polyamide (Künzel 1999)] where the vapor permeability rises substantially with ambient relative humidity because of slight changes in the pore structure: water molecules squeeze between polymer molecules and thereby create new passages through the material. This effect is called solution diffusion. Moisture transport by solution diffusion can be described by Equation (23) using humidity-dependent vapor permeability functions determined by cup tests at several average relative humidity steps.
(25)
where W ma p Pa
= = = =
humidity ratio of moving air air flux, kg/(s·m2) partial water vapor pressure in air, Pa atmospheric air pressure, Pa
Even small air fluxes can carry much larger volumes of water vapor compared to vapor diffusion. However, potentially damaging airflow mostly occurs through cracks and leaky joints rather than through the entire area of a building component. Exceptions include masonry, tiled roofs, slated roofs, mineral and glass wool boards, wood wool, and cement boards.
Water Flow by Capillary Suction Within small pores of an equivalent diameter less than 0.1 mm, molecular attraction between the pore wall and the water molecules causes capillary suction (Figure 9), defined as 2 cos s = -------------------r
(26)
where s r
= = = =
capillary suction, Pa surface tension of water, N/m equivalent radius of capillary, m contact wetting angle, degrees
The contact angle is the angle between the water meniscus and capillary surface. The smaller the contact angle, the larger the capillary suction. In hydrophilic (water-attracting) materials, the contact wetting angle is less than 90°; in hydrophobic (water-repelling) materials, it is between 90 and 180°. Capillary water movement is governed by the gradient in capillary suction s:
Fig. 9 Capillary Rise in Hydrophilic Materials
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Heat, Air, and Moisture Control in Building Assemblies—Fundamentals ml = –km grad(s)
25.13
(27)
where ml = liquid flux, kg/(s·m2) km = water permeability, kg/(Pa·s·m)
Alternatively, with relative humidity as the driving factor [for the conversion, see Kelvin’s Equation (23)]:
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ml = – grad()
(28)
where is the liquid transport coefficient related to the relative humidity as driving potential, in kg/(m · s). Capillary suction is greater in smaller capillaries, so water moves from larger to smaller capillaries. In pores with constant equivalent radius, water moves toward zones with smaller contact angles. Although surface tension is a decreasing function of temperature (the higher the temperature, the lower the surface tension) and water moves toward zones with lower temperature, that effect is small compared to the effect of equivalent pore diameter and contact angle. Capillary suction increases linearly with the inverse of the radius [see Equation (26)], but the flow resistance increases proportionally to the fourth power of the inverse radius. Therefore, larger pores have a much greater liquid transport capacity than smaller pores. Because larger pores can only be filled with water once the smaller pores are saturated, the liquid transport capacity is a function of moisture content. Thus, water permeability km and liquid transport coefficient are also functions of water content. Determination of these functions is, however, quite difficult because it requires the measurement of suction with respect to relative humidity distributions during transient water absorption and drying tests (Plagge et al. 2007). Whereas measuring suction requires experience and special preparation of material samples, determining one-dimensional moisture content distributions in porous building materials can be done accurately with state-of-the-art scanning technologies using nuclear magnetic resonance (NMR), or gamma ray or x-ray attenuation (Krus 1996; Kumaran 1991; van Besien et al. 2002). Transient water content profiles recorded during such scanning tests serve to determine the liquid diffusivity Dw of the examined material, which is defined by ml = –Dw grad(w)
(29)
where w = moisture content, kg/m3 Dw = liquid diffusivity, m2/s
For most hygroscopic building materials, Dw is a function of moisture content. Although Equation (29), which resembles Fick’s law for diffusion, would seem a natural choice for calculating liquid flow, its use is not recommended because water content is not a continuous potential in building envelopes consisting of different materials. Using Equation (27) or (28) is recommended because relative humidity and capillary suction s are considered to be continuous potentials (no jumps at material interfaces). Where diffusivity functions are available, the liquid transport coefficient in Equation (28) can be determined by = Dw dw/d
(30)
where dw/d is the slope of the moisture retention curve, in kg/m3.
Liquid Flow at Low Moisture Content The explanation of liquid flow at low moisture content is still a matter of controversy. Some researchers assume it is surface diffusion (e.g., Krus 1996), whereas others believe liquid flow only fully starts beyond critical moisture content (Carmeliet et al. 1999; Kumaran et al. 2003; Vos and Coelman 1967). Liquid flow begins within
Fig. 10 Moisture Fluxes by Vapor Diffusion and Liquid Flow in Single Capillary of Exterior Wall under Winter Conditions the hygroscopic range, and is often mistaken for a part of vapor diffusion. In porous materials with a fixed pore structure, the apparent increase in vapor permeability during a wet-cup test may be partly because of liquid transport phenomena, and partly to shorter diffusion paths among water islands in the porous system formed by capillary condensation. Surface diffusion is defined as molecular movement of water adsorbed at the pore walls of the material. The driving potential is the mobility of the molecules, which depends on relative humidity in the pores (i.e., the adsorbed water migrates from zones of high to low relative humidity). Liquid flow, if present at low moisture content, can be described by Equations (28) or (29), as for capillary flow. Under isothermal conditions, it is impossible to differentiate between vapor and liquid flow at low moisture content. However, in the presence of a temperature gradient, both transport processes may oppose each other in a pore; the fluxes may go in opposite directions (Künzel 1995). This can be explained by looking at the physical processes in a single capillary going through a wall, as shown in Figure 10. For heating climates in winter, the indoor vapor pressure is usually higher than outdoors while the indoor humidity is lower than outdoors. Therefore, the partial vapor pressure gradient is opposed to the relative humidity gradient over the cross section of a exterior wall. Looking at one capillary in that wall under very dry conditions (Figure 10), the only moisture transport mechanism is vapor diffusion and the total flux is directed towards the exterior. If the average humidity in the wall rises to 50 to 80% rh, liquid water begins to move in the opposite direction either by surface diffusion or by capillary suction in the nanopores. Under these conditions, the total moisture flux may go to zero if both fluxes are of the same magnitude (Krus 1996). When conditions are very wet (e.g., from wind-driven rain), most of the capillary pores are filled with water, and the dominant transport mechanism is flow by capillary suction.
Transient Moisture Flow It is difficult to experimentally distinguish between liquid flow by suction and water vapor flow by diffusion in porous, hygroscopic materials. Because these materials have a very complex porous system and each surface is transversed by liquid-filled pore fractions and vapor-filled pore fractions, vapor and liquid flow are often treated as parallel processes. This allows expression of moisture flow as the summation of the two transport equations, one using
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2017 ASHRAE Handbook—Fundamentals (SI)
water vapor pressure to drive water vapor flow by diffusion, and the other using either capillary suction or relative humidity to drive liquid moisture flow. The conservation equation in that case can be written as w ------- = – div(mw + mv) + Sw t
(31)
present, practitioners can only use simplified tools or hygrothermal models that do not yet cover all airflow-, gravity-, and pressuregradient-induced moisture flow aspects.
6.
SIMPLIFIED HYGROTHERMAL DESIGN CALCULATIONS AND ANALYSES
where w mv mw Sw div
= = = = =
moisture content of building material, kg/m3 water vapor flux, kg/(m2 ·s) liquid water flux, kg/(m2 ·s) moisture source or sink, kg/(m3 ·s) divergence (resulting inflow or outflow per unit volume of solid), m–1
Vapor and liquid fluxes are given by Equations (23), (27), and (28), which may be rewritten in terms of only two driving forces capillary suction pressure s and partial vapor pressure p: w s ------- ----- = div k m grad s + p grad p + Sw s t
(32)
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where s = capillary suction pressure, Pa p = partial vapor pressure, Pa p = vapor permeability (related to partial vapor pressure), kg/(Pa·s·m) km = water permeability (related to partial suction pressure), kg/(Pa·s·m) Sw = moisture source or sink, kg/(m3 ·s)
Alternatively, suction pressure s in Equation (32) can be replaced by relative humidity as the sole variable, with saturation pressure psat only a function of temperature: w (33) ------- ------ = div grad + p grad p sat + Sw where = relative humidity, % psat = saturation vapor pressure, Pa p = vapor permeability (related to partial vapor pressure), kg/(Pa·s·m) = liquid transport coefficient (related to relative humidity), kg/(s·m)
Because of the strong temperature dependence of vapor pressure with respect to saturation vapor pressure, Equation (32) with respect to (33) must be coupled with Equation (3) to describe nonisothermal moisture flow. Under isothermal conditions, Equation (32) with respect to (33) could be solved independently. However, pure isothermal conditions hardly ever exist in reality; as soon as water evaporates or condenses, the latent heat effect leads to temperature differences. Other potentials may be used if material properties appropriate to those potentials are available.
5.
COMBINED HEAT, AIR , AND MOISTURE TRANSFER
Combined heat, air, and moisture transfer can have a detrimental effect on a building. Air in- and exfiltration short-circuit the U-factor as a designed wall performance. Wind washing, indoor air washing, and stack-induced air movement may increment the U-factor by a factor of 2.5 or more. High moisture levels in building materials may also have a negative effect on the building envelope’s thermal performance. Therefore, it is advisable to analyze the combined heat, air, and moisture transfer through building assemblies. However, some of these transport phenomena, especially those involving airflow, are three-dimensional in nature and difficult to predict because they mostly occur through accidental gaps, cracks, or imperfect joints. Research into these effects is ongoing, but at
6.1
SURFACE HUMIDITY AND CONDENSATION
Surface condensation occurs when water vapor contacts a nonporous surface that has a temperature lower than the dew point of the surrounding air. Insulation should therefore be thick enough to ensure that the surface temperature on the warm side of an insulated assembly always exceeds the dew-point temperature there. However, even without reaching the dew point, relative humidity at the surface may become so high that, given enough time, mold growth occurs. According to Hens (1990), a simple design rule is that surface relative humidity in layers warmer than 5°C should not exceed 80% on a monthly mean basis. The temperature ratio fhi is useful for calculating the surface temperature: ts – to f h = ------------i ti – to
(34)
where ts = surface temperature on warm side, °C to = ambient temperature on cold side, °C ti = ambient temperature on warm side, °C
The minimum temperature ratio to avoid surface condensation is t d ,i – t o f h ,min = ----------------i ti – to
(35)
where td,i is the dew point of ambient air on the warm side, °C. The minimum insulation thickness to avoid surface condensation on a flat element can be calculated from f h ,min i Lmin = k --------------------------------- – R add h i 1 – f h ,min
(36)
i
where Radd is the thermal resistance between the surface on the warm side and the cold ambient for the wall without thermal insulation, (m2 ·K)/W. The condensation resistance of glazing is often estimated from outdoor and indoor design temperatures, U-factor of the window assembly, and air film resistance. A window assembly may have different U-factors at the glass, frame, and edge where the glass meets the frame; condensation resistance must be calculated at each of these locations. A procedure for these calculations can be found in NFRC (2004). The likelihood of window condensation depends strongly on the indoor air film resistance. This resistance may be reduced by washing the window with supply air. It may be increased by using window treatments indoors such as blinds or curtains, or by attaching self-adhesive infrared (IR) reflective films. Condensation on glazing is not inherently damaging, unless water is allowed to run onto painted or other surfaces that can be damaged by water.
6.2
INTERSTITIAL CONDENSATION AND DRYING
Dew-Point Method The best-known simple steady-state design tools for evaluating interstitial condensation and drying within exterior envelopes (walls, roofs, and ceilings) are the dew-point method and the Glaser method
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Heat, Air, and Moisture Control in Building Assemblies—Fundamentals (which uses the same underlying principles as the dew-point method, but uses graphic rather than computational methods). These methods assume that steady-state conduction governs heat flow and steadystate diffusion governs water vapor flow. Both analyses compare partial water vapor pressures in the envelope, as calculated by steadystate water vapor diffusion, with saturation water vapor pressures, which are based on calculated steady-state temperatures in the envelope. The condition where the calculated partial water vapor pressure is greater than saturation has been called condensation. Strictly speaking, condensation is the change in phase from vapor to liquid, as occurs on glass, metal, synthetic foils, etc. For porous and hygroscopic building materials (e.g., wood, gypsum, masonry materials), vapor may be adsorbed or absorbed and only forms the droplets usually associated with true condensation when the moisture content passes capillary saturation. Nevertheless, the term condensation is used for this method to indicate vapor pressure in excess of saturation vapor pressure, although this could be misleading about actual water conditions on porous and hygroscopic surfaces. This is one of the unfortunate simplifications inherent in a steady-state analytic tool. Steady-state heat conduction and vapor diffusion impose severe limitations on applicability and interpretation. The greatest one is that the main focus is on preventing sustained interstitial condensation, as indicated by vapor pressures beyond saturation vapor pressures. Many building failures (e.g., mold, buckling of siding, paint failure) are not necessarily related to interstitial condensation; conversely, limited interstitial condensation can often be tolerated, depending on the materials involved, temperature conditions, and speed at which the material dries out. (Drying can only be approximated because both the dew-point and Glaser methods neglect moisture storage and capillary flow.) Because all moisture transfer mechanisms except water vapor diffusion are excluded, results should be considered as approximations and should be used with extreme care. Their validity and usefulness depend on judicious selection of boundary conditions, initial conditions, and material properties. Specifically, the methods should be used to estimate monthly or seasonal mean conditions only, rather than daily or weekly means. Furthermore, water vapor permeances may vary with relative humidity, and rain, flashing imperfections, leaky or poorly formed joints, rain exposure, airflow, and sunshine can have overriding effects. The dew-point and Glaser methods, however, are still used by design professionals and actually form the basis for most codes dealing with moisture control and vapor retarders. For those who want to use this simple tool despite its shortcomings, a description of the dew-point method is presented in this chapter, with two application examples in Chapter 27. A comprehensive description of the dew-point and Glaser methods can be found in TenWolde (1994). The dew-point method uses the equations for steady-state heat conduction and diffusion in a flat component, with the vapor flux in a layer written as p p –mv = p ------ = -----Z d
(37)
where mv p p d Z
= = = = =
water vapor flux through layer of material, ng/(s·m2) partial water vapor pressure difference across layer, Pa water vapor permeability of material, ng/(Pa·s·m) thickness of layer, m water vapor resistance, (Pa·s·m2)/ng
Over time, the dew-point method has been upgraded: (1) the concept of critical moisture content allows accounting for moisture build-up and upgraded calculation of drying, and (2) carried vapor flow has been included, underlining the importance of airtightness to avoid moisture deposition by condensation in building assemblies.
25.15
Calculations have also been based on monthly mean outdoor weather data, corrected for solar gains, long-wave losses, the nonlinear relation between temperature and water vapor saturation pressure, and monthly mean indoor environmental data rather than on daily extremes (Hens 2007, 2016; Vos and Coelman 1967).
7.
TRANSIENT COMPUTATIONAL ANALYSIS
Computer models can analyze and predict the heat, air, and moisture response of building components. These transient models can predict the varying hygrothermal situations in building components for different design configurations under various conditions and climates, and their capabilities are continually improved. Hens (1996) reviewed the state of the art of heat, air, and moisture transport modeling for buildings and identified 37 different models, most of which were research tools that are not readily available and may have been too complex for use by practitioners. Some, however, were available either commercially, free of charge, or through a consultant. Trechsel (2001) provided an update on existing tools and approaches. For many applications and for design guide development, the actual behavior of an assembly under transient climatic conditions must be simulated, to account for short-term processes such as driving rain absorption, summer condensation, and phase changes. Understanding the application limits of such models is an important part of that process. The features of a complete moisture analysis model include transient heat, air, and moisture transport formulation, incorporating the physics of contact conditions between layers and materials. Interfaces may be bridgeable for vapor diffusion, airflow, and gravity or pressure liquid flow only. They may be ideally capillary (no flow resistance from one layer to the next) or behave as a real contact (have an additional capillary resistance at the interface). Not all these features are required for every analysis, though additional features may be needed in some applications (e.g., moisture flow through unintentional cracks and intentional openings, rain penetration through veneer walls and exterior cladding). To model these phenomena accurately, experiments may be needed to define subsystem performance under various loads (Straube and Burnett 1997). It is usually preferable to take performance measurements of system and subsystems in field situations, because only then are all exterior loads and influences captured. Transient models enable timestep-by-timestep analysis of heat, air, and moisture conditions in building components, and give much more realistic results than steady-state conduction/diffusion and conduction/diffusion/airflow models. However, they are complex and usually not transparent, and require judgment and expertise on the part of the user. Existing models are one, two, or three dimensional, requiring the user to devise a realistic representation of the building component to be analyzed. Users should be aware which transport phenomena and types of boundary conditions are included and which are not. For instance, some models cannot handle air transport or rain wetting of the exterior. Results also tend to be very sensitive to the choice of indoor and outdoor conditions. Usually, exact conditions are not known. Indoor and outdoor conditions to be used were established by ASHRAE Standard 160. More extensive data on material properties are available [e.g., Kumaran (2006)], but it can be problematic finding accurate data for all the materials in a component. Validation, verification, and benchmarking of combined heat, air, and moisture models is a formidable task. Currently, only limited internationally accepted experimental data exist. The main difficulty lies in the fact that it is difficult to measure air and moisture fluxes and moisture transport potentials, even under laboratory conditions. In addition, even an already validated model should be verified for each new application.
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25.16 In most full hygrothermal models, common outputs are vapor pressure; temperature; moisture content; relative humidity; and air, heat, and moisture fluxes. Results must be checked for consistency, accuracy, grid independence, and sensitivity to parameter changes. The results may be used to evaluate the moisture tolerance of an envelope system subjected to various interior and exterior loads. Heat fluxes may be used to determine thermal performance under the influence of moisture and airflow. Furthermore, the transient output data may be used for durability and indoor air quality assessment. Postprocessing tools concerning durability (e.g., corrosion, mold growth, freeze and thaw, hygrothermal stress and strain, indoor air humidity) have been developed or are under development. For instance, Carmeliet (1992) linked full hygrothermal modeling to probability-based fracture mechanics to predict the risk of crack development and growth in an exterior insulation finish system (EIFS) by weathering. A transient model to estimate the rate of mold growth was developed by Sedlbauer (2001). Combined heat, air, and moisture models also have limitations. Rain absorption, for example, can be modeled, but rainwater runoff and its consequences at joints, sills, and parapets cannot, although runoff followed by gravity-induced local penetration is one of the main causes of severe moisture problems. Even an apparently simple problem, such as predicting rain leakage through a brick veneer, is beyond many tools’ capabilities. In such cases, simple qualitative schemes and field tests still are the way to proceed (Hens 2007).
7.1
CRITERIA TO EVALUATE HYGROTHERMAL SIMULATION RESULTS
At the building assembly and whole-building level, combined heat, air, and moisture transfer has consequences for thermal comfort, perceived indoor air quality, health, durability, and energy efficiency. Hygrothermal conditions in a building or within a building envelope assembly can be crucial for overall performance of the construction and its mechanical systems. Therefore, simulation results should be compared to limit conditions and widely accepted performance criteria determined for the following performance issues.
Thermal Comfort Thermal comfort, defined as a condition of mind that expresses satisfaction with the thermal environment (ASHRAE Standard 55), depends on two human parameters (clothing and metabolism) and a set of environmental variables, among them relative humidity. At effective temperatures below 25°C, relative humidity’s effect on thermal comfort is minimal, but above 25°C, its importance increases as latent heat loss becomes a main mechanism in getting rid of metabolic heat. If, at those temperatures, the air feels too moist, the thermal environment is perceived as uncomfortable. At low relative humidity, polluted air can irritate the mucosa, and electric discharges when touching insulators (e.g., plastic chairs) are felt. However, in most residential buildings and in many offices, temperature is controlled but not relative humidity, except in hot and humid climates. Its instantaneous value depends on the equilibrium between vapor release indoors, ventilation, airflow among rooms, and temporary vapor storage by finishes and furnishings (often called moisture buffering). The average value over longer periods depends on ventilation and vapor release only, influenced somewhat by moisture buffering.
Perceived Air Quality Air quality may be defined exactly by measuring the pollutants present. However, occupants typically perceive drier, cooler air as smelling “fresher” than humid, warmer air. Thus, temperature and relative humidity affect perception of air freshness. Together, they define the air’s enthalpy. Testing shows that higher enthalpy lowers
2017 ASHRAE Handbook—Fundamentals (SI) the perception of freshness (Fang et al. 1998). Despite this fact, in most buildings, relative humidity is uncontrolled.
Human Health Mold in buildings is of concern to occupants. Mold can grow on most surfaces if the relative humidity at the surface is above a critical value, the surface temperature is conducive to growth, and the substrate provides nutritional value to the organism. The growth rate depends on the magnitude and duration of surface relative humidity. Surface relative humidity is a complex function of material moisture content, local surface temperature, and humidity conditions in the space. In recognition of the issue’s complexity, the International Energy Agency established a surface relative humidity criterion for design purposes: monthly average values should remain below 80% (Hens 1990). Other proposals include the Canada Mortgage and Housing Corporation’s stringent requirement of always keeping surface relative humidity below 65% (CMHC 1999). Although there still is no agreement on which criterion is most appropriate, mold growth can usually be avoided by allowing surface relative humidity over 80% only for short time periods. The relative humidity criterion may be relaxed for nonporous surfaces that are regularly cleaned. Most molds only grow at temperatures above 5°C. Moisture accumulation below 5°C may not cause mold growth if the material is allowed to dry out below the hygroscopic moisture content for a relative humidity of 80% before the temperature rises above 5°C. Mathematical models for predicting a mold growth index were developed by Hukka and Viitanen (1999) and Sedlbauer (2001); these can be linked to results from hygrothermal analysis. Dust mites trigger allergies and asthma. Dust mites thrive at high relative humidities (over 70%) at room temperature, but will not survive sustained relative humidities below 50% (Burge et al. 1994). Note that these values relate to local conditions in the places that mites tend to inhabit (e.g., mattresses, carpets, soft furniture).
Durability of Finishes and Structure Moisture behind paint films may cause paint failure, and water or condensation may also cause streaking or staining. Excessive changes in moisture content of wood-based panels or boards may cause buckling or warp. Excessive moisture in masonry and concrete may cause salt efflorescence, or, when combined with low temperatures, freeze/thaw damage and spalling (chipping). Structural failures caused by wood decay are rare but have occurred (Merrill and TenWolde 1989). Decay generally requires wood moisture content at fiber saturation (usually about 30% by mass) or higher and temperatures between 10 and 40°C. Such high wood moisture contents are possible in green lumber or by absorption of liquid water from condensation, leaks, groundwater, or saturated materials in contact with the wood. To maintain a safety margin, 20% moisture content by mass is sometimes used as the maximum allowable. Because wood moisture content can vary widely with sample location, a local moisture content of 20% or higher may indicate fiber saturation elsewhere. Once established, decay fungi produce water that enables them to maintain moisture conditions conducive to their growth. Rusting of nails, nail plates, or other metal building components is also a potential cause of structural failure. Corrosion may occur at relative humidities near the metal surface above 60% or as a result of liquid water from elsewhere. Wood moisture content over 20% encourages corrosion of steel fasteners in the wood, especially if the wood is treated with preservatives. In buildings, metal fasteners are often the coldest surfaces, encouraging condensation and corrosion.
Energy Efficiency Moisture can significantly degrade thermal performance of most insulation materials. Moisture contributes to heat transfer in both
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sensible and latent forms, as well as through mass transfer. The effect depends on the type of insulation material, moisture content, temperature of the insulation material and its thermal history, location of moisture in the insulation material, and the building envelope’s interior and exterior environments. Reported relationships between thermal performance of the insulation material and moisture content vary significantly. Kyle and Desjarlais (1994) estimated that water distribution accounts for a difference of up to 25% in heat flux in some cases. Evaporation on the warm side and condensation or adsorption on the cold side add important latent heat components to the heat flux (Kumaran 1987). 5Under conditions where water vapor pressure gradients change slowly or where the insulation layer has an extremely low water vapor permeance, little water vapor is transported, but moisture still affects sensible heat transfer in the building envelope component. Epstein and Putnam (1977) and Larsson et al. (1977) showed a nearly linear increase in sensible heat transfer of approximately 3 to 5% for each volume percent increase in moisture content in cellular plastic insulations. For example, an insulation material with about a 5% moisture content by volume has 15 to 25% greater heat transfer than when dry. Other field studies by Dechow and Epstein (1978) and Ovstaas et al. (1983) showed similar results for insulations installed in below-grade applications such as foundation walls.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. ASHRAE. 2010. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2010. ASHRAE. 2007. Method of test for the evaluation of building energy analysis computer programs. ANSI/ASHRAE Standard 140-2007. ASHRAE. 2009. Criteria for moisture-control design analysis in buildings. ANSI/ASHRAE Standard 160-2009. ASTM. 2010. Terminology relating to thermal insulation. Standard C168-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Test method for thermal performance of building materials and envelope assemblies by means of a hot box apparatus. Standard C1363-11. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Test methods for water vapor transmission of materials. Standard E96/E96M-10. American Society for Testing and Materials, West Conshohocken, PA. BSI. 1992. Code of practice for assessing exposure of walls to wind-driven rain. Standard BS 8104:1992. British Standards Institution, London. Burge, H.A., H.J. Su, and J.D. Spengler. 1994. Moisture, organisms, and health effects. Chapter 6 in Moisture control in buildings, ASTM Manual MNL 18. American Society for Testing and Materials, West Conshohocken, PA. Burnett, E., J. Straube, and A. Karagiozis. 2004. Development of design strategies for rainscreen and sheathing membrane performance in wood frame walls. ASHRAE Research Project RP-1091, Final Report. CAN/CSA. 2005. Windows. Standard A440-00 (R2005). Canadian Standards Association, Mississauga, ON. Carmeliet, J. 1992. Durability of fiber-reinforced rendering for exterior insulation systems. Ph.D. dissertation, Catholic University–Leuven, Belgium. Carmeliet, J., G. Houvenaghel, and F. Descamps. 1999. Multiscale network for simulating liquid water and water vapour transfer properties of porous materials. Transport in Porous Materials 35:67-88. Christian, J.E., A.O. Desjarlais, and T.K. Stovall. 1998. Straw bale wall hot box test results and analysis. Thermal Performance of the Exterior Envelopes of Buildings VII, ASHRAE. CMHC. 1999. Best practice guide, wood frame envelopes. Canada Mortgage and Housing Corporation and Canada Wood Council, Ottawa.
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Dalehaug, A., O. Aunronning, and B. Time. 2005. Measurement of water retention properties of plaster: A parameter study of the influence on moisture balance of an external wall construction from variations of this parameter. Proceedings of the 7th Symposium on Building Physics in the Nordic Countries, Reykjavik, pp. 94-101. Dechow, F.J., and K.A. Epstein. 1978. Laboratory and field investigations of moisture absorption and its effect on thermal performance of various insulations. ASTM Special Technical Publication STP 660:234-260. Douglas, J.S., T.H. Kuehn, and J.W. Ramsey. 1992. A new moisture permeability measurement method and representative test data. ASHRAE Transactions 98(2):513-519. Epstein, K.A., and L.E. Putnam. 1977. Performance criteria for the protected membrane roof system. Proceedings of the Symposium on Roofing Technology. National Institute of Standards and Technology, Gaithersburg, MD, and National Roofing Contractors Association, Rosemont, IL. Fang, L., G. Clausen, and P.O. Fanger. 1998. Impact of temperature and humidity on the perception of indoor air quality. Indoor Air 8:80-90. Galbraith, G.H., R.C. McLean, and J.S. Guo. 1998. Moisture permeability data presented as a mathematical relationship. Building Research & Information 20(6):364-372. Hedlin, C.P. 1988. Heat flow through a roof insulation having moisture contents between 0 and 1% by volume, in summer. ASHRAE Transactions 94(2):1579-1594. Hens, H. 1990. Guidelines & practice. International Energy Agency Annex XIV, Leuven, Belgium. Hens, H. 1992. Air/wind tightness of pitched roofs—How they really behave. (In German.) Bauphysik 14(6):161-174. Hens, H. 1996. Heat, air and moisture transfer in highly insulated envelope parts, task 1: Modelling. Final Report, vol. 1, International Energy Agency, Annex 24. Catholic University–Leuven, Laboratorium for Building Physics, Belgium. Hens, H. 2007. Does heat, air moisture modeling really help in solving hygrothermal problems? Proceedings Rakennusfysiikka, Technical University of Tampere, Finland. Hens, H. 2016. Applied building physics: Boundary conditions, building performance, material properties, 2nd ed. Ernst & Sohn, Berlin. Hens, H., A. Janssens, W. Depraetere, J. Carmeliet, and J. Lecompte. 2007a. Brick cavity walls: A performance analysis based on measurements and simulation. Journal of Building Physics 31(2):95-124. Hens, H., F. Vaes, A. Janssens, and G. Houvenaghel. 2007b. A flight over a roof landscape: Impact of 40 years of roof research on roof practices in Belgium. Thermal Performance of the Exterior Envelopes of Whole Buildings X. ASHRAE. Hukka, A., and H. Viitanen. 1999. A mathematical model of mold growth on wooden material. Wood Science and Technology 33(6):475-485. ISO/DIN. 2009. Hygrothermal performance of buildings—Calculation and presentation of climatic data—Part 3: Calculation of a driving rain index for vertical surfaces from hourly wind and rain data. Standard 159273:2009. International Organization for Standardization, Geneva, and Deutsches Institut für Normung, Berlin. Janssens, A. 1998. Reliable control of interstitial condensation in lightweight roof systems. Ph.D. dissertation, Catholic University–Leuven, Belgium. Kronvall, J. 1982. Air flows in building components. Report TVBH-1002. Division of Building Technology, Lund University of Technology, Sweden. Krus, M. 1996. Moisture transport and storage coefficients of porous mineral building materials: Theoretical principles and new test methods. Fraunhofer IRB Verlag, Stuttgart. Kumaran, M.K. 1987. Vapor transport characteristics of mineral fiber insulation from heat flow measurements. In Water vapor transmission through building materials and systems: Mechanisms and measurements. ASTM Special Technical Publication STP 1039:19-27. Kumaran, M.K. 1991. Application of gamma-ray spectroscopy for determination of moisture distribution in insulating materials. Proceedings of the International Centre for Heat and Mass Transfer, pp. 95-103. Kumaran, M.K. 2006. A thermal and moisture transport database for common building and insulating materials (RP-1018). ASHRAE Transactions 112(2):485-497. Kumaran, M., J. Lackey, N. Normandin, F. Tariku, and D. Van Reenen. 2003. Variations in the hygrothermal properties of several wood-based building products. In Research in Building Physics: Proceedings of the
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Second International Conference on Building Physics, Leuven, Belgium, pp. 35-42. J. Carmeliet, H. Hens, and G. Vermeir, eds. Taylor and Francis, London. Künzel, H.M. 1995. Simultaneous heat and moisture transport in building components: One- and two-dimensional calculation using simple parameters. Fraunhofer IRB Verlag, Stuttgart. Künzel, H.M. 1999. More moisture load tolerance of construction assemblies through the application of a smart vapor retarder. Thermal Performance of Exterior Envelopes of Buildings VII, pp. 129-132. ASHRAE. Künzel, H.M. 2007. Factors determining surface moisture on external walls. Thermal Performance of the Exterior Envelopes of Whole Buildings X. ASHRAE. Künzel, H.M., and T. Grosskinski. 1989. Non-ventilated and fully insulated—The best solution for the pitched roof. (In German). Warme- und Kalteschutz im Bau 27. Künzel, H.M., and A. Holm. 2001. Simulation of heat and moisture transfer in construction assemblies. Fraunhofer IBP, Holzkirchen. publica .fraunhofer.de/documents/N-26888.html. Künzel, H.M., and A. Karagiozis. 2004. Vapor control in cold and coastal climate zones. Proceedings of the Canadian Conference on Building Energy Simulation, eSim 2004, pp. 55-60. Kyle, D.M., and A.O. Desjarlais. 1994. Assessment of technologies for constructing self-drying low-slope roofs. Oak Ridge National Laboratory Report ORNL/CON-380. Oak Ridge, TN. Lacy, R.E. 1965. Driving-rain maps and the onslaught of rain on buildings. Proceedings of the RILEM/CIB Symposium on Moisture Problems in Buildings, Helsinki, Finland. Larsson, L.E., J. Ondrus, and B.A. Petersson. 1977. The protected membrane roof (PMR)—A study combining field and laboratory tests. Proceedings of the Symposium on Roofing Technology. National Institute of Standards and Technology, Gaithersburg, MD, and National Roofing Contractors Association, Rosemont, IL. Merrill, J.L., and A. TenWolde. 1989. Overview of moisture-related damage in one group of Wisconsin manufactured homes. ASHRAE Transactions 95(1):405-414. METEOTEST. 2007. Handbook of METEONORM—Global meteorological database for engineers, planners and education. METEOTEST, Bern, Switzerland. Morrison Hershfield. 2011. Thermal performance of building envelope details for mid- and high-rise buildings. ASHRAE Research Project RP1365, Report. NFRC. 2004. Procedure for determining fenestration product condensation resistance values. Technical Document 500. National Fenestration Rating Council, Silver Spring, MD. Ovstaas, G., S. Smith, W. Strzepek, and G. Titley. 1983. Thermal performance of various insulations in below-earth-grade perimeter application. ASTM Special Technical Publication STP 789:435-454. Palfey, A.J. 1980. Thermal performance of low emittance building sheathing. Journal of Thermal Insulation (now Journal of Building Physics) 3:129-141. Pedersen, C.R. 1990. Combined heat and moisture transfer in building constructions. Report 214. Technical University of Denmark. Pedersen-Rode, C., T.W. Petrie, G.E. Courville, P.W. Childs, and K.E. Wilkes. 1991. Moisture migration and drying rates for low slope roofs— Preliminary results. Proceedings of the 3rd International Symposium on Roofing Technology. National Roofing Contractors Association, Rosemont, IL. Plagge, R., G. Scheffler, and A. Nicolai. 2007. Experimental methods to derive hygrothermal material functions for numerical simulation tools. Thermal Performance of Exterior Envelopes of Buildings X. ASHRAE. Robinson, H.E., F.J. Powlitch, and R.S. Dill. 1954. The thermal insulating value of airspaces. Housing and Home Finance Agency, Housing Research Paper 32, U.S. Government Printing Office, Washington, D.C.
Roels, S., J. Carmeliet, and H. Hens. 2003. Hamstad, WP 1: Moisture transfer properties and material characterisation. Final Report (GRD1-19992007), KUL2003-18, Catholic University–Leuven, Belgium. Salonvaara, M. 2011. Environmental weather loads for hygrothermal analysis and design of buildings. ASHRAE Research Project RP-1325, Report. Sanders, C. 1996. Environmental conditions. IEA Annex 24 Report, vol. 2, Catholic University–Leuven, Belgium. Schwarz, B. 1971. Die Wärme- und Stoffübertragung an Außenwandoberflächen. (Heat and mass transfer at exterior wall surfaces.) Dissertation, University of Stuttgart. Sedlbauer, K. 2001. Prediction of mould fungus formation on the surface of and inside building components. Ph.D. dissertation, University of Stuttgart. Shuman, E.C. 1980. Field measurement of heat flux through a roof with saturated thermal insulation and covered with black and white granules. ASTM Special Technical Publication STP 718:519-539. Straube, J., and E. Burnett. 1997. Rain control and screened wall systems. 7th Conference on Building Science and Technology, Durability of Buildings—Design, Maintenance, Codes and Practices, pp. 17-37. Straube J., and E. Burnett. 2000. Simplified prediction of driving rain on buildings. Proceedings of the First International Building Physics Conference, Technische Universiteit–Eindhoven, the Netherlands, pp. 375382. TenWolde, A. 1994. Design tools. Chapter 11 in Moisture control in buildings. ASTM Manual MNL 18. American Society for Testing and Materials, West Conshohocken, PA. TenWolde, A., and I. Walker. 2001. Interior moisture design loads for residences. In Thermal Performance of Exterior Envelopes of Buildings VIII. ASHRAE. Trechsel, H. 2001. Moisture analysis and condensation control in building envelopes. ASTM Manual MNL 40. American Society for Testing and Materials, West Conshohocken, PA. Van Besien, T., S. Roels, and J. Carmeliet. 2002. Experimental determination of moisture: Diffusivity of porous building materials using x-ray radiography. Proceedings of the 6th Nordic Symposium on Building Physics, Trondheim, Norway. Vos, B.H., and E.J.W. Coelman. 1967. Condensation in structures. Report BI-67-33/23, TNO-IBBC, Rijswijk, the Netherlands. Zheng, R., A. Janssens, J. Carmeliet, W. Bogaerts, and H. Hens. 2004. An evaluation of highly insulated cold zinc roofs in a moderate humid climate, Part 2—Corrosion behaviour of zinc sheeting. Construction and Building Materials 18(1):61-71.
BIBLIOGRAPHY ASTM. 2015. Test method for steady-state thermal transmission properties by means of the heat flow meter apparatus. Standard C518. American Society for Testing and Materials, West Conshohocken, PA. Hens, H. 2012. Building physics: Heat, air and moisture—Fundamentals and engineering methods with examples and exercises, 2nd ed. Ernst & Sohn, Berlin. Kumaran, M.K. 1999. Moisture diffusivity of building materials from water absorption measurements. Journal of Thermal Envelope and Building Science 22:349-355. Sedlbauer, K., M. Krus, C. Fitz, and H.M. Künzel. 2011. Reducing the risk of microbial growth on insulated walls by PCM enhanced renders and IR reflecting paints. Proceedings 12DBMC—International Conference on Durability of Building Materials and Components. Tomlinson, J., C. Jotshi, and D. Goswami. 1992. Solar thermal energy storage in phase change materials. Proceedings of Solar ’92: The American Solar Energy Society Annual Conference, Cocoa Beach, FL.
Related Commercial Resources
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Related Commercial Resources CHAPTER 26
HEAT, AIR, AND MOISTURE CONTROL IN BUILDING ASSEMBLIES—MATERIAL PROPERTIES INSULATION MATERIALS AND INSULATING SYSTEMS.............................................................................. Apparent Thermal Conductivity............................................... Materials and Systems ............................................................. AIR BARRIERS ........................................................................ WATER VAPOR RETARDERS ................................................ DATA TABLES......................................................................... Thermal Property Data............................................................
26.1 26.1 26.3 26.5 26.6 26.7 26.7
HIS chapter contains material property data related to the thermal-, air-, and moisture-related performance of building assemblies. The information can be used in simplified calculation methods as applied in Chapter 27, or in software-based methods for transient solutions. Heat transfer under steady-state and transient conditions is covered in Chapter 4, and Chapter 25 discusses combined heat,- air, and -moisture transport in building assemblies. For information on thermal insulation for mechanical systems (including insulation used in a range of temperatures), see Chapter 23. For information on insulation materials used in refrigerant piping systems and cryogenic or low-temperature applications, see Chapters 10 and 47 of the 2014 ASHRAE Handbook—Refrigeration. For properties of materials not typically used in building construction, see Chapter 33 of this volume. Density and thermal properties such as thermal conductivity, thermal resistance, specific heat capacity, and emissivity for longwave radiation are provided for a wide range of building materials, insulating materials, and insulating systems. Air and moisture properties (e.g., air permeance, water vapor permeance or permeability, capillary water-absorption coefficients, sorption isotherms) are given for several materials, with a brief description of how to use the tabulated data. Data on soil thermal conductivity, air cavity resistances, and surface film coefficients, which are also important when considering performance of building assemblies, are also provided.
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T
1.
INSULATION MATERIALS AND INSULATING SYSTEMS
The main purpose of using thermal insulation materials is to reduce conductive, convective, and radiant heat flows. When properly applied in building envelopes, insulating materials do at least one of the following: • Increase energy efficiency by reducing the building’s heat loss or gain • Control surface temperatures for occupant comfort • Help to control temperatures within an assembly, to reduce the potential for condensation • Modulate temperature fluctuations in unconditioned or partly conditioned spaces The primary property of a thermal insulation material is a low apparent thermal conductivity. Additional functions may be served, such as providing support for a surface finish, impeding water The preparation of this chapter is assigned to TC 4.4, Building Materials and Building Envelope Performance.
26.1 Copyright © 2017, ASHRAE
Surface Emissivity and Emittance Data................................... 26.7 Thermal Resistance of Plane Air Spaces ................................. 26.7 Air Permeance Data................................................................. 26.7 Water Vapor Permeance Data ............................................... 26.12 Moisture Storage Data........................................................... 26.13 Soils Data ............................................................................... 26.13 Surface Film Coefficients/Resistances ................................... 26.16 Codes and Standards.............................................................. 26.21
vapor transmission and air leakage into or out of controlled spaces, reducing damage to structures from exposure to fire and freezing conditions, and providing better control of noise and vibration. These functions, of course, should be consistent with the capabilities of the materials. ASTM Standard C168 defines terms related to thermal insulating materials.
1.1
APPARENT THERMAL CONDUCTIVITY
The primary property of a thermal insulation material is a low apparent thermal conductivity, though selection of the appropriate material for a given application also involves consideration of the other performance characteristics mentioned previously. Thermal conductivity (symbol k, in Europe) is a property of a homogeneous, nonporous material. Thermal insulation materials are highly porous, however, with porosities typically exceeding 90%. As a consequence, heat transmission involves conduction in the solid matrix but mainly gas conduction and radiation in the pores (even convection can occur in larger pores). This is why the term apparent thermal conductivity is used. That property is affected by structural parameters such as density, matrix type (fibrous or cellular), and thickness. Each sample of a given insulation material has a unique value of apparent thermal conductivity for a particular combination of temperature, temperature difference, moisture content, and age, a value that is not representative for other conditions. For more details, refer to ASTM Standards C168, C177, C335, C518, C976, and C1045.
Influencing Conditions Density and Structure. Figure 1 shows the variation of the apparent thermal conductivity with density at one mean temperature (i.e., 24°C) for a number of insulation materials used in building envelopes. For most mass-type insulations, there is a minimum that not only depends on the type and form of the material but also on temperature and direction of heat flow. For fibrous materials, the values of density at which the minimum occurs increase as the fiber diameter [or cell size; see Figure 2 (Lotz 1969)] and mean temperature increase. Structural factors also include compaction and settling of insulation, air permeability, type and amount of binder used, additives that influence the bond or contact between fibers or particles, and type and form of the radiation transfer inhibitor, if any. In cellular materials, most factors that influence strength also control the apparent thermal conductivity: size, shape, and orientation of cells, and thickness of cell walls. As Figures 1 and 2 suggest, a specific combination of cell size, density, and gas composition in those materials produces optimum thermal conductivity.
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26.2
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 1 Apparent Thermal Conductivity Versus Density of Several Thermal Insulations Used as Building Insulations
Fig. 2 Variation of Apparent Thermal Conductivity with Fiber Diameter and Density (Lotz 1969)
Temperature. At most normal operating temperatures, the apparent thermal conductivity of insulating materials generally increases with temperature. The rate of change varies with material type and density. Some materials have an inflection in the curve where the blowing agent changes phase from gas to liquid. The apparent thermal conductivity of a sample at one mean temperature (average of the two surface temperatures) only applies to the material at the particular thickness tested. Further testing is required to obtain values suitable for all thicknesses.
Insulating materials that allow a large percentage of heat transfer by radiation, such as low-density fibrous and cellular products, show the greatest change in apparent thermal conductivity with temperature and surrounding surface emissivity. The effect of temperature on structural integrity is unimportant for most insulation materials in low-temperature applications. At very low temperatures, however, some polymeric compounds may undergo glass transition, which is characterized by a marked increase in thermal conductivity. For urethanes and butyl-based compounds, this occurs at approximately –40°C, but for silicones the glass transition temperature is more in the range of –90°C, which is not normally encountered in building applications. In any case, decomposition, excessive linear shrinkage, softening, or other effects limit the maximum suitable temperature for a material. Moisture Content. The apparent thermal conductivity of insulation materials increases with moisture content. If moisture condenses in the insulation, it not only reduces thermal resistance, but it may also physically damage the system, because some insulation materials deteriorate with exposure to water. Most materials would be damaged if moisture were allowed to freeze in the material, because water expands when it freezes. The increase in apparent thermal conductivity depends on the material, temperature, moisture content, and moisture distribution. Section A3 of the CIBSE Guide A (CIBSE 2006) covers thermal properties of building structures affected by moisture. Thickness. Radiant heat transfer in pores of some materials increases the measured apparent thermal conductivity. For low-density insulation (e.g., 5.5 kg/m3), the effect becomes more pronounced with installed thickness) (Pelanne 1979). The effect on thermal resistance is small, even negligible for building applications. No thickness effect is observed in foam insulation. Age. As mentioned previously, most heat transfer in insulation materials at temperatures encountered in buildings and outdoors occurs by conduction through air or another gas in the pores (Lander 1955; Rowley et al. 1952; Simons 1955; Verschoor and Greebler 1952). In fact, heat transfer in dry insulation materials can be closely approximated by combining gas conduction with conduction through the matrix and radiation in the pores, each determined separately. If air in the pores of a cellular insulation material is replaced by a gas with a different thermal conductivity, the apparent thermal conductivity changes by an amount approximately equal to the difference between the thermal conductivity of air and the gas. For example, replacing air with an inert gas can lower the apparent thermal conductivity by as much as 50%. Cellular plastic foams with a high proportion (i.e., more than 90%) of closed cells retain the blowing agent for extended periods of time. Newly produced, they have apparent thermal conductivities of approximately 0.022 W/(m·K) at 24°C. This value increases with time as air diffuses into the cells and the gas gradually dissolves in the polymer or diffuses out. Diffusion rates and increase in apparent thermal conductivity depend on several factors, including permeance of cell walls to the gases involved, foam age, temperature, geometry of the insulation (thickness), and integrity of the surface facing or covering provided. Brandreth (1986) and Tye (1988) showed that aging of unfaced polyurethane and polyisocyanurate is reasonably well understood analytically and confirmed experimentally. The dominant parameters for minimum aging are • • • • • • •
Closed-cell content 90%, preferably 95% Small, uniform cell diameter 1 mm Small anisotropy in cell structure High density Increased thickness High initial pressure of blowing agent in the cells Polymer highly resistant to gas diffusion and solubility
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• Larger proportion of polymer evenly distributed in struts and windows between cells • Low temperature For laminated and spray-applied products, aging is further reduced with higher-density polymer skins, or by well-adhered facings and coverings with low gas and moisture permeance. An oxygen diffusion rate of less than 3.5 mm3/(m2·day) for a 25 m thick facing is one criterion used by some industry organizations for manufacturers of laminated products. Adhesion of the facing must be continuous, and every effort must be made during manufacturing to eliminate or minimize the shear plane layer at the foam/substrate interface (Ostrogorsky and Glicksman 1986). Before 1987, chlorinated fluorocarbons were commonly used as cell gas. Because of their high ozone-depleting potential, chlorofluorocarbons (CFCs) were phased out during the 1990s in accordance with the Montreal Protocol of 1987. Alternatives used today are fluorinated hydrocarbons, CO2, n-pentane, and c-pentane. Closed-cell phenolic-type materials and products, which are blown with similar gases, age differently and much more slowly because of their closed-cell structure. Other Influences. Convection and air infiltration in or through some insulation systems may increase heat transfer. Low-density, loose-fill, large open-cell, and fibrous insulation, and poorly designed or installed reflective systems are the most susceptible. The temperature difference across the insulation and the height and width of the insulated space influence the amount of convection. In some cases, natural convection may be inherent to the system (Wilkes and Childs 1992; Wilkes and Rucker 1983), but in many cases it is a consequence of careless design and/or construction of the insulated structure (Donnelly et al. 1976). Gaps between board- and batt-type insulations lower their effectiveness. Board-type insulation may not be perfectly square, may be installed improperly, and may be applied to uneven surfaces. A 4% void area around batt insulation can produce a 50% loss in effective thermal resistance for ceiling application with R = 3.4 (m2 ·K)/W (Verschoor 1977). Similar and worse results have been obtained for wall configurations (Brown et al. 1993; Hedlin 1985; Lecompte 1989; Lewis 1979; Rasmussen et al. 1993; Tye and Desjarlais 1981). As a solution, preformed joints in boardtype insulation allow boards to fit together without air gaps. Boards and batts can be installed in two layers, with joints between layers offset and staggered. The requirements of ASHRAE Standard 90.1 provide additional guidance on proper installation of insulating materials, as does Chapter 44 in the 2015 ASHRAE Handbook— HVAC Applications. Measurement. Apparent thermal conductivity for insulation materials and systems is obtained by the measuring methods listed in ASTM (2008). These methods apply mainly to laboratory measurements on dried or conditioned samples at specific mean temperatures and temperature gradient conditions. Although fundamental heat transmission characteristics of a material or system can be determined accurately, actual performance in a structure may vary from laboratory results. Only field measurements can clarify the differences. Field-test procedures continue to be developed. Envelope design, construction, and material may all affect the procedure to be followed, as detailed in ASTM (1985a, 1985b, 1988, 1990, 1991).
1.2
MATERIALS AND SYSTEMS
Glass Fiber and Mineral Wool Glass fiber is produced using recycled glass, whereas mineral wool uses diabase stone. Glass and stone are melted, after which a spinning head stretches the melt into fibers with diameter 10 m. These fall through a spray of phenol or silicon binder onto the facings for blankets and batts, which lie on a conveyor belt. The fiber blankets, batts, or boards pass a heated press where the binder
26.3
hardens and the insulation gets its final density and thickness. After passing through the press, the blankets, batts, or boards are cut to size. The spectrum of finished products includes loose fill; over blankets and batts; and soft, semidense, and dense boards. Blankets cannot take any extra load, except their own weight. Dense boards are moderately compression resistant, with a modulus of 10, or about 0.04 to 0.08 MPa. Mineral wool and glass fiber may look similar, but there are important differences. Glass fiber consists of well-ordered, long fibers, whereas mineral wool is composed of unordered, shorter fibers. Glass is also amorphous, whereas diabase stone is crystalline. The thermal conductivity of glass fiber is somewhat lower than for mineral wool (see Table 1), with lower values for higher-density blankets in both materials. Glass and mineral fiber are very vapor permeable. The coefficient of thermal expansion is low for both materials, at ~7 × 10–6 K–1, and irreversible hygrothermal deformation does not occur. The two are also very temperature resistant, although the binder may start evaporating above 250°C and degrades above 600°C for glass fiber and above 850°C for mineral wool (consequently, mineral wool is preferred for high-temperature applications). Both insulation materials are quite moisture tolerant, although wet batts and blankets lose their shape, and the stiffness and compression strength of some dense boards degrade when wet. Glass fibers slowly pulverize when exposed to a combination of high temperature, moisture, and oxygen. Neither glass fiber nor mineral wool burn, but the binder may be combustible. Binder concentrations below 4% simply evaporate, but more concentrated binders can burn. Also, most facing layers are flammable. Glass fiber and mineral wool are widely used insulation materials. Applications range from low-slope roofs (dense boards) and pitched roofs (blankets, batts, and soft boards) to cavity fill (semidense water-repellant boards), timber-frame insulation, exterior insulation finishing systems (EIFS) (dense boards), floor insulation (dense boards), and perimeter insulation (dense boards). Manufacturers modify specific products for many applications, including boards with improved water-repellent properties for fullcavity fill and boards with a dense upper layer for low-slope roofing.
Cellulose Fiber The base material for cellulose fiber is unsold or recycled newspaper and cardboard. Borax salts are added during processing to reduce flammability and mold sensitivity. The material is available in loose form, and is placed into the desired space by wet or dry blowing, which results in densities between 24 to 60 kg/m3, depending on blowing pressure. The material is also available in board form. Cellulose fibers are very hygroscopic, and the borax salt magnifies that property. The fibers are also capillary active, and quite vapor permeable. Avoid loading beyond the mass of a loose fill. Less densely packed cellulose fibers show irreversible settling and are sensitive to hygric swelling and shrinking; wet-blown fibers exhibit shrinkage upon drying. Long-lasting moisture content above 20% (as a percentage of dry mass) should be avoided, because this ultimately leads to decay. Despite the borax salts, cellulose fibers are rather combustible. The borax salts are also not benign: simple exposure may cause respiratory and skin irritation, and ingestion could induce gastrointestinal distress (including nausea, persistent vomiting, abdominal pain, and diarrhea); less common effects on the vascular system and brain are headaches and lethargy. Cellulose fibers are typically used as an alternative for glass fiber and mineral wool. Typical applications are insulation of timberframed walls, and insulation of the ceiling between living spaces and attics. The boards are used for insulating pitched roofs. One important limitation: never apply cellulose fiber (or any insulation, for that matter) where continued long-term wetness may be expected.
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26.4
2017 ASHRAE Handbook—Fundamentals (SI)
Plastic Foams
Cellular Glass
Expanded Polystyrene (EPS). The basic material for EPS is pentane-blown polystyrene pearls. The pearls are first heated above 100°C, at which temperature the evaporating pentane causes expansion. The expanded pearls are then stored for a few days, allowing diffusion of the remaining pentane. Then they are poured into molds and steam heated, so that the expanded pearls coagulate in their own melt. Once cool, the blocks are cut into boards and stored until initial shrinkage ends. EPS is a thermoplastic with a problematic fire reaction: it melts and drips when burning. Consequently, additives are used to slow down flammability. Extruded Polystyrene (XPS). The basic material for XPS is polystyrene pearls, which are melted, blown with a blowing agent, and extruded into boards with a dense skin. The extruded foam is then stabilized in a water bed and cut into separate boards. As with EPS, XPS is stored for several weeks before application. XPS is also a thermoplastic with additives used to slow down flammability. The water vapor resistance of XPS boards is very high, allowing their use in inverted roofs and as perimeter insulation in humid soils. Polyurethane (PUR and Polyisocyanurate (PIR). PUR and PIR are the only insulation materials produced chemically by isocyanate reacting with polyolefin in the presence of a catalyst, a blowing agent, and additives. The difference between the two is the isocyanate ratio: in PIR, this ratio is high enough (60 to 65% kg/kg instead of 50 to 55% kg/kg) to form autopolymers. The main result is a better reaction in combustion. Because the explosive isocyanate/ polyolefin reaction is highly sensitive to temperature and relative humidity, strict control of both parameters is needed. The reaction product is also very sticky, which allows the mixture to be sprayed on many kinds of substrates, or to be used to produce sandwich panels (Figure 3). Once the reaction is finished, the boards are cut to the desired size and stored. R-11 [a CFC with ozone depletion potential (ODP) = 1] was used as a blowing agent until the early 1990s. Since then, blowing agents with zero ODP are preferred, such as hydrofluorocarbons (HFCs) for PUR insulation boards, HFCs or CO2 for spray-applied PUR, and pentane for PIR boards.
Cellular glass is a light, expanded-glass insulation with closedcell pores. It is water- and vaportight, allowing neither vapor diffusion nor capillary suction through the material. Depending on the production process, cellular glass is delivered either as insulating boards or as loose-fill aggregate. The boards are used for roof, wall, basement, and foundation insulation; loose fill is only used for foundation or basement insulation. Cellular glass boards must be protected from frost damage caused by water freezing in its open surface pores. The R-value of cellular glass boards is not affected by moisture, but the thermal resistance of loose fill decreases in moist conditions because of water clinging to the surface of the aggregates.
Capillary-Active Insulation Materials (CAIMs) CAIMs are used as interior wall insulation for existing buildings. Despite being rather vapor permeable, they are applied without a vapor-retarding layer because condensing moisture is supposed to be wicked away from the dew point toward the interior wall surface (Figure 3). In contrast to conventional insulation systems that need a vapor retarder to protect the wall structure from harmful condensation, CAIMs provide condensation control without reducing the drying potential towards the indoors. Because of increasing demand in Europe, several capillary-active insulation systems made of calcium silicate, foamed concrete, or hydrophilic glass fiber have appeared on the market. Tests show that these materials may differ in their wicking ability, but most of them succeed in redistributing condensate by capillary suction. However, even the best-performing CAIM cannot prevent increased relative humidity at the interface between interior insulation and original wall surface of 95% or more in winter. The benefits of CAIMs for insulating existing wall structures are therefore still debatable. Other potential concerns are that their thermal resistances are also generally inferior to those of conventional insulation [k = 0.05 to 0.06 W/(m·K)].
Transparent Insulation Transparent insulation material (TIM) combines transparency for short-wave radiation with low heat conduction, extremely low convection, and opacity for long-wave radiation. The material comprises thin parallel transparent plastic tubes or transparent glass fibers sandwiched between two glass sheets. TIM has a higher thermal conductivity than classic insulation materials [between 0.049 and 0.063 W/(m·K)] but allows solar gains into the conditioned space, so the net heat balance (equilibrium between losses and gains) may be more favorable. Still, use of this material remains limited because of soiling and overheating. The plastic tubes slowly yellow, and, if the space between the two glass sheets is not vaportight, water vapor may diffuse into the panels and condense against the coldest sheet. Dust may enter the TIM boards through spacer leaks and be fixed in the condensate. Also the exterior surface of the panels can become soiled. Overheating is moderated by combining the TIM with solar shading, but this is currently too expensive to be economically viable.
Vacuum Insulation Panels
Fig. 3 Working Principle of Capillary-Active Interior Insulation
Vacuum insulation (sometimes called modified-atmosphere insulation because the interior is not a hard vacuum) is available in rigid and semirigid panels of various sizes. Vacuum insulation panels consist of an interior filler material and an exterior barrier material. Heat conduction through the center of the panel is typically less than 0.007 W/(m·K); some panels have been manufactured with a centerof-panel thermal conductivity less than 0.0025 W/(m·K) However, heat is also transported around the edges of the panel, and that heat transport (often referred to as edge effect) can significantly reduce the thermal resistance of the whole panel compared to the thermal
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties resistance of the center region. For that reason, the resistance of the whole assembly should be considered, and larger panels are generally preferred. In buildings, vacuum panels may be used when the space available for thermal insulation is tightly constricted, such as in historic building retrofits. Vacuum insulation panels are also used in appliances and shipping containers. Vacuum insulation panels rely on reduced gaseous conduction, via reduced air pressure, for their thermal performance and must therefore be protected from puncture or other physical damage. Depending on the permeability of the barrier material and the nature of the seam joining the barrier material around the unit’s perimeter, panels age through air diffusion. To delay this phenomenon, most barrier materials incorporate a very thin metallic layer (often produced using vapor deposition methods). Another way to slow aging is to incorporate getter materials (i.e., any reactive material that absorbs small amounts of gas in an evacuated space) within the panel; some filler materials act as getters themselves. The filler material supports the exterior atmospheric pressure load on the panel and reduces both radiative and gaseous heat transport across the panel. To reduce gaseous heat transfer, voids in the filler material must be smaller than the mean free path of the gas molecules, which is in turn determined by the air pressure in the panel. Therefore, filler materials with finer void sizes retain their heat transfer reduction abilities at higher pressures than fillers with greater void sizes do.
Reflective Insulation Systems Reflective insulation consists of surfaces having high reflectivity (and low emissivity) for long-wave radiation, thus reducing radiant heat transfer. To be effective, these surfaces must face an air layer, or no radiant heat transfer is available to be reduced. Conventional calculation methods ascribe the radiative properties to the facing air layer, rather than to the reflective insulation system, to avoid double-counting the reflective effect. Multiple layers of reflective materials facing smooth and parallel sealed air spaces increase overall thermal resistance, though thermal conductivity will never be less than that of still air. In any case, air exchange and movement must be inhibited or the reduction in radiative heat transfer will be overshadowed by increased convection. Conventional insulation can be combined with reflective surfaces facing air spaces to increase thermal resistance. However, each design must be evaluated, because thermal performance of these systems depends on factors such as condition of insulation, shape and form of construction, means to avoid air leakage and movement, and condition and aging characteristics of reflective surfaces. Values for foil insulation products supplied by manufacturers must be used with caution because they apply only to systems that are identical to the configuration in which the product was tested. In addition, surface oxidation, dust accumulation, condensation, and other factors that change the condition of the low-emittance surface can reduce the thermal effectiveness (Hooper and Moroz 1952). Deterioration results from contact with acidic or basic solutions (e.g., wet cement mortar, preservatives found in decay-resistant lumber). Polluted environments may cause rapid and severe degradation. However, Hooper and Moroz found that site inspections showed a predominance of well-preserved reflective surfaces, with only a small number of cases of rapid and severe deterioration. An extensive review of the reflective building insulation system performance literature is provided by Goss and Miller (1989).
2.
AIR BARRIERS
The main characteristic of an air barrier system is reduced air permeance. To create that performance, the barrier must • Meet material permeability requirements.
26.5
• Be continuous when installed (i.e., tight joints in air barrier assembly; effective bonds in air barrier materials at intersections such as wall/roof, wall/foundation, and wall/windows; tightly sealed penetrations). • Accommodate dimensional changes caused by temperature or shrinkage without damaging joints or air barrier material. • Be strong enough to support stresses applied to air barrier material or assembly. The air barrier must not be ruptured or excessively deformed by wind and stack effect. Where an adhesive is used to complete a joint, the assembly must be designed to withstand forces that might gradually peel away the air barrier material. Where the material is not strong enough to withstand anticipated wind and other loads, it must be supported on both sides to account for positive and negative wind gust pressures. In addition, the following properties can be important, depending on the application: • • • • •
Elasticity Thermal stability Fire and flammability resistance Inertness to deteriorating elements Ease of fabrication, application, and joint sealing
Air barriers may control both vapor and airflow (i.e., they may act as an air/vapor retarder), depending on the characteristics of the materials used. Many designs are based on this idea, with measures taken to ensure that the layer with vapor-retarding properties is continuous to control airflow. Some designs treat airflow and vapor retarders as separate entities, but an airflow retarder should not be where it can cause moisture to condense if it also has vapor-retarding properties. For example, a vapor-retarding air barrier placed on the cold side of a building envelope may cause condensation, particularly if the vapor retarder at the other side of the building is ineffective. Instead, a carefully installed, sealed cold-side air retarder that has sufficient thermal resistance may lower the potential for condensation by raising the temperature at its inside surface during the cold season (Ojanen et al. 1994). Air leakage characteristics can be determined with the ASTM Standard E1186 test method for air barriers on the interior side of the building envelope, and described according to ASTM Standard E1677. Specific air leakage criteria for air barriers in cold heating climates are found in Di Lenardo et al. (1995). These specifications provide classes for air leakage rates of 0.05, 0.10, 0.15, and 0.20 L/(s · m2) when measured with an air pressure difference of 75 Pa, depending on the water vapor permeance of the outermost layer of the building envelope. The highest leakage rate applies if the permeance of the outermost layer is greater than 600 ng/(Pa · s · m2); the lowest rate applies if the permeance is less than 60 ng/(Pa ·s · m2). Intermediate values are also provided. The recommendations apply only to heating climates. The required air permeance of an air barrier material has been set by some building codes at 0.02 L/(s ·m2) at a pressure difference of 75 Pa. A 2008 addendum to ASHRAE Standard 90.1 also references this value. ASTM Standard E1677 provides an alternative minimum air barrier test and criteria specifically suitable for framed walls of low-rise buildings. Air leakage characteristics of an air barrier assembly can be determined with the ASTM Standard E2357 test method, which measures the air leakage of three wall specimens: (1) with the air barrier material installed using air barrier accessories alone, (2) with the air barrier material installed and connected to air barrier components (window, doors, and other premanufactured elements) using air barrier accessories, and (3) with an air barrier wall assembly connected to a foundation assembly and roof assembly using air barrier accessories. The test method reports the air leakage rate at a reference pressure difference of 75 Pa, not because it is necessarily
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26.6 representative of in-service conditions, but because it provides a more accurate measurement that can then be adjusted for actual conditions. Building assemblies are constructed and the various air barrier assemblies are connected to form an air barrier system for the whole building. The building’s air leakage characteristics can be determined with the ASTM Standard E779 test method. A 2008 addendum to ASHRAE Standard 90.1 requires 2.0 L/(s· m2) at 75 Pa pressure difference. The effectiveness of an air barrier is greatly reduced by openings and penetrations, even small ones. These openings can be caused by poor design, poor workmanship during application, insufficient coating thickness, improper caulking and flashing, uncompensated thermal expansion, mechanical forces, aging, and other forms of degradation. Faults or leaks typically occur at electrical boxes, plumbing penetrations, telephone and television wiring, and other unsealed openings in the structure. This is especially true if telephone, television, or other services are installed after the envelope has been inspected and/or tested. A ceiling air barrier should be continuous at chases for plumbing, ducts, flues, and electrical wiring. In flat roofing, mechanical fasteners are sometimes used to adhere the system to the deck, and often penetrate the air barrier. In heating climates, the resulting holes may allow air exfiltration and accompanying water vapor leakage into the roof. ASTM Standard E1186 describes several techniques for locating air leakage sites in building envelopes and air barrier systems. As noted previously, air barrier assemblies must withstand pressures exerted by stack effects, wind, or both during construction and over the building’s life. The magnitude of pressure varies, depending on building type and sequence of construction. At one extreme, single-family dwellings may be built with exterior cladding partly or entirely installed and insulation in place before the air barrier is added. Chimney effects in these buildings are small, even in cold weather, so stresses on the air barrier during construction are small. At the other extreme, wind and chimney effect forces in tall buildings are much greater. A fragile, unprotected sheet material should not be used as an air barrier because it will probably be torn by wind before construction is completed. Calculations of water vapor flow, interstitial condensation, and related moisture accumulation using only water vapor resistances are useless when airflow is involved. More information on air leakage in buildings may be found in Chapter 16.
3. WATER VAPOR RETARDERS The main characteristic of a water vapor retarder is low vapor permeance. The following properties are also important, depending on the application: • Mechanical strength in tension, shear, impact, and flexure • Adhesion • Elasticity Thermal stability • Fire and flammability resistance • Resistance to other deteriorating elements [e.g., chemicals, ultraviolet (UV) radiation] • Ease of fabrication, application, and joint sealing Although a flow of dry air may accelerate drying of a wet building component (Karagiozis and Salonvaara 1999a, 1999b), vapor retarders are completely ineffective without effective airflow control, A single layer may serve both purposes, of course: the designer must assess the needs for control of water vapor and air movement in a building envelope, and devise a system that guarantees the required vapor retarder and air barrier properties. Water vapor retarders demand consideration in every building design. The need for and type of system depend on the climate zone,
2017 ASHRAE Handbook—Fundamentals (SI) construction type, building usage, and moisture sources other than indoor water vapor to be considered. Water vapor retarders were originally designed to protect building elements from water vapor diffusing through building materials and condensing against and in layers at the cold side of the thermal insulation. It is now recognized that it is just as important to allow a building assembly to dry as it is to keep the building assembly from getting wet by vapor diffusion. In some cases, to allow the building assembly to dry, a water vapor retarder may not be needed, or should be semipermeable. In other cases, the environmental conditions, building construction, and building usage may dictate that a material with very low water vapor permeance should be installed to protect building components. A balanced design approach is required: a vapor retarder can reduce the potential for an assembly to dry, but can also reduce the potential for the assembly getting wet. ASHRAE Standard 160 should be followed to determine the need for and placement of a vapor retarder. The 2007 supplement to the International Codes (ICC 2007) lists three water vapor retarder classes: • Class I: 5.7 ng/(Pa ·s · m2) or less • Class II: more than 5.7 ng/(Pa· s · m2) but less than or equal to 57 ng/(Pa· s · m2) • Class III: more than 57 ng/(Pa · s · m2) but less than or equal to 570 ng/(Pa · s · m2) The designer should determine the type of water vapor retarder needed and its location in the envelope assembly, based on climatic conditions, other materials used in the assembly, additional sources of humidity, and the building’s use (e.g., intended relative humidity). A vapor retarder typically slows the rate of water vapor diffusion, but does not totally prevent it. In most cases, requirements for vapor retarders in envelope assemblies are not extremely stringent: because conditions on the inside and outside of buildings vary continually, air movement and ventilation can provide wetting as well as drying at various times, and water vapor entering one side of an envelope assembly can be stored temporarily as hygroscopic moisture and released later. A vapor barrier is often used to try to stop water vapor transport when the real problem is transport of water vapor by air transport. This causes confusion between the use and function of vapor barriers/retarders and those of air barriers. However, if conditions are conducive to excessive humidification, water vapor retarders help to (1) keep the thermal insulation dry; (2) prevent structural damage from rot, corrosion, freeze/thaw, and other environmental actions; and (3) reduce paint problems on exterior walls (although rain absorption through cracks in the paint may be a more probable cause of paint problems) (ASTM Standard C755). Judicious placement of a vapor retarder may also help an assembly to dry out. Another way to look at a vapor retarder is that it is the most vapor-resistant layer in the assembly; a capable designer knows where this layer is and ensures that it does not promote excessive moisture accumulation or prevent the assembly from drying. Therefore, all building envelope assemblies should be assessed to ensure that an unintentional water vapor retarder does not create problems. The vapor retarder’s effectiveness depends on its vapor permeance, installation, and location in the insulation. The retarder should be at or near the surface exposed to higher water vapor pressure and higher temperature. In heating climates, this is usually the winterwarm side. Water vapor retarders are classified as rigid, flexible, or coating materials. Rigid retarders include reinforced plastics, aluminum, and stainless steel. These usually are mechanically fastened in place and are vapor-sealed at the joints. Flexible retarders include metal foils, laminated foil and treated papers, coated felts and papers, and plastic films or sheets. They are supplied in roll form or as an integral part of a building material (e.g., insulation). Accessory materials are required for sealing joints. Coating retarders may be
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties semifluid or mastic; paint (called surface coatings); or hot melt, including thermofusible sheet materials. Their basic composition may be asphaltic, resinous, or polymeric, with or without pigments and solvents, as required to meet design conditions. They can be applied by spray, brush, trowel, roller, dip, or mop, or in sheet form, depending on the type of coating and surface to which it is applied. Potentially, each of these materials is an air barrier; however, to meet air barrier specifications, it must satisfy requirements for strength, continuity, and air permeance. A construction of several materials, some perhaps of substantial thickness, can also constitute a vapor retarder system. In fact, designers have many options. For example, airflow and moisture movement might be controlled using an interior finish, such as drywall, to provide strength and stiffness, along with a low-permeability coating, such as a vapor-retarding paint, to provide the required low permeance. Other designs may use more than one component. However, (1) any component that qualifies as a vapor retarder usually also impedes airflow, and is thus subject to air pressure differences that it must resist; and (2) any component that impedes airflow may also retard vapor movement and promote condensation or frost formation if it is at the wrong location in the assembly. Several studies found a significant increase in apparent permeance as a result of small holes in the vapor retarder. For example, Seiffert (1970) reported a hundredfold increase in the vapor permeance of aluminum foil when it is 0.014% perforated, and a 4000fold increase when 0.22% of the surface is perforated. In general, penetrations particularly degrade a vapor retarder’s effectiveness if it has very low permeance (e.g., polyethylene or aluminum foil). In addition, perforations may lead to air leakage, which further erodes effectiveness. “Smart” vapor retarders allow substantial summer drying while functioning as effective vapor retarders during the cold season. One type of smart vapor retarder has low vapor permeance but conducts liquid water, allowing moisture that condenses on the retarder to dry. Korsgaard and Pedersen (1989, 1992) describe such a retarder composed of synthetic fabric sandwiched between staggered strips of plastic film. The fabric wicks liquid water while the plastic film retards vapor flow. Another type of smart vapor retarder provides low vapor permeance at low relative humidities, but much higher permeance at high relative humidity. During the heating season in cold and moderate climates, the indoor humidity usually is below 50% and the smart vapor retarder’s permeance is low. In the summer or on winter days with high solar heat gains, when the temperature gradient is inward, moisture moving from exterior parts of the wall or roof raises the relative humidity at the vapor retarder. This increases vapor permeance and potential for the wall or roof to dry out. One such vapor retarder is described by Kuenzel (1999). Below 50% rh, the film’s permeance is less than 57 ng/(Pa · s· m2), but it increases above 60% rh, reaching 2050 ng/(Pa · s · m2) at 90% rh.
4. 4.1
DATA TABLES
THERMAL PROPERTY DATA
Steady-state thermal resistances (R-values) of building assemblies (walls, floors, windows, roof systems, etc.) can be calculated from thermal properties of the materials in the component, provided by Table 1, or heat flow through the assembled component can be measured directly with laboratory equipment such as the guarded hot box (ASTM Standard C1363). Direct measurement is the most accurate method of determining the overall thermal resistance for a combination of building materials combined as a building envelope assembly. However, not all combinations may be conveniently or economically tested in this manner. For many simple constructions, calculated R-values (see Chapter 25) agree reasonably well with values determined by hot-box measurement.
26.7
Values in Table 1 were developed by testing under controlled laboratory conditions. In practice, overall thermal performance can be reduced significantly by factors such as improper installation, quality of workmanship and shrinkage, settling, or compression of the insulation (Tye 1985, 1986; Tye and Desjarlais 1983). Good workmanship becomes increasingly important as the insulation requirement becomes greater. Therefore, some engineers include additional insulation or other safety factors based on experience in their design Values in Table 1 are recorded at 24°C, and are intended to be representative values of generic materials. The tabulated thermal conductivities are either relatively constant as tested, or vary over a range of densities. For the most part, thermal conductivity varies directly with density, which provides some guidance for users here a range is presented. A conservative design might use values at the higher end of the range (unless moisture content is a concern, in which case low-conductivity materials might reduce the assembly’s ability to dry out, and would thus be a more conservative choice). References are provided for each material, so users can investigate the as-tested conditions, and additional information regarding the test specimens. Caution: values in Table 1 should not be used without referring to the footnotes, which define limitations and some of the as-tested conditions for the materials listed. Because commercially available materials vary, not all values apply to specific products.
4.2
SURFACE EMISSIVITY AND EMITTANCE DATA
Table 2 provides measured long-wave emissivities for various surfaces, which are used to characterize radiant heat transfer to or from these surfaces. To simplify radiant heat transfer calculations, the combined emittance for two surfaces is also provided, although these values can be calculated using eff = 1/(1/1 + 1/2 – 1). As described previously, surface oxidation, dust accumulation, condensation, and other factors can impair the emissivity of highly reflective surfaces, so slightly higher values should be used.
4.3
THERMAL RESISTANCE OF PLANE AIR SPACES
Table 3 provides effective resistance values for plane (i.e., generally flat) air spaces that are enclosed within an assembly. Where an assembly incorporates reflective insulation, the effect of the reflective surface is ascribed to the air space, not to the material component. It should be understood that the reflective surface must face an air space to have any effect in reducing thermal transmittance, and assigning the value of the reflective surface to the air space in a design calculation reinforces this concept. Note that “reflective insulation systems” are bounded by an enclosed air space within an assembly, whereas “radiant barrier systems” feature a reflective surface facing an open airspace. Reflective insulation may be described as modifying the effective R-value of the assembly, but a radiant barrier system may not. This includes reflective surfaces behind siding, which should not be considered as “reflective insulation” (in most cases, the nature of the heat transfer will be dominated by wind-driven convection, rather than radiant exchange). Thermal resistance values for siding with reflective foil backing are provided in Table 1.
4.4
AIR PERMEANCE DATA
Table 4 provides measured air permeability of different materials, tested in accordance with Bomberg and Kumaran (1986), to be used in assessing the suitability of these materials in an air barrier assembly. As discussed previously, low air permeance is not sufficient to
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26.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 1
Building and Insulating Materials: Design Valuesa
Description
Density, kg/m3
Conductivityb k, Resistance R, Specific Heat, kJ/(kg·K) Referenceo W/(m·K) (m2 ·K)/W
Insulating Materials Blanket and battc,d Glass-fiber batts...................................................................
Rock and slag wool batts.....................................................
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Mineral wool, felted ............................................................
7.5 to 8.2 9.8 to 12 13 to 14 22 — 32 to 37 45 16 to 48 16 to 130
Board and slabs Cellular glass ....................................................................... 120 Cement fiber slabs, shredded wood with Portland cement binder............................................................................. 400 to 430 with magnesia oxysulfide binder................................... 350 Glass fiber board.................................................................. — 24 to 96 Expanded rubber (rigid) ...................................................... 64 Extruded polystyrene, smooth skin ..................................... — aged per CAN/ULC Standard S770-2003..................... 22 to 58 aged 180 days ................................................................ 22 to 58 European product........................................................... 30 aged 5 years at 24°C ...................................................... 32 to 35 blown with low global warming potential (GWP) (<5) blowing agent ............................................................. Expanded polystyrene, molded beads ................................. — 16 to 24 29 Mineral fiberboard, wet felted............................................. 160 Rock wool board.................................................................. — floors and walls.............................................................. 64 to 130 roofing ........................................................................... 160 to 180. Acoustical tileg .................................................................... 340 to 370 Perlite board......................................................................... 140 Polyisocyanurate.................................................................. — unfaced, aged per CAN/ULC Standard S770-2003 ...... 26 to 37 with foil facers, aged 180 days ...................................... — — Phenolic foam board with facers, agedf............................... Loose fill Cellulose fiber, loose fill ..................................................... — attic application up to 100 mm ...................................... 16 to 19 attic application > 100 mm ............................................ 19 to 26 wall application, dense packed ...................................... 56 Perlite, expanded ................................................................. 32 to 64 64 to 120 120 to 180 Glass fiberd attics, ~100 to 600 mm ................................................. 6.4 to 8.0 attics, ~600 to 1100 mm ................................................ 8 to 9.6 closed attic or wall cavities............................................ 29 to 37 Rock and slag woold attics, ~90 to 115 mm .................................................... 24 to 26 attics, ~125 to 430 mm .................................................. 24 to 29 closed attic or wall cavities............................................ 64 Vermiculite, exfoliated ........................................................ 112 to 131 64 to 96 Spray-applied Cellulose, sprayed into open wall cavities .................... 26 to 42 Glass fiber, sprayed into open wall or attic cavities...... 16 29 to 37 Polyurethane foam .............................................................. — low density, open cell .................................................... 7.2 to 10 medium density, closed cell, aged 180 days.................. 30 to 51
0.046 to 0.048 0.040 to 0.043 0.037 to 0.039 0.033 — 0.036 to 0.037 0.033 to 0.035 0.040 0.035
— — — — — — — — —
0.8 — — — — 0.8 — — — —
Kumaran (2002) Four manufacturers (2011) Four manufacturers (2011) Four manufacturers (2011) Four manufacturers (2011) Kumaran (1996) One manufacturer (2011) One manufacturer (2011) CIBSE (2006), NIST (2000) NIST (2000)
0.042
—
0.8
One manufacturer (2011)
0.072 to 0.076 0.082 — 0.033 to 0.035 0.029 — 0.026 to 0.029 0.029 0.030 0030
— — — — — — —
— 1.3 0.8 — 1.7 1.5 —
—
—
0.035 to 0.036 — 0.035 to 0.037 0.033 0.037 — 0.033 to 0.036 0.039 to 0.042 0.052 to 0.053 0.052 — 0.023 to 0.025 0.022 to 0.023 0.020 to 0.023
— — — — — — — — — — — — — —
— 1.5 — — 0.8 0.8 — 0.8 0.6 to 0.8 — 1.5 — — —
One manufacturer (2011) Kumaran (1996) Independent test reports (2008) Independent test reports (2008) Kumaran (1996) Kumaran (1996) Five manufacturers (2011) Five manufacturers (2011)
— 0.045 to 0.046 0.039 to 0.040 0.039 to 0.040 0.039 to 0.045 0.045 to 0.052 0.052 to 0.061
— — — — — — —
1.4 — — — 1.1 — —
NIST (2000), Kumaran (1996) Four manufacturers (2011) Four manufacturers (2011) One manufacturer (2011) (Manufacturer, pre 2001) (Manufacturer, pre 2001) (Manufacturer, pre 2001)
0.052 to 0.055 0.049 to 0.052 0.035 to 0.036
— — —
— — —
Four manufacturers (2011) Four manufacturers (2011) Four manufacturers (2011)
0.049 0.046 to 0.048 0.039 to 0.042 0.068 0.063
— — — — —
— — — 1.3 —
Three manufacturers (2011) Three manufacturers (2011) Three manufacturers (2011) Sabine et al. (1975) Manufacturer (pre 2001)
0.039 to 0.040 0.039 to 0.042 0.033 to 0.037 — 0.037 to 0.042 0.020 to 0.029
— — — — — —
— — — 1.5 — —
Two manufacturers (2011) Manufacturers’ association (2011) Four manufacturers (2011) Kumaran (2002) Three manufacturers (2011) Five manufacturers (2011)
Kumaran (1996) One manufacturer (2011) Nottage (1947) Kumaran (1996) Four manufacturers (2011) One manufacturer (2011) One manufacturer (2011) One manufacturer (2011)
One manufacturer (2010) Kumaran (1996) Seven manufacturers (2011) Two manufacturers (2011) One manufacturer (2011)
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties Table 1
Building and Insulating Materials: Design Valuesa (Continued)
Description Building Board and Siding Board Asbestos/cement board........................................................ Cement board ...................................................................... Fiber/cement board..............................................................
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26.9
Gypsum or plaster board ..................................................... Oriented strand board (OSB).............................9 to 11 mm ......................................................................... 12.7 mm Plywood (douglas fir) ............................................ 12.7 mm ......................................................................... 15.9 mm Plywood/wood panels............................................ 19.0 mm Vegetable fiber board sheathing, regular density................................ 12.7 mm intermediate density .................................. 12.7 mm nail-based sheathing ........................................ 12.7 mm shingle backer.................................................... 9.5 mm sound deadening board .................................... 12.7 mm tile and lay-in panels, plain or acoustic laminated paperboard homogeneous board from repulped paper Hardboard medium density ............................................................. high density, service-tempered grade and service grade high density, standard-tempered grade.......................... Particleboard low density .................................................................... medium density ............................................................. high density ................................................................... underlayment .................................................. 15.9 mm Waferboard .......................................................................... Shingles Asbestos/cement ............................................................ Wood, 400 mm, 190 mm exposure ............................... Wood, double, 400 mm, 300 mm exposure .................. Wood, plus ins. backer board .............................. 8 mm Siding............................................................................. Asbestos/cement, lapped .................................. 6.4 mm Asphalt roll siding ......................................................... Siding Asphalt insulating siding (12.7 mm bed) ...................... Hardboard siding ............................................... 11 mm Wood, drop, 200 mm......................................... 25 mm Wood, bevel 200 mm, lapped .......................................... 13 mm 250 mm, lapped .......................................... 19 mm Wood, plywood, lapped.................................... 9.5 mm Aluminum, steel, or vinyl,j,k over sheathing ................. hollow-backed ....................................................... insulating-board-backed ........................... 9.5 mm foil-backed................................................ 9.5 mm insulated vinyl siding........................ 19 to 32 mm Architectural (soda-lime float) glass Building Membrane Vapor-permeable felt........................................................... Vapor: seal, 2 layers of mopped 0.73 kg/m2 felt................. Vapor: seal, plastic film....................................................... Finish Flooring Materials Carpet and rebounded urethane pad ........................ 19 mm Carpet and rubber pad (one-piece) ........................ 9.5 mm Pile carpet with rubber pad......................... 9.5 to 12.7 mm Linoleum/cork tile .................................................. 6.4 mm PVC/rubber floor covering .................................................. rubber tile .......................................................... 25 mm terrazzo .............................................................. 25 mm
Density, kg/m3
Conductivityb k, Resistance R, Specific Heat, kJ/(kg·K) Referenceo W/(m·K) (m2 ·K)/W
1900 1150 1400 1000 400 300 640 650 650 460 540 450 650 290 350 400 290 240 290 480 480
0.57 0.25 0.25 0.19 0.07 0.06 0.16 — — — — — — — — — — — 0.058 0.072 0.072
— — — — — — — 0.11 0.12 0.14 0.15 0.19 0.11 0.23 0.19 0.19 0.17 0.24 — — —
1.00 0.84 0.84 0.84 1.88 1.88 1.15 1.88 1.88 1.88 1.88 1.88 1.88 1.30 1.30 1.30 1.30 1.26 0.59 1.38 1.17
Nottage (1947) Kumaran (2002) Kumaran (2002) Kumaran (1996) Kumaran (1996) Kumaran (1996) Kumaran (2002) Kumaran (2002) Kumaran (2002) Kumaran (2002) Kumaran (2002) Kumaran (2002) Kumaran (2002) Lewis (1967) Lewis (1967)
800 880 1010
0.105 0.12 0.144
— — —
1.30 1.34 1.34
Lewis (1967) Lewis (1967) Lewis (1967)
590 800 1000 640 700
0.102 0.135 1.18 — 0.072
— — — 1.22 —
1.30 1.30 — 1.21 1.88
Lewis (1967) Lewis (1967) Lewis (1967) Lewis (1967) Kumaran (1996)
1900 — — —
— — — —
0.037 0.15 0.21 0.25
— —
— —
0.037 0.026
— 1.30 1.17 1.30 — 1.01 1.47
— —
— — —
0.26 — 0.14
1.47 0.12 1.17
— — —
— — —
1.17 1.17 1.22
— — —
— — —
2500
1.0
0.14 0.18 0.10 — 0.11 0.32 0.52 0.35 to 0.48 —
— — —
— — —
0.011 0.21 Negligible
— — —
110 320 290 465 — 1900 —
— — — — 0.40 — —
0.42 0.12 0.28 0.09 — 0.06 0.014
— — — — — — 0.80
Lewis (1967)
1.17
1.22i 1.34 — 0.84
NIST (2000) NIST (2000) NIST (2000) NIST (2000) CIBSE (2006) NIST (2000)
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26.10
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Building and Insulating Materials: Design Valuesa (Continued)
Description
Density, kg/m3
Conductivityb k, Resistance R, Specific Heat, kJ/(kg·K) Referenceo W/(m·K) (m2 ·K)/W
Metals (See Chapter 33, Table 3) Roofing Asbestos/cement shingles.................................................... Asphalt (bitumen with inert fill) .........................................
Licensed for single user. © 2017 ASHRAE, Inc.
Asphalt roll roofing ............................................................. Asphalt shingles................................................................... Built-up roofing ....................................................... 10 mm Mastic asphalt (heavy, 20% grit) ........................................ Reed thatch .......................................................................... Roofing felt.......................................................................... Slate ......................................................................... 13 mm Straw thatch ......................................................................... Wood shingles, plain and plastic-film-faced ....................... Plastering Materials Cement plaster, sand aggregate ........................................... Sand aggregate......................................................... 10 mm ........................................................................... 20 mm Gypsum plaster.................................................................... Lightweight aggregate ............................................. 13 mm ............................................................................ 16 mm on metal lath ...................................................... 19 mm Perlite aggregate .................................................................. Sand aggregate..................................................................... on metal lath ...................................................... 19 mm Vermiculite aggregate .........................................................
Perlite plaster ....................................................................... Pulpboard or paper plaster................................................... Sand/cement plaster, conditioned........................................ Sand/cement/lime plaster, conditioned................................ Sand/gypsum (3:1) plaster, conditioned .............................. Masonry Materials Masonry units Brick, fired clay ...................................................................
Clay tile, hollow 1 cell deep.......................................................... 75 mm .................................................................... 100 mm 2 cells deep ...................................................... 150 mm .................................................................... 200 mm .................................................................... 250 mm 3 cells deep ...................................................... 300 mm Lightweight brick ................................................................ Concrete blocksh.i Limestone aggregate ~200 mm, 16.3 kg, 2200 kg/m3 concrete, 2 cores ......... with perlite-filled cores............................................ ~300 mm, 25 kg, 2200 kg/m3 concrete, 2 cores ............ with perlite-filled cores............................................ Normal-weight aggregate (sand and gravel) ~200 mm, 16 kg, 2100 kg/m3 concrete, 2 or 3 cores… with perlite-filled cores............................................
1920 1600 1900 2300 920 920 920 950 270 2250 — 240 —
— 0.43 0.58 1.15 — — — 0.19 0.09 1.20 — 0.07 —
0.037 — — — 0.027 0.078 0.059 — — — 0.009 — 0.166
1.00 — — — 1.51 1.26 1.47 — — — 1.26 — 1.30
1860 — — 1120 1280 — 720 1680 — 480 600 720 840 960 400 600 600 1560 1440 1550
0.72 — — 0.38 0.46 720 720 — 0.22 0.81 — 0.14 0.20 0.25 0.26 0.30 0.08 0.19 0.07 0.63 0.48 0.65
— 0.013 0.026 — — — — 0.083 — — 0.023 — — — — — — — — — — —
0.84 0.84 0.84 — — 0.056 0.066 — 1.34 0.84 — — — — — — — — — — — —
CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006)
2400 2240 2080 1920 1760 1600 1440 1280 1120
1.21 to 1.47 1.07 to 1.30 0.92 to 1.12 0.81 to 0.98 0.71 to 0.85 0.61 to 0.74 0.52 to 0.62 0.43 to 0.53 0.36 to 0.45
— — — — — — — — —
— — — 0.80 — — — — —
Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988)
— — — — — — 800 770
— — — — — — 0.20 0.22
0.14 0.20 0.27 0.33 0.39 0.44 — —
0.88 — — — — — — —
Rowley and Algren (1937) Rowley and Algren (1937) Rowley and Algren (1937) Rowley and Algren (1937) Rowley and Algren (1937) Rowley and Algren (1937) Kumaran (1996) Kumaran (1996)
— — — —
— — —
— 0.37 — 0.65
— — — —
— —
— —
0.20 to 0.17 0.35
0.92 —
CIBSE (2006) CIBSE (2006) CIBSE (2006)
CIBSE (2006) CIBSE (2006) CIBSE (2006) CIBSE (2006)
CIBSE (2006) CIBSE (2006) — —
Valore (1988) Valore (1988) Van Geem (1985) Van Geem (1985)
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties Table 1
Licensed for single user. © 2017 ASHRAE, Inc.
Description
26.11
Building and Insulating Materials: Design Valuesa (Continued) Density, kg/m3
Conductivityb k, Resistance R, Specific Heat, kJ/(kg·K) Referenceo W/(m·K) (m2 ·K)/W
with vermiculite-filled cores ................................... — — — — ~300 mm, 22.7 kg, 2000 kg/m3 concrete, 2 cores ....... Medium-weight aggregate (combinations of normal and lightweight aggregate) — ~200 mm, 13 kg, 1550 to 1800 kg/m3 concrete, 2 or 3 cores — with perlite-filled cores ........................................... — — with vermiculite-filled cores ................................... — — with molded-EPS-filled (beads) cores..................... — — with molded EPS inserts in cores ............................ — — Low-mass aggregate (expanded shale, clay, slate or slag, pumice) — — ~150 mm, 7 1/2 kg, 1400 kg/m2 concrete, 2 or 3 cores with perlite-filled cores ........................................... — — with vermiculite-filled cores ................................... — — — — 200 mm, 8 to 10 kg, 1150 to 1380 kg/m2 concrete ..... with perlite-filled cores ........................................... — — with vermiculite-filled cores ................................... — — with molded-EPS-filled (beads) cores..................... — — with UF foam-filled cores ....................................... — — with molded EPS inserts in cores ............................ — — — — 300 mm, 16 kg, 1400 kg/m3, concrete, 2 or 3 cores.... with perlite-filled cores ........................................... — — with vermiculite-filled cores ................................... — — Stone, lime, or sand ............................................................. 2880 10.4 Quartzitic and sandstone...................................................... 2560 6.2 2240 3.46 1920 1.88 Calcitic, dolomitic, limestone, marble, and granite............. 2880 4.33 2560 3.17 2240 2.31 1920 1.59 1600 1.15 Gypsum partition tile 75 by 300 by 760 mm, solid ........................................ — — 4 cells ................................... — — 100 by 300 by 760 mm, 3 cells ................................... — — Limestone ............................................................................ 2400 0.57 2600 0.93 Concretesi Sand and gravel or stone aggregate concretes..................... 2400 1.4 to 2.9 (concretes with >50% quartz or quartzite sand have 2240 1.3 to 2.6 conductivities in higher end of range) 2080 1.0 to 1.9 Low-mass aggregate or limestone concretes....................... 1920 0.9 to 1.3 expanded shale, clay, or slate; expanded slags; cinders; 1600 0.68 to 0.89 pumice (with density up to 1600 kg/m3); scoria (sanded 1280 0.48 to 0.59 concretes have conductivities in higher end of range) 960 0.30 to 0.36 640 0.18 Gypsum/fiber concrete (87.5% gypsum, 12.5% wood chips) 800 0.24 Cement/lime, mortar, and stucco......................................... 1920 1.40 1600 0.97 1280 0.65 Perlite, vermiculite, and polystyrene beads......................... 800 0.26 to 0.27 640 0.20 to 0.22 480 0.16 320 0.12 Foam concretes.................................................................... 1920 0.75 1600 0.60 1280 0.44 1120 0.36 Foam concretes and cellular concretes ................................ 960 0.30 640 0.20 320 0.12 Aerated concrete (oven-dried) ............................................ 430 to 800 0.20 Polystyrene concrete (oven-dried) ...................................... 255 to 800 0.37 Polymer concrete ................................................................. 1950 1.64 2200 1.03 Polymer cement ................................................................... 1870 0.78 Slag concrete ....................................................................... 960 0.22 1280 0.32
0.34 to 0.24 0.217
— 0.92
Valore (1988) Valore (1988)
0.30 to 0.22 0.65 to 0.41 0.58 0.56 0.47
— — — — —
Van Geem (1985) Van Geem (1985) Van Geem (1985) Van Geem (1985) Van Geem (1985)
0.34 to 0.29 0.74 0.53 0.56 to 0.33 1.20 to 0.77 0.93 to 0.69 0.85 0.79 0.62 0.46 to 0.40 1.6 to 1.1 1.0 — — — — — — — — —
— — — 0.88 — — — — — — — — — — — 0.88 — — — 0.88 —
Van Geem (1985) Van Geem (1985) Van Geem (1985) Van Geem (1985) Van Geem (1985) Shu et al. (1979) Shu et al. (1979) Shu et al. (1979) Shu et al. (1979) Van Geem (1985) Van Geem (1985) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988)
0.222 0.238 0.294 — —
0.79 — — 0.84 0.84
Rowley and Algren (1937) Rowley and Algren (1937) Rowley and Algren (1937) Kumaran (2002) Kumaran (2002)
— — — — — — — — — — — — — — — — — — — — — — — — — — — — — —
— 0.80 to 1.00 — — 0.84 0.84 — — 0.84 — — — — 0.63 to 0.96 — — — — — — — — — 0.84 0.84 — — — — —
Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Rowley and Algren (1937) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Valore (1988) Kumaran (1996) Kumaran (1996) Kumaran (1996) Kumaran (1996) Kumaran (1996) Touloukian et al (1970) Touloukian et al. (1970)
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26.12
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Building and Insulating Materials: Design Valuesa (Continued) Density, kg/m3
Licensed for single user. © 2017 ASHRAE, Inc.
Description
Woods (12% moisture content)l Hardwoods Oak ...................................................................................... Birch .................................................................................... Maple................................................................................... Ash....................................................................................... Softwoods Southern pine....................................................................... Southern yellow pine........................................................... Eastern white pine ............................................................... Douglas fir/larch.................................................................. Southern cypress.................................................................. Hem/fir, spruce/pine/fir ....................................................... Spruce .................................................................................. Western red cedar ................................................................ West coast woods, cedars .................................................... Eastern white cedar.............................................................. California redwood.............................................................. Pine (oven-dried) ................................................................ Spruce (oven-dried) ............................................................
Conductivityb k, Resistance R, Specific Heat, kJ/(kg·K) Referenceo W/(m·K) (m2 ·K)/W
1600 2000
0.43 1.23
— —
— —
— 660 to 750 680 to 725 635 to 700 615 to 670 — 570 to 660 500 400 535 to 580 500 to 515 390 to 500 400 350 350 to 500 360 390 to 450 370 395
— 0.16 to 0.18 0.17 to 0.18 0.16 to 0.17 0.15 to 0.16 — 0.14 to 0.16 0.13 0.10 0.14 to 0.15 0.13 0.11 to 0.13 0.09 0.09 0.10 to 0.13 0.10 0.11 to 0.12 0.092 0.10
— — — — — — — — — — — — — — — — — — —
1.63n — — — — 1.63n — — — — — — — — — — — 1.88 1.88
Touloukian et al. (1970) Touloukian et al. (1970) Wilkes (1979) Cardenas and Bible (1987) Cardenas and Bible (1987) Cardenas and Bible (1987) Cardenas and Bible (1987) Wilkes (1979) Cardenas and Bible (1987) Kumaran (2002) Kumaran (2002) Cardenas and Bible (1987) Cardenas and Bible (1987) Cardenas and Bible (1987) Kumaran (2002) Kumaran (2002) Cardenas and Bible (1987) Kumaran (2002) Cardenas and Bible (1987) Kumaran (1996) Kumaran (1996)
Notes for Table 1 aValues are for mean temperature of 24°C. Representative values for dry materials are intended
kVinyl
as design (not specification) values for materials in normal use. Thermal values of insulating materials may differ from design values depending on in situ properties (e.g., density and moisture content, orientation, etc.) and manufacturing variability. For properties of specific product, use values supplied by manufacturer or unbiased tests. bSymbol also used to represent thermal conductivity. cDoes not include paper backing and facing, if any. Where insulation forms boundary (reflective or otherwise) of airspace, see Tables 2 and 3 for insulating value of airspace with appropriate effective emittance and temperature conditions of space. dConductivity varies with fiber diameter (see Chapter 25). Batt, blanket, and loose-fill mineral fiber insulations are manufactured to achieve specified R-values, the most common of which are listed in the table. Because of differences in manufacturing processes and materials, the product thicknesses, densities, and thermal conductivities vary over considerable ranges for a specified R-value. eValues are for aged products with gas-impermeable facers on the two major surfaces. An aluminum foil facer of 25 m thickness or greater is generally considered impermeable to gases. For change in conductivity with age of expanded polyisocyanurate, see SPI Bulletin U108. fCellular phenolic insulation may no longer be manufactured. gInsulating values of acoustical tile vary, depending on density of board and on type, size, and depth of perforations. hValues for fully grouted block may be approximated using values for concrete with similar unit density. iValues for concrete block and concrete are at moisture contents representative of normal use. jValues for metal or vinyl siding applied over flat surfaces vary widely, depending on ventilation of the airspace beneath the siding; whether airspace is reflective or nonreflective; and on thickness, type, and application of insulating backing-board used. Values are averages for use as design guides, and were obtained from several guarded hot box tests (ASTM Standard C1363) on hollow-backed types and types made using backing of wood fiber, foamed plastic, and glass fiber. Departures of ±50% or more from these values may occur.
lSee
ensure a reliable air barrier assembly: the system must be properly fastened and supported (on both sides) to resist wind loads, and all materials must be durable for the expected service life of the assembly. The air barrier must also be continuous, and should be installed in such a way as to discourage wind washing (i.e., air movement that reduces the thermal resistance of insulation layers in the assembly).
4.5
WATER VAPOR PERMEANCE DATA
Table 5 gives typical water vapor permeance and permeability values for common building materials. These values can be used to calculate water vapor flow through building components and assemblies using equations in Chapter 25.
specific heat = 1.0 kJ/(kg·K) Adams (1971), MacLean (1941), and Wilkes (1979). Conductivity values listed are for heat transfer across the grain. Thermal conductivity of wood varies linearly with density, and density ranges listed are those normally found for wood species given. If density of wood species is not known, use mean conductivity value. For extrapolation to other moisture contents, the following empirical equation developed by Wilkes (1979) may be used: –2
–4
1.874 10 + 5.753 10 M k = 0.1791 + --------------------------------------------------------------------------------1 + 0.01M where is density of moist wood in kg/m3, and M is moisture content in percent. mDimension
referenced is taken at the maximum siding profile thickness. The range of R values and associated thicknesses represent values for products tested to ASTM Standard D7793, which requires applying 6.7 m/s airstream perpendicular to surface of siding during testing.
nFrom
Wilkes (1979), an empirical equation for specific heat of moist wood at 24°C is as follows: 0.299 + 0.01M c p = ---------------------------------------- + c p 1 + 0.01M where cp accounts for heat of sorption and is denoted by c p = M 1.921 10
–3
–5
– 3.168 10 M
where M is moisture content in percent by mass. oBlank
space in reference column indicates historical values from previous volumes of ASHRAE Handbook. Source of information could not be determined.
Water vapor permeability of most building materials is a function of moisture content, which, in turn, is a function of ambient relative humidity. Permeance values at various relative humidities are presented in Table 6 for several building materials. Figure 4 depicts the increase in permeability with increasing relative humidity for oriented strand board (OSB) and plywood samples (Kumaran 2002). Users of the dew-point method may use constant values found in Table 5. However, if condensation in the assembly is predicted, then a more appropriate value should be used. Transient hygrothermal modeling typically uses vapor permeability values that vary with relative humidity. Vapor permeability of homogeneous materials can be calculated from thickness and vapor permeance (as given in Table 6).
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties
26.13
Table 2 Emissivity of Various Surfaces and Effective Emittances of Facing Air Spacesa Effective Emittance eff of Air Space
Licensed for single user. © 2017 ASHRAE, Inc.
Surface Aluminum foil, bright Aluminum foil, with condensate just visible (>0.5 g/m2) Aluminum foil, with condensate clearly visible (>2.0 g/m2) Aluminum sheet Aluminum-coated paper, polished Brass, nonoxidized Copper, black oxidized Copper, polished Iron and steel, polished Iron and steel, oxidized Lead, oxidized Nickel, nonoxidized Silver, polished Steel, galvanized, bright Tin, nonoxidized Aluminum paint Building materials: wood, paper, masonry, nonmetallic paints Regular glass
Average One Surface’s Both Emissivity Emittance ; Surfaces’ Other, 0.9 Emittance 0.05
0.05
0.03
0.30b
0.29
—
0.70b
0.65
—
0.12
0.12
0.06
0.20
0.20
0.11
0.04 0.74 0.04 0.2 0.58 0.27 0.06 0.03 0.25 0.05 0.50
0.038 0.41 0.038 0.16 0.35 0.21 0.056 0.029 0.24 0.047 0.47
0.02 0.59 0.02 0.11 0.41 0.16 0.03 0.015 0.15 0.026 0.35
0.90
0.82
0.82
0.84
0.77
0.72
Fig. 4 Permeability of Wood-Based Sheathing Materials at Various Relative Humidities
aValues
apply in 4 to 40 m range of electromagnetic spectrum. Also, oxidation, corrosion, and accumulation of dust and dirt can dramatically increase surface emittance. Emittance values of 0.05 should only be used where the highly reflective surface can be maintained over the service life of the assembly. Except as noted, data from VDI (1999). bValues based on data in Bassett and Trethowen (1984).
Fig. 5 Sorption/Desorption Isotherms, Cement Board
4.6
MOISTURE STORAGE DATA
Transient analysis of assemblies requires consideration of the materials’ moisture storage capacity. Some materials (hygroscopic) adsorb or reject moisture to achieve equilibrium with adjacent air. Storage capacity of these materials is typically shown by graphs of moisture content versus humidity. The curve showing uptake of moisture (the sorption isotherm) is usually above the curve showing drying (the desorption isotherm) because the material’s uptake and release of moisture are inhibited by surface tension. Table 7 provides data for these curves for several hygroscopic materials, and Kumaran (1996, 2002) and McGowan (2007) provide actual curves, additional data, and conditions under which they were determined. Table 7 expresses moisture content as percentage of dry mass, followed by a subscript value of the relative air humidity at which this moisture content occurs. Note that these values are based on measurement of materials that have reached equilibrium with their surroundings, which in some cases can take many weeks. Most hygrothermal simulation software programs that use these values assume that equilibrium is achieved instantaneously. Maximum values in Table 7 are those that could be realistically measured in laboratory conditions, so not all materials have a listing for maximum moisture content at 100% relative humidity. For those that do, there may be two listings: the moisture content measured when the material’s capillary pores were saturated (shown as 100c), and the value at total saturation (shown as 100t). Note that
the moisture content of any material is 0.0 at a theoretical relative humidity of 0%, so this point is not shown in the table. Figure 5 shows an example of a conventional sorption isotherm graph. Curves show sorption (wetting) and desorption (drying) for data in Table 7 and from Kumaran (2002). (Data from Table 7 were selectively used to provide an accurate representation of the sorption isotherm; not all data from the original source are represented.)
4.7
SOILS DATA
Apparent soil thermal conductivity is difficult to estimate and may change in the same soil at different times because of changed moisture conditions and freezing temperatures. Figure 6 shows typical apparent soil thermal conductivity as a function of moisture content for different general types of soil. The figure is based on data presented in Salomone and Marlowe (1989) using envelopes of thermal behavior coupled with field moisture content ranges for different soil types. In Figure 6, “well graded” applies to granular soils with good representation of all particle sizes from largest to smallest. “Poorly graded” refers to granular soils with either uniform gradation, in which most particles are about the same size, or skip (or gap) gradation, in which particles of one or more intermediate sizes are not present. Although thermal conductivity varies greatly over the complete range of possible moisture contents, this range can be narrowed if it
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26.14
2017 ASHRAE Handbook—Fundamentals (SI) Table 3 Thermal Resistances of Plane Air Spaces,a,b,c (m2 ·K)/W Air Space Effective Emittance eff d,e
Position of Air Space
Horiz.
45° Slope
Licensed for single user. © 2017 ASHRAE, Inc.
Vertical
45° Slope
Horiz.
Direction of Heat Flow
Temp. Mean Temp.d, °C Diff.,d K
0.03
13 mm Air Spacec 0.05 0.2 0.5
0.82
0.03
20 mm Air Spacec 0.05 0.2 0.5
0.82
Up
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.37 0.29 0.37 0.30 0.37 0.30 0.36
0.36 0.28 0.36 0.30 0.36 0.29 0.35
0.27 0.23 0.28 0.26 0.30 0.26 0.31
0.17 0.17 0.20 0.20 0.22 0.22 0.25
0.13 0.13 0.15 0.16 0.18 0.18 0.20
0.41 0.30 0.40 0.32 0.39 0.31 0.38
0.39 0.29 0.39 0.32 0.38 0.31 0.37
0.28 0.24 0.30 0.27 0.31 0.27 0.32
0.18 0.17 0.20 0.20 0.23 0.22 0.26
0.13 0.14 0.15 0.16 0.18 0.19 0.21
Up
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.43 0.36 0.45 0.39 0.46 0.37 0.46
0.41 0.35 0.43 0.38 0.45 0.36 0.45
0.29 0.27 0.32 0.31 0.36 0.31 0.38
0.19 0.19 0.21 0.23 0.25 0.25 0.29
0.13 0.15 0.16 0.18 0.19 0.21 0.23
0.52 0.35 0.51 0.37 0.48 0.36 0.45
0.49 0.34 0.48 0.36 0.46 0.35 0.43
0.33 0.27 0.35 0.30 0.37 0.31 0.37
0.20 0.19 0.23 0.23 0.26 0.25 0.29
0.14 0.14 0.17 0.18 0.20 0.20 0.23
Horiz.
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.43 0.45 0.47 0.50 0.52 0.51 0.56
0.41 0.43 0.45 0.48 0.50 0.50 0.55
0.29 0.32 0.33 0.38 0.39 0.41 0.45
0.19 0.22 0.22 0.26 0.27 0.31 0.33
0.14 0.16 0.16 0.20 0.20 0.24 0.26
0.62 0.51 0.65 0.55 0.66 0.51 0.65
0.57 0.49 0.61 0.53 0.63 0.50 0.63
0.37 0.35 0.41 0.41 0.46 0.42 0.51
0.21 0.23 0.25 0.28 0.30 0.31 0.36
0.15 0.17 0.18 0.21 0.22 0.24 0.27
Down
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.44 0.46 0.47 0.51 0.52 0.56 0.57
0.41 0.44 0.45 0.49 0.50 0.54 0.56
0.29 0.33 0.33 0.39 0.39 0.44 0.45
0.19 0.22 0.22 0.27 0.27 0.33 0.33
0.14 0.16 0.16 0.20 0.20 0.25 0.26
0.62 0.60 0.67 0.66 0.73 0.67 0.77
0.58 0.57 0.63 0.63 0.69 0.64 0.74
0.37 0.39 0.42 0.46 0.49 0.51 0.57
0.21 0.24 0.26 0.30 0.32 0.36 0.39
0.15 0.17 0.18 0.22 0.23 0.28 0.29
Down
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.44 0.47 0.47 0.52 0.52 0.57 0.58
0.41 0.45 0.45 0.50 0.50 0.55 0.56
0.29 0.33 0.33 0.39 0.39 0.45 0.46
0.19 0.22 0.22 0.27 0.27 0.33 0.33
0.14 0.16 0.16 0.20 0.20 0.26 0.26
0.62 0.66 0.68 0.74 0.75 0.81 0.83
0.58 0.62 0.63 0.70 0.71 0.78 0.79
0.37 0.42 0.42 0.50 0.51 0.59 0.60
0.21 0.25 0.26 0.32 0.32 0.40 0.40
0.15 0.18 0.18 0.23 0.23 0.30 0.30
40 mm Air Spacec
Air Space
Horiz.
45° Slope
Vertical
90 mm Air Spacec
Up
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.45 0.33 0.44 0.35 0.43 0.34 0.42
0.42 0.32 0.42 0.34 0.41 0.34 0.41
0.30 0.26 0.32 0.29 0.33 0.30 0.35
0.19 0.18 0.21 0.22 0.24 0.24 0.27
0.14 0.14 0.16 0.17 0.19 0.20 0.22
0.50 0.27 0.49 0.40 0.48 0.39 0.47
0.47 0.35 0.47 0.38 0.46 0.38 0.45
0.32 0.28 0.34 0.32 0.36 0.33 0.38
0.20 0.19 0.23 0.23 0.26 0.26 0.29
0.14 0.15 0.16 0.18 0.20 0.21 0.23
Up
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.51 0.38 0.51 0.40 0.49 0.39 0.48
0.48 0.36 0.48 0.39 0.47 0.38 0.46
0.33 0.28 0.35 0.32 0.37 0.33 0.39
0.20 0.20 0.23 0.24 0.26 0.26 0.30
0.14 0.15 0.17 0.18 0.20 0.21 0.24
0.56 0.40 0.55 0.43 0.52 0.41 0.51
0.52 0.38 0.52 0.41 0.51 0.40 0.49
0.35 0.29 0.37 0.33 0.39 0.35 0.41
0.21 0.20 0.24 0.24 0.27 0.27 0.31
0.14 0.15 0.17 0.19 0.20 0.22 0.24
Horiz.
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.70 0.45 0.67 0.49 0.62 0.46 0.58
0.64 0.43 0.62 0.47 0.59 0.45 0.56
0.40 0.32 0.42 0.37 0.44 0.38 0.46
0.22 0.22 0.26 0.26 0.29 0.29 0.34
0.15 0.16 0.18 0.20 0.22 0.23 0.26
0.65 0.47 0.64 0.51 0.61 0.50 0.60
0.60 0.45 0.60 0.49 0.59 0.48 0.58
0.38 0.33 0.41 0.38 0.44 0.40 0.47
0.22 0.22 0.25 0.27 0.29 0.30 0.34
0.15 0.16 0.18 0.20 0.22 0.24 0.26
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties
26.15
Table 3 Thermal Resistances of Plane Air Spaces,a,b,c (m2 ·K)/W (Continued) Air Space Effective Emittance eff d,e Position of Air Space
45° Slope
Horiz.
0.03
40 mm Air Spacec 0.05 0.2 0.5
0.82
0.03
90 mm Air Spacec 0.05 0.2 0.5
Down
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
0.89 0.63 0.90 0.68 0.87 0.64 0.82
0.80 0.59 0.82 0.64 0.81 0.62 0.79
0.45 0.41 0.50 0.47 0.56 0.49 0.60
0.24 0.25 0.28 0.31 0.34 0.35 0.40
0.16 0.18 0.19 0.22 0.24 0.27 0.30
0.85 0.62 0.83 0.67 0.81 0.66 0.79
0.76 0.58 0.77 0.64 0.76 0.64 0.76
0.44 0.40 0.48 0.47 0.53 0.51 0.58
0.24 0.25 0.28 0.31 0.33 0.36 0.40
0.16 0.18 0.19 0.22 0.24 0.28 0.30
Down
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
1.07 1.10 1.16 1.24 1.29 1.36 1.42
0.94 0.99 1.04 1.13 1.17 1.27 1.32
0.49 0.56 0.58 0.69 0.70 0.84 0.86
0.25 0.30 0.30 0.39 0.39 0.50 0.51
0.17 0.20 0.20 0.26 0.27 0.35 0.35
1.77 1.69 1.96 1.92 2.11 2.05 2.28
1.44 1.44 1.63 1.68 1.82 1.85 2.03
0.60 0.68 0.72 0.86 0.89 1.06 1.12
0.28 0.33 0.34 0.43 0.44 0.57 0.59
0.18 0.21 0.22 0.29 0.29 0.38 0.39
Direction of Heat Flow
Temp. Mean Temp.d, °C Diff.,d K
143 mm Air Spacec
Licensed for single user. © 2017 ASHRAE, Inc.
Air Space
Horiz.
45° Slope
Vertical
45° Slope
Horiz.
Up
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 5.6
Up
32.2 10.0 10.0 17.8 17.8 45.6 45.6
Horiz.
32.2 10.0 10.0 17.8 17.8 45.6 45.6
Down
32.2 10.0 10.0 17.8 17.8 45.6 45.6
Down
32.2 10.0 10.0 17.8 17.8 45.6 45.6
5.6 16.7 5.6 11.1 5.6 11.1 11.1 5.6 16.7 5.6 11.1 5.6 11.1 11.1 5.6 16.7 5.6 11.1 5.6 11.1 11.1 5.6 16.7 5.6 11.1 5.6 11.1 11.1
aSee Chapter 25. Thermal resistance values were determined from R
0.82
0.53 0.39 0.52 0.42 0.51 0.41 0.49 0.57 0.39 0.56 0.41 0.53 0.38 0.49 0.66 0.50 0.66 0.54 0.64 0.53 0.63 0.865 0.68 0.87 0.75 0.87 0.75 0.87 2.06 1.87 2.24 2.13 2.43 2.19 2.57
0.50 0.38 0.50 0.41 0.49 0.40 0.48 0.54 0.37 053 0.40 0.51 0.37 0.48 0.61 0.47 0.61 0.52 0.61 0.52 0.61 0.78 0.64 0.80 0.71 0.82 0.73 0.83 1.63 1.57 1.82 1.84 2.05 1.96 2.26
0.33 0.29 0.36 0.33 0.38 0.34 0.40 0.35 0.29 0.37 0.33 0.39 0.32 0.40 0.38 0.35 0.42 0.40 0.46 0.43 0.49 0.44 0.43 0.49 0.50 0.56 0.56 0.62 0.63 0.71 0.75 0.90 0.95 1.10 1.18
= 1/C, where C = hc + eff hr , hc is conduction/convection coefficient, eff hr is radiation coefficient 0.227eff [(tm + 273)/100]3, and tm is mean temperature of air space. Values for hc were determined from data developed by Robinson et al. (1954). Equations (5) to (7) in Yarbrough (1983) show data in this table in analytic form. For extrapolation from this table to air spaces less than 12.5 mm (e.g., insulating window glass), assume hc = 21.8(1 + 0.00274tm)/l, where l is air space thickness in mm, and hc is heat transfer in W/(m2 ·K) through air space only. bValues based on data presented by Robinson et al. (1954). (Also see Chapter 4, Tables 5 and 6, and Chapter 33.) Values apply for ideal conditions (i.e., air spaces of uniform thickness bounded by plane, smooth, parallel surfaces with no air leakage to or from the space). This table should not be used for hollow siding or profiled cladding: see Table 1. For greater accuracy, use overall Ufactors determined through guarded hot box (ASTM Standard C1363) testing. Thermal resistance values for multiple air spaces must be based on careful estimates of mean temperature differences for each air space.
0.20 0.20 0.23 0.24 0.27 0.27 0.30 0.21 0.20 0.24 0.24 0.27 0.26 0.30 0.22 0.23 0.25 0.28 0.30 0.32 0.35 0.24 0.26 0.28 0.32 0.34 0.39 0.41 0.28 0.34 0.35 0.44 0.46 0.58 0.61 cA
0.14 0.15 0.17 0.19 0.20 0.22 0.24 0.15 0.15 0.17 0.18 0.20 0.21 0.24 0.15 0.16 0.18 0.21 0.22 0.25 0.27 0.16 0.18 0.19 0.23 0.24 0.29 0.31 0.18 0.22 0.22 0.29 0.29 0.39 0.40
single resistance value cannot account for multiple air spaces; each air space requires a separate resistance calculation that applies only for established boundary conditions. Resistances of horizontal spaces with heat flow downward are substantially independent of temperature difference. dInterpolation is permissible for other values of mean temperature, temperature difference, and effective emittance eff. Interpolation and moderate extrapolation for air spaces greater than 90 mm are also permissible. eEffective emittance of air space is given by 1/ = 1/ + 1/ 1, eff eff 1 2 where 1 and 2 are emittances of surfaces of air space (see Table 2). Also, oxidation, corrosion, and accumulation of dust and dirt can dramatically increase surface emittance. Emittance values of 0.05 should only be used where the highly reflective surface can be maintained over the service life of the assembly.
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26.16
2017 ASHRAE Handbook—Fundamentals (SI) Table 4 Air Permeability of Different Materials Mean Air Permeability, kg/(Pa · s· m)
Material kg/m3
Licensed for single user. © 2017 ASHRAE, Inc.
Cement board, 12.5 mm, 1140 Fiber cement board, 6.3 mm, 1380 kg/m3 Gypsum wall board, 12.5 mm, 625 kg/m3 with one coat primer with one coat primer/two coats latex paint Hardboard siding, 9.5 mm, 740 kg/m3 Oriented strand board (OSB), 1140 kg/m3, 9.5 mm 11 mm 12.5 mm Douglas fir plywood, 12.5 mm, 455 kg/m3 16 mm, 545 kg/m3 Canadian softwood plywood, 19 mm, 450 kg/m3 Wood fiber board, 9.5 mm, 320 kg/m3 Masonry Materials Aerated concrete, 460 kg/m3 Cement mortar, 1600 kg/m3 Clay brick, 100 by 100 by 200 mm, 1990 kg/m3 Limestone, 2500 kg/m3 Portland stucco mix, 1990 kg/m3 Eastern white cedar, (transverse) 19 mm, 465 kg/m3 Eastern white pine, (transverse) 19 mm, 465 kg/m3 Southern yellow pine, (transverse) 19 mm, 500 kg/m3 Spruce, (transverse) 19 mm, 400 kg/m3 Western red cedar, (transverse) 19 mm, 350 kg/m3 Cellulose insulation, dry blown, 32 kg/m3 Glass fiber batt, 16 kg/m3 Polystyrene expanded, 16 kg/m3 sprayed foam, 38 kg/m3 6.5 to 19 kg/m3 Polyisocyanurate insulation, 26.5 kg/m3 Bituminous paper (#15 felt), (transverse) 0.7 mm, 865 kg/m3 Asphalt-impregnated paper #10, (transverse) 0.13 mm, 95 kg/m3 #30, (transverse) 0.15 mm, 130 kg/m3 #60, (transverse) 0.23 mm, 260 kg/m3 Spun bonded polyolefin (SBPO) (transverse) 0.1 mm, 14 kg/m3 with crinkled surface, (transverse) 0.075-0.1 mm, 15 kg/m3 Wallpaper, vinyl, (transverse) 0.13 mm, 94 kg/m3 Exterior insulated finish system (EIFS), 1 mm, 1140 kg/m3
10–8
3 3 10–12 4.2 10–9 2.2 10–8 2.5 10–9 4.5 10–9 1 10–9 2 10–9 1 10–9 4 10–11 1 10–9 2 10–11 2.5 10–7
Table 4 Air Permeability of Different Materials Mean Air Permeability, kg/(Pa · s · m)
Material Source: Kumaran (2002).
5 10–9 1.5 10–9 2 to 5 10–10 negligible 1 10–11 negligible 1 10–12 3 10–11
Fig. 6 Trends of Apparent Thermal Conductivity of Moist Soils
5 10–11 < 1 10–12 2.9 10–4 2.5 10–4 1.1 10–8 1 10–11 4.2 10–9 negligible 2.5 10–6
is assumed that the moisture contents of most field soils lie between the wilting point of the soil (i.e., the moisture content of a soil below which a plant cannot alleviate its wilting symptoms) and the field capacity of the soil (i.e., the moisture content of a soil that has been thoroughly wetted and then drained until the drainage rate has become negligibly small). After a prolonged dry spell, moisture is near the wilting point, and after a rainy period, soil has moisture content near its field capacity. Moisture contents at these limits have been studied by many agricultural researchers, and data for different types of soil are given by Kersten (1949) and Salomone and Marlowe (1989). Shaded areas in Figure 6 approximate (1) the full range of moisture contents for different soil types and (2) a range between average values of each limit. Table 8 summarizes design values for thermal conductivities of the basic soil classes. Table 9 gives ranges of thermal conductivity for some basic classes of rock. The value chosen depends on whether heat transfer is calculated for minimum heat loss through the soil, as in a ground heat exchange system, or a maximum value, as in peak winter heat loss calculations for a basement. Hence, high and low values are given for each soil class. effect of the latent heat of fusion of water. Energy released during this phase change significantly retards the progress of the frost front in moist soils. For further information, see Chapter 17, which includes a method for estimating heat loss through foundations.
1.1 10–6 6.6 10–6 7.1 10–6 4.6 10–7 3 10–7 5 10–9 0
As heat flows through soil, moisture tends to move away from the heat source. This moisture migration provides initial mass transport of heat, but it also dries the soil adjacent to the heat source, thus lowering the apparent thermal conductivity in that zone of soil. Typically, when other factors are held constant, k increases with moisture content and with dry density of a soil, but decreases with increasing organic content of a soil and for uniform gradations and rounded soil grains (because grain-to-grain contacts are reduced). The k of a frozen soil may be higher or lower than that of the same unfrozen soil (because the conductivity of ice is higher than that of water but lower than that of typical soil grains). Differences in k below moisture contents of 7 to 8% are quite small. At approximately 15% moisture content, k may vary up to 30% from unfrozen values. When calculating annual energy use, choose values that represent typical mean site conditions. In climates where ground freezing is significant, accurate heat transfer simulations should include the
4.8
SURFACE FILM COEFFICIENTS/ RESISTANCES
As explained in Chapter 25, the overall thermal resistance of an assembly comprises its surface-to-surface thermal resistance Rs and the surface film resistances between the assembly’s surfaces and the interior and exterior environment (Ri and Ro). Table 10 gives typical values for the surface film coefficients hi and ho and their reciprocals, the surface resistances Ri and Ro. As shown, the indoor values depend on position of the surface, direction of heat transfer, and the
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties
26.17
Table 5 Typical Water Vapor Permeance and Permeability for Common Building Materialsa Material
Mass, kg/m2
Plastic and Metal Foils and Filmsb Aluminum foil
0.025 0.009 0.051 0.1 0.15 0.2 0.25 0.051 0.1 0.025 0.09 0.19 0.25 3.2
Polyethylene
Polyvinylchloride, unplasticized Polyvinylchloride, plasticized Polyester
Licensed for single user. © 2017 ASHRAE, Inc.
Cellulose acetate Liquid-Applied Coating Materials Commercial latex paints (dry film thickness) Vapor retarder paint Primer-sealer Vinyl acetate/acrylic primer Vinyl/acrylic primer Semigloss vinyl/acrylic enamel Exterior acrylic house and trim Paint, 2 coats Asphalt paint on plywood Aluminum varnish on wood Enamels on smooth plaster Primers and sealers on interior insulation board Various primers plus 1 coat flat oil paint on plaster Flat paint on interior insulation board Water emulsion on interior insulation board Paint, 3 coats Exterior paint, white lead and oil on wood siding Exterior paint, white lead/zinc oxide and oil on wood Styrene/butadiene latex coating Polyvinyl acetate latex coating Chlorosulfonated polyethylene mastic Asphalt cutback mastic 1.6 mm, dry 4.8 mm, dry Hot-melt asphalt Building Paper, Felts, Roofing Papersc Duplex sheet, asphalt laminated, aluminum foil one side Saturated and coated roll roofing Kraft paper and asphalt laminated, reinforced Blanket thermal insulation back-up paper, asphalt coated Asphalt, saturated and coated vapor retarder paper Asphalt, saturated, but not coated, sheathing paper asphalt felt, 0.73 kg/m2 tar felt, 0.73 kg/m2 Single kraft, double Polyamide film, 2 mil
Thickness, mm
Permeance, ng/(Pa · s · m2) Dry-Cup
Wet-Cup
Other Method
Permeability, ng/(Pa · s · m)
1.7
4.7 10–4 4.7 10–4 4.7 10–4 4.7 10–4 4.7 10–4
0.0 2.9 9.1 4.6 3.4b 2.3b 39b 46 to 80 42 13 4.6 263 18
0.07 0.03 0.05 0.04 0.06 0.04
26 360 424 491 378 313 23 17 to 29 29 to 86 51 to 120 91 to 172 229 1716 to 4863
0.6 1.2 1.1 2.2
17 to 57 51 629 315 97 3.4
0.6 1.1
8.0 0.0 29 5.7
0.42 3.18 0.33 0.30 0.42 0.21 0.68 0.68 0.16
0.1 2.9 17 23 11 to 17 190 57 230 1170 62.9
Source: Lotz (1964). aThis table allows comparisons of materials, but when selecting vapor retarder materials, exact values for permeance or permeability should be obtained from manufacturer or from laboratory tests. Values shown indicate variations among mean values for materials that are similar but of different density, orientation, lot, or source. Values should not be used as design or specification data. Values from dry- and wet-cup methods were usually obtained from investigations using ASTM Standards C355 and E96; other values were obtained by two-temperature, special cell, and air velocity methods.
10 14 103 34 to 240 34 1160 320 1040 2400 1174 bUsually
installed as vapor retarders, although sometimes used as exterior finish and elsewhere near the cold side, where special considerations are then required for warmside barrier effectiveness. cLow-permeance sheets used as vapor retarders. High permeance used elsewhere in construction.
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26.18
2017 ASHRAE Handbook—Fundamentals (SI) Table 6 Water Vapor Permeance at Various Relative Humidities and Capillary Water Absorption Coefficient
Licensed for single user. © 2017 ASHRAE, Inc.
Material Building Board and Siding Asbestos cement board with oil-base finishes Cement board, 1130 kg/m3 Fiber cement board, 1380 kg/m3 Gypsum board, asphalt impregnated Gypsum wall board, 625 kg/m3 with one coat primer with one coat primer/two coats latex paint Hardboard siding, 740 kg/m3 Oriented strand board (OSB), 660 kg/m3 650 kg/m3 650 kg/m3 Particleboard, 762 kg/m3 Plywood Douglas fir, 470 kg/m3 Douglas fir, 550 kg/m3 Canadian softwood, 445 kg/m3 Exterior-grade, 580 kg/m3 Exterior-grade, 510 kg/m3 Wood fiber board, 320 kg/m3 300 kg/m3
Thickness, mm
Permeance at Various Relative Humidities, ng/ (Pa·s·m2) 10%
3
30%
50%
70%
Water Absorption Coefficient, 90% References/Comments kg/(m2·s1/2)
740 200
980 590
1290 1850
0.013 0.025
2340 680 110
230-460 17-29 600 73 2300 2720 1490 210
3190 2200 400
3760 2890 800
4470 3590 1650
0.0019
10.8 9.9
360 0.65
400 18
430 49
470 140
520 390
0.00072 0.0016
Kumaran (2002) Kumaran (2002)
10.8 12.3 19 12
2.4 3.6
110 73 220 120
210 140 280 270
380 220 490 540
0.0022 0.0016
16
56 28 230 49
Kumaran (2002) Kumaran (2002) Burch et al. (1992) Kumaran (2002)
15 17.8 12.1
10 3.3 21
27 32 22
73 130 30
190 340 96
530 750 620
0.0031 0.0037
13 11.8 25.1
1050 2710
67 1150 2780
78 1270 2900
180 1390 3070
880 1530 3330
0.00094
780 1270 360 50
1130 1550 380 56 140
1640 1880 410 100
2460 2320 440 260
0.036 0.02 0.17 0.018
460 10 1300 2550
750 11 3300 2550
12.5 8 12.5 12.5
600 26
0.0042
Dry cup* Dry cup* Kumaran (2002) Kumaran (2002) Dry cup* Kumaran (2002) Kumaran (2002) Kumaran (2002)
Kumaran (2002) Kumaran (2002) Burch and Desjarlais (1995) Burch et al. (1992) Kumaran (2002) Burch and Desjarlais (1995)
Masonry Materials Aerated concrete, 460 kg/m3 Cement mortar, 1600 kg/m3 Clay brick, 1980 kg/m3 Concrete, 2200 kg/m3 Concrete block (cored, limestone aggregate) Lightweight concrete, 1330 kg/m3 Limestone, 2500 kg/m3 Perlite board, 160 kg/m3 173 kg/m3
20.3 13 12.4 25 200
Plaster, on metal lath on wood lath on plain gypsum lath (with studs) Polystyrene concrete, 260800 kg/m3 Portland stucco mix, 1985 kg/m3 Tile masonry, glazed
19
Woods Cedar Eastern white cedar, 360 kg/m3 (transverse) Western red cedar, 350 kg/m3 (transverse) Pine 340 kg/m3 (longitudinal) Eastern white pine, 460 kg/m3 (transverse) Southern yellow pine, 500 kg/m3 (transverse) Sugar pine, 365 kg/m3 (transverse) Spruce 450 kg/m3 (longitudinal)
25 25 25 25.1
550 1050 330
2550
490 10 1100 2550
10 2550
0.00033
860 630 1150
Kumaran (2002) Kumaran (2002) Kumaran (2002) Kumaran (1996) * Kumaran (1996) Kumaran (2002) Kumaran (1996) Burch and Desjarlais (1995) * * *
25
650
680
720
800
940
14 100
58
82
120 6.9
160
230
0.012
Kumaran (2002) *
19
0.66
4.1
26
160
1100
0.0016
Kumaran (2002)
18
5.9
13
27
59
130
0.0010
Kumaran (2002)
1200
1600
3000
4800
0.0163
Kumaran (1996)
25
Kumaran (1996)
19
2.5
9.4
35
140
540
0.0066
Kumaran (2002)
19.5
6.2
21
70
240
870
0.0014
Kumaran (2002)
13 13.2
26 2300
31 5600
54 6400
130 6900
480 7000
0.0096
Burch et al. (1992) Kumaran (1996)
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties Table 6
Water Vapor Permeance at Various Relative Humidities and Capillary Water Absorption Coefficient (Continued)
Material kg/m3 (transverse)
410 400 kg/m3 (transverse) Insulation Air (still) Cellular glass Cellulose, dry blown, 30 kg/m3 Corkboard Glass fiber batt, 11.5 kg/m3 Glass-fiber insulation board, 120 kg/ m3 facer, 880 kg/m3
Licensed for single user. © 2017 ASHRAE, Inc.
26.19
Mineral fiber insulation, 148 to 172 kg/m3 Mineral wool (unprotected) Phenolic foam (covering removed) Polyisocyanurate insulation, 26.5 kg/m3 32.5 kg/m3 Polyisocyanurate glass-mat facer, 430 kg/m3 Polystyrene expanded, 14.8 kg/m3 extruded, 28.6 kg/m3 Polyurethane expanded board stock sprayed foam, 39.0 kg/m3 6.5 to 8.5 kg/m3 Structural insulating board, sheathing quality interior, uncoated Unicellular synthetic flexible rubber foam
10%
30%
50%
70%
Water Absorption Coefficient, 90% References/Comments kg/(m2·s1/2)
11.5 19.5
19
60 55
120 160
500 480
1700 1510
25
Other Construction Materials Built-up roofing (hot-mopped) Exterior insulated finish system (EIFS), 1140 kg/m3 Glass fiber reinforced sheet, acrylic polyester
Kumaran (1996) Kumaran (2002)
1960 8300
2600 550 1960 8300
1960 8300
0.73
2.2
6.6
20
5080
5080
5080
5080
230
260
* * Kumaran (2002)
1740 1960 8300
2170 120-150 1960 8300
1.6
0.24
25
5080
7000 0.0 2420
0.0020
* * Kumaran (2002) * Kumaran (2002) Burch and Desjarlais (1995) Burch and Desjarlais (1995) Kumaran (1996)
64.5 25 88 23.6
2760
0.10
25 25 25
160
180
6660 1500 210
24.6
120
130
140
160
190
0.8
600
860
1260
1860
2760
24.3
120
140
160
190
230
25.4 25
48
48 23 to 92
48
48
48
Kumaran (2002) Dry cup*
25 25 25
94 3500
100 3500
110 3500 11002900 29005200
120 3500
130 3500
Kumaran (2002) Kumaran (2002) *
13 25
Foil, Felt, Paper (transverse) 0.20 Asphalt-impregnated paper, 10 min rating, 170 g/m2 0.22 30 min rating, 200 g/m2 60 min rating, 280 g/m2 0.34 0.72 Bituminous paper (#15 felt), 515 g/m2 Polyamide film 0.050 0.14 to 0.15 Spun bonded polyolefin (SBPO), 65 g/m2 0.10 to 0.11 with crinkled surface, 67 g/m2 Wallpaper paper, 151-168 g/m2 textile, 291-333 g/m2 vinyl, 170 g/m2
Permeance at Various Relative Humidities, ng/ (Pa·s·m2)
Thickness, mm
Burch and Desjarlais (1995) Burch and Desjarlais (1995) Kumaran (2002)
*
1.1-8.6
Dry cup*
240
430
780
1480
3060
0.00099
Kumaran (2002)
440 1510 290
1280 2440 290 240 4370
2310 3180 400 610 4370
4670 4240 1170 2000 4370
0.00093 0.0011 0.00051
4370
740 1910 290 53 4370
0.00031
Kumaran (2002) Kumaran (2002) Kumaran (2002) Gatland II (2005) Kumaran (2002)
3.17
3.17
3.17
3.17
3.17
0.00024
Kumaran (2002)
0.28 0.425-0.70 0.205
85
5000-7000 620-1100 140
210
14000-22000 9200-29000 320
460
0.00025
Kumaran (1996) Kumaran (1996) Kumaran (2002)
**
95
95
0.0 95
95
95
0.00053
* Kumaran (2002)
1.4 1.2
7 3
Dry cup* Dry cup*
*Historical data, no reference available **EIFS vapor permeance was tested with polymer cement base coat and latex acrylic finish coat of 4.4 mm thickness applied to expanded polystyrene of 34 mm thickness.
surface’s long-wave emissivity. Outdoors, the values depends on air speed and the surface’s long-wave emissivity. Table 10 reflects standard situations, with an assumed (approximate) interior surface tem-
perature representative of wall or roof assemblies. For situations that deviate substantially from standard conditions, including interior surface temperatures for fenestration systems, use ASHRAE (1998)
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
26.20
2017 ASHRAE Handbook—Fundamentals (SI) Table 7 Sorption/Desorption Isotherms of Building Materials at Various Relative Humidities Sorption, % Moisture Content at % Relative Humidity
Licensed for single user. © 2017 ASHRAE, Inc.
Material Building Board and Siding Cement board, 13 mm, 1130 kg/m3 Fiber cement board, 8 mm, 1380 kg/m3 Gypsum wall board, 13 mm, 625 kg/m3 Hardboard siding, 11 mm, 740 kg/m3 Oriented strand board (OSB), 9.5 mm, 660 kg/m3 11.1 mm, 650 kg/m3 12.7 mm, 650 kg/m3 Particle board, 19 mm, 760 kg/m3 Plywood, 13 mm 16 mm 19 mm Plywood (exterior-grade), 12 mm, 580 kg/m3 Wood fiber board, 11 mm, 320 kg/m3 25 mm, 300 kg/m3
143
1.970
450.6
5.870.4 16.889.9 34.7100t
Insulation Cellulose, dry-blown, 30 kg/m3 Glass fiber batt, 11.5 kg/m3 Glass-fiber board, 24 mm, 120 kg/m3 Glass-fiber board facer, 1.6 mm, 880 kg/m3 Mineral fiber, 40 kg/m3 Polystyrene, expanded, 14.8 kg/m3 extruded, 28.6 kg/m3 Polyurethane, sprayed foam, 39 kg/m3 6.5 to 8.5 kg/m3 Polyisocyanurate, 26.5 kg/m3 Polyisocyanurate glass facer, 1 mm, 430 kg/m3
6.193
0.450.5 0.6570.5 1.890.8 4.294
42.7100t
1.643
3.270
4.681
6.293
1899.27 2899.93
References Kumaran (2002)
6.650.5 12.370.5 19.690.6 31.395.32 32.599.49 33.999.93 Kumaran (2002) 68.9100c 113100t 0.9950.4 1.3271.5 1.6984.8 1.8288.3
Kumaran (2002)
4.750.3 6.969.6 13.191.3 90100t
4.450.3 7.669.2 13.491.3 3891.3
4.648.9 7.669.1 14.788.6 126100c
6.949.9 9.169.4 16.290.3 17.392.3 39.399.3 60.699.8 Kumaran (2002)
Kumaran (2002)
5.448.9 8.269.1 14.788.6 160100t 7.949.9 9.969.4 17.490.3 39.199.3 62.799.8 Kumaran (2002) 4.648.9 7.869.1 14.888.6 124100t 7.949.9 1069.4 17.690.3 2092.3 4299.3 59.5 Kumaran (2002) 1.211.3 6.357.6 9.778.6 11.384.1 15.993.6 21.597.3 1.711.3 8.857.6 1478.6 16.684.1 1993.6 23.397.6 Kumaran (1996) 748.9 6.848.9 6.748.9 1.8311.3
9.269.1 9.669.1 10.169.1 6.958
15.888.6 16.888.6 17.688.6 9.578.7
170100t 140100t 190100t 12.184.5 17.993.8 22.1
4.650.6 7.470.5 15.891.1 304 0.6311.3 5.758
Masonry Materials Aerated concrete, 460 kg/m3 1.150.6 600 kg/m3 1.817.8 Cement mortar, 1600 kg/m3 0.4249.9 Clay brick, 100 × 100 × 200 mm, 0.0850 1980 kg/m3 Concrete, 2200 kg/m3 0.8825.2 Lightweight concrete, 2.924.4 1100 kg/m3 Limestone, 2500 kg/m3 050 Perlite board 13033 Portland stucco mix, 1985 kg/m3 350 Woods Eastern white cedar, 25 mm, 360 kg/m3 Eastern white pine, 25 mm, 460 kg/m3 Southern yellow pine, 25 mm, 500 kg/m3 Spruce (transverse) Western red cedar, 25 mm, 350 kg/m3
3.481
Desorption, % Moisture Content at % Relative Humidity
2.171.5 3.275.8 2.370.1 0.1269.1
8.449.9 8.649.9 8.949.9 2.0911.3
10.869.4 11.369.4 11.369.4 9.358
18.290.3 19.890.3 19.390.3 13.778.7
3.950.6 7.471.1 1590.6
1992.3 19.392.3 20.792.3 15.284.5
7099.3 4799.3 6699.3 19.893.8
101 79 9999.8 23.4
23099.71 23099.85 23099.93 Kumaran (2002)
9.278.7 11.384.5 16.493.8 24.697.4 1.2611.3 7.658
1278.7
14.684.5 20.693.8 28.197.4
588.1 4.690.3 5.389.9 0.191.2
6.388.1 455.2 6.189.9 4.598.9
3497.81 6.675.6 1798.9 699.63
83100c 172 6.492.4 9.695.9 17.598.4 26100t 9.9100t
1.150.6 2.317.8 3.449.9 050
2.271.5 2.833 4.470.2 091.2
7299.85 15.491.6 2299.63 8.299.71
9299.99 36.598 2599.93 9.199.93
1.1544.9 1.7465 2.6280 3.3589.8 4.4598.2 0.9420 2.1945.4 2.9865.6 3.8584.8 4.5794.8 3.445.2 465.2 4.685 6.698 3.119.6 4.440 5.259.8 679.6 7.194.7 070 0.188.5 1.8100t 16052 26075 38086 80097 3.770.3 5.889.9 12100t
117099.8
070.5
Kumaran (2002) Kumaran (1996) Kumaran (2002) Kumaran (2002) Kumaran (1996) Kumaran (1996)
4.250
0.188.6 0.2195.3 0.598.9 0.699.27 1.399.93 Kumaran (2002) Kumaran (1996) 5.270.3 790.3 10.395.29 11.698.9 11.799.93 Kumaran (2002)
3.449.8 7.670
12.888.5 228100t
1.750
7.470.5 11.988.7 8598.9
3.249.8 7.670
1288.5
192100t
3.250
970.5
12.488.7 8499.78
3.649.8 8.170
15.288.5 158100t
4.350
1070.5
15.688.7 5799.78
4.149.8 9.270 3.449.8 670
16.788.5 228100t 9.688.5 228100t
4.950 150
11.370.5 17.788.7 14895.96 18799.78 970.5 13.388.7 11399.78
550.2
1272.8
6.150.5 9.671.5 2488.1
Kumaran (2002) Kumaran (2002) Kumaran (2002) Burch et al.
11899.63 17699.92
2688
Kumaran (2002)
0.2150.6 0.3471.5 0.7588.1 0.2450.4 0.3571.4 0.6788.2 Kumaran (2002) 0.1611.3 0.75 0.8278.7 0.9684.5 1.393.8 2.0397.4 0.4311.3 0.8632.8 1.1158 1.2684.5 1.7493.8 2.1697.4 Burch et al. 0.0911.3 0.5358 0.7678.7 0.8484.5 1.1493.8 1.5497.4 0.1811.3 0.5658 0.8778.7 1.0984.5 1.4593.8 1.8197.4 Burch et al. 0.520.1 0.5545.4 0.5965 0.785.2 0.7694.5 0.897.5 0.450.4 0.368.3 0.288.3
0.520.1 0.5844.9 0.6364.9 0.8184.5 1.194.7 1.697.8 0.450.1 0.567.9 0.587.9
Kumaran (1996) Kumaran (2002)
0.650.4 0.568.3 0.488.3
0.550.1 0.567.9 0.487.9
Kumaran (2002)
1.350.4 1.768.3 288.4
1.150.1 1.567.9 1.887.9
Kumaran (2002)
0.550.4 170.2 1.690.3 1.350.4 1.768.3 2.188.3 1.3611.3 4.558 6.878.7 984.5
150.5 2.170.9 791.3 1.150.1 1.567.9 1.987.9 0.8911.3 5.858 8.378.7 10.9
Kumaran (2002) Kumaran (2002) Burch et al.
12.593.8 17.997.4
14.493.8 18.497.4
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties or values from Chapter 15 to determine the surface film coefficients/ resistances. Table 11 lists the standard surface film coefficient values used in European standards.
Licensed for single user. © 2017 ASHRAE, Inc.
4.9
CODES AND STANDARDS
ASHRAE. 2010. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2010. ASTM. 2010. Standard terminology relating to thermal insulation. Standard C168-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test method for steady-state heat flux measurements and thermal transmission properties by means of the guarded-hot-plate apparatus. Standard C177-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test method for steady-state heat transfer properties of pipe insulation. Standard C335/C335M-10e1. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test method for steady-state thermal transmission properties by means of the heat flow meter apparatus. Standard C518-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard practice for selection of water vapor retarders for thermal insulation. Standard C755-10 (R2015). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2005. Standard classification of potential health and safety concerns associated with thermal insulation materials and accessories. Standard C930-05. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2013. Standard practice for calculating thermal transmission properties under steady-state conditions. Standard C1045-07 (R2013). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard test method for thermal performance of building materials and envelope assemblies by means of a hot box apparatus. Standard C1363-11. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2016. Standard specification for insulated vinyl siding. Standard D7793-16. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test methods for water vapor transmission of materials. Standard E96/E96M-10. American Society for Testing and Materials, West Conshohocken, PA.
Table 8
Typical Apparent Thermal Conductivity Values for Soils, W/(m·K) Recommended Values for Designa
Sands Silts Clays Loams
Normal Range
Lowb
Highc
0.6 to 2.5 0.9 to 2.5 0.9 to 1.6 0.9 to 2.5
0.78 1.64 1.12 0.95
2.25 2.25 1.56 2.25
a Reasonable
values for use when no site- or soil-specific data are available. conservative values for minimum heat loss through soil (e.g., use in soil heat exchanger or earth-contact cooling calculations). Values are from Salomone and Marlowe (1989). c Moderately conservative values for maximum heat loss through soil (e.g., use in peak winter heat loss calculations). Values are from Salomone and Marlowe (1989). b Moderately
Table 9
Typical Apparent Thermal Conductivity Values for Rocks, W/(m·K) Normal Range
Pumice, tuff, obsidian Basalt Shale Granite Limestone, dolomite, marble Quartzose sandstone
0.5 to 2.2 0.5 to 2.6 0.9 to 4.0 1.7 to 4.3 1.2 to 4.3 1.4 to 7.8
26.21
ASTM. 2009. Standard practices for air leakage site detection in building envelopes and air barrier systems. Standard E1186-03 (2009). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard specification for air barrier (AB) material or system for low-rise framed building walls. Standard E1677-11. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard test method for determining air leakage of air barrier assemblies. Standard E2357-11. American Society for Testing and Materials, West Conshohocken, PA. CAN/ULC. 2003. Standard for determination of log-term thermal resistance of closed-cell thermal insulating foams. CAN/ULC Standard S7702003. Standards Council of Canada, Ottawa, ON, and Underwriters Laboratories Canada, Toronto, ON. VDI. 1999. Environmental meteorology—Interactions between atmosphere and surfaces—Calculation of short-wave and long-wave radiation. Standard 3789 Part 2. Verein Deutscher Ingenieure (Association of German Engineers), Dusseldorf.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. Adams, L. 1971. Supporting cryogenic equipment with wood. Chemical Engineering (May):156-158.
Table 10
Surface Film Coefficients/Resistances Surface Emittance,
Position of Surface Indoor Horizontal Sloping at 45° Vertical Sloping at 45° Horizontal
Direction of Heat Flow Upward Upward Horizontal Downward Downward
Outdoor (any position) Any Wind (for winter) at 6.7 m/s Any Wind (for summer) at 3.4 m/s
Nonreflective = 0.90
Reflective = 0.20
= 0.05
hi
Ri
hi
Ri
hi
Ri
9.26 9.09 8.29 7.50 6.13
0.11 0.11 0.12 0.13 0.16
5.17 5.00 4.20 3.41 2.10
0.19 0.20 0.24 0.29 0.48
4.32 4.15 3.35 2.56 1.25
0.23 0.24 0.30 0.39 0.80
ho 34.0
Ro 0.030
—
—
—
—
22.7
0.044
—
—
—
—
Notes: 1. Surface conductance hi and ho measured in W/(m2·K); resistance Ri and Ro in (m2·K)/W. 2. No surface has both an air space resistance value and a surface resistance value. 3. Conductances are for surfaces of the stated emittance facing virtual blackbody surroundings at same temperature as ambient air. Values based on surface/air temperature difference of 5.6 K and surface temperatures of 21°C. 4. See Chapter 4 for more detailed information. 5. Condensate can have significant effect on surface emittance (see Table 2). Also, oxidation, corrosion, and accumulation of dust and dirt can dramatically increase surface emittance. Emittance values of 0.05 should only be used where highly reflective surface can be maintained over the service life of the assembly.
Table 11
European Surface Film Coefficients/Resistances
Position of Surface Indoors Horizontal, sloping till 45° Vertical, sloping beyond 45° Outdoors
Direction of Heat Flow Upward Downward Any direction
h, R, W/(m2·K) (m2·K)/W 10 6 7.7 25
0.1 0.17 0.13 0.04
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Licensed for single user. © 2017 ASHRAE, Inc.
26.22 ASHRAE. 1998. Standard method for determining and expressing the heat transfer and total optical properties of fenestration products. SPC 142. ASTM. 1985a. Guarded hot plate and heat flow meter methodology. Special Technical Publication STP 879. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1985b. Building applications of heat flux transducers. Special Technical Publication STP 885. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1988. Thermal insulation: Material and systems. Special Technical Publication STP 922. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1990. Insulation materials: Testing and applications. Special Technical Publication STP 1030. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1991. Insulation materials: Testing and applications, 2nd vol. Special Technical Publication STP 1116. American Society for Testing and Materials, West Conshohocken, PA. ASTM. Annual. Annual book of ASTM standards, vol. 04.06, Thermal insulation; building and environmental acoustics. American Society for Testing and Materials, West Conshohocken, PA. Bassett, M.R., and H.A. Trethowen. 1984. Effect of condensation on emittance of reflective insulation. Journal of Thermal Insulation 8(October):127. Bomberg, M.T., and M.K. Kumaran. 1986. A test method to determine air flow resistance of exterior membranes and sheathings. Journal of Thermal Insulation 9:224-235. Brandreth, D.A., ed. 1986. Advances in foam aging—A topic in energy conservation series. Caissa Editions, Yorklyn, DE. Brown, W.C., M.T. Bomberg, J. Rasmussen, and J. Ullett. 1993. Measured thermal resistance of frame walls with defects in the installation of mineral fibre insulation. Journal of Thermal Insulation and Building Envelopes 16(April):318-339. Burch, D.M., and A.O. Desjarlais. 1995. Water vapor measurements of lowslope roofing materials. Report NISTIR 5681. National Institute of Standards and Technology, Gaithersburg, MD. Burch, D.M., W.C. Thomas, and A.H. Fanney. 1992. Water vapor permeability measurements of common building materials. ASHRAE Transactions 98(2). CIBSE. 2006. Thermal properties of building structures. Chapter 3 in CIBSE Guide A: Environmental Design. The Chartered Institution of Building Services Engineers, London, U.K. Construction Specifications Canada. 1990. Tek-AID on air barrier systems. Toronto. Di Lenardo, B., W.C. Brown, W.A. Dalgleish, K. Kumaran, and G.F. Poirier. 1995. Technical guide for air barrier systems for exterior walls of lowrise buildings. Canadian Construction Materials Centre, National Research Council Canada, Ottawa, Ontario. Donnelly, R.G., V.J. Tennery, D.L. McElroy, T.G. Godfrey, and J.O. Kolb. 1976. Industrial thermal insulation: An assessment. Oak Ridge National Laboratory Reports TM-5283, TM-5515, and TID-27120. Glaser, P.E., I.A. Black, R.S. Lindstrom, F.E. Ruccia, and A.E. Wechsler. 1967. Thermal insulation systems—A survey. NASA Report SP5027. Goss, W.P., and R.G. Miller. 1989. Literature review of measurement and prediction of reflective building insulation system performance: 19001989. ASHRAE Transactions 95(2). Hedlin, C.P. 1985. Effect of insulation joints on heat loss through flat roofs. ASHRAE Transactions 91(2B):608-622. Hooper, F.C., and W.J. Moroz. 1952. The impact of aging factors on the emissivity of reflective insulations. ASTM Bulletin (May):92-95. ICC. 2007. 2007 supplement to the International Codes. International Code Council, Washington, D.C. ISO. 2003. Thermal performance of windows, doors, and shading devices— Detailed calculations. Standard 15099. International Organization for Standardization, Geneva. Karagiozis, A.N., and H.M. Salonvaara. 1999a. Hygrothermal performance of EIFS-clad walls: Effect of vapor diffusion and air leakage on the drying of construction moisture. Special Technical Publication STP 1352, pp. 32-51. American Society for Testing and Materials, West Conshohocken, PA. Karagiozis, A.N., and H.M. Salonvaara. 1999b. Whole building hygrothermal performance: Proceedings of the 5th Symposium on Building Physics in the Nordic Countries, Goteborg, vol. 2, pp. 745-753. C.E. Hagentoft and P.I. Sandberg, eds.
2017 ASHRAE Handbook—Fundamentals (SI) Kersten, M.S. 1949. Thermal properties of soils. University of Minnesota, Engineering Experiment Station Bulletin 28 (June). Korsgaard, V., and C.R. Pedersen. 1989. Transient moisture distribution in flat roofs with hygro diode vapor retarder. Proceedings of ASHRAE/ DOE/BTECC/CIBSE Conference on Thermal Performance of Exterior Envelopes of Buildings IV. Korsgaard, V., and C.R. Pedersen. 1992. Laboratory and practical experience with a novel water-permeable vapor retarder. Proceedings of ASHRAE/ DOE/BTECC/CIBSE Conference on Thermal Performance of Exterior Envelopes of Buildings V, pp. 480-490. Kuenzel, H.M. 1999. More moisture load tolerance of construction assemblies through the application of a smart vapor retarder. Proceedings of Thermal Performance of the Exterior Envelopes of Buildings VII, pp. 129-132. ASHRAE. Kumaran, M.K. 1989. Experimental investigation on simultaneous heat and moisture transport through thermal insulation. Proceedings of the Conseil International du Batiment/International Building Council (CIB) 11th International Conference 2:275-284. Kumaran, M.K. 1996. Heat, air and moisture transport. Final Report, vol. 3, task 3: Material properties. International Energy Agency Annex 24. Kumaran, M.K. 2002. A thermal and moisture transport database for common building and insulating materials. ASHRAE Research Project RP-1018, Final Report. National Research Council, Canada. Lander, R.M. 1955. Gas is an important factor in the thermal conductivity of most insulating materials. ASHRAE Transactions 61:151. Lecompte, J. 1989. The influence of natural convection in an insulated cavity on the thermal performance of the wall. Special Technical Publication STP 1000:397-420. American Society for Testing and Materials, West Conshohocken, PA. Lewis, W.C. 1967. Thermal conductivity of wood-base fiber and particle panel materials. Forest Products Laboratory, Research Paper FPL 77, June. Lewis, J.E. 1979. Thermal evaluation of the effects of gaps between adjacent roof insulation panels. Journal of Thermal Insulation (October):80-103. Lotz, W.A. 1964. Vapor barrier design, neglected key to freezer insulation effectiveness. Quick Frozen Foods (November):122. Lotz, W.A. 1969. Facts about thermal insulation. ASHRAE Journal (June): 83-84. MacLean, J.D. 1941. Thermal conductivity of wood. ASHVE Transactions 47:323. McGowan, A.G. 2007. Catalog of material thermal property data (RP-905). ASHRAE Research Project, Final Report. NIST. 2000. NIST standard reference database 81: NIST heat transmission properties of insulating and building materials. U.S. Department of Commerce, National Institute of Standards and Materials, Gaithersburg, MD. srdata.nist.gov/insulation/. Nottage, H.B. 1947. Thermal properties of building materials used in heat flow calculations. ASHVE Transactions 53:215-243. Ojanen, T., R. Kohonen, and M.K Kumaran. 1994. Modeling heat, air, and moisture transport through building materials and components. Chapter 2 in Manual MNL 18, Moisture control in buildings. American Society for Testing and Materials, West Conshohocken, PA. Ostrogorsky, A.G., and L.R. Glicksman. 1986. Laboratory tests of effectiveness of diffusion barriers. Journal of Cellular Plastics 22:303. Pelanne, C.M. 1979. Thermal insulation heat flow measurements: Requirements for implementation. ASHRAE Journal 21(3):51. Rasmussen J., W.C. Brown, M. Bomberg, and J.M. Ullett. 1993. Measured thermal performance of frame walls with defects in the installation of mineral fibre insulation. Proceedings of the 3rd Symposium on Building Physics in the Nordic Countries, pp. 209-217. Robinson, H.E., F.J. Powlitch, and R.S. Dill. 1954. The thermal insulation value of airspaces. Housing Research Paper 32, Housing and Home Finance Agency. Robinson, H.E., F.J. Powell, and L.A. Cosgrove. 1957. Thermal resistance of airspaces and fibrous insulations bounded by reflective surfaces. National Bureau of Standards, Building Materials and Structures Report BMS 151. Rowley, F.B., and A.B. Algren. 1937. Thermal conductivity of building materials. University of Minnesota Bulletin #12, Minneapolis. Rowley, F.B., R.C. Jordan, C.E. Lund, and R.M. Lander. 1952. Gas is an important factor in the thermal conductivity of most insulating materials. ASHVE Transactions 58:155.
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Heat, Air, and Moisture Control in Building Assemblies—Material Properties Sabine, H.J., M.B. Lacher, D.R. Flynn, and T.L. Quindry. 1975. Acoustical and thermal performance of exterior residential walls, doors and windows. NBS Building Science Series 77. National Institute of Standards and Technology, Gaithersburg, MD. Salomone, L.A., and J.I. Marlowe. 1989. Soil and rock classification according to thermal conductivity: Design of ground-coupled heat pump systems. EPRI CU-6482. Electric Power Research Institute, Palo Alto, CA. Seiffert, K. 1970. Damp diffusion and buildings. Elsevier, Amsterdam, the Netherlands. Shu, L.S., A.E. Fiorato, and J.W. Howanski. 1979. Heat transmission coefficients of concrete block walls with core insulation. Proceedings of the ASHRAE/DOE-ORNL Conference on Thermal Performance of the Exterior Envelopes of Buildings, ASHRAE SP 28, pp. 421-435. Simons, E. 1955. In-place studies of insulated structures. Refrigerating Engineering 63:40, 128. Touloukian, Y.S., R.W. Powell, C.Y. Ho, and I.G. Clemens. 1970. Thermophysical properties of matter. Thermal conductivity data tables of nonmetallic solids. IFI/Plenum, New York. Tye, R.P. 1985. Upgrading thermal insulation performance of industrial processes. Chemical Engineering Progress (February):30-34. Tye, R.P. 1986. Effects of product variability on thermal performance of thermal insulation. Proceedings of the First Asian Thermal Properties Conference, Beijing, People’s Republic of China. Tye, R.P. 1988. Aging of cellular plastics: A comprehensive bibliography. Journal of Thermal Insulation 11:196-222. Tye, R.P., and A.O. Desjarlais. 1981. Performance characteristics of foam-inplace urea formaldehyde insulation. ORNL/Sub-78/86993/1. Oak Ridge National Laboratory, Oak Ridge, TN. Tye, R.P., and A.O. Desjarlais. 1983. Factors influencing the thermal performance of thermal insulations for industrial applications. In Thermal insulation, materials, and systems for energy conservation in the ’80s, F.A. Govan, D.M. Greason, and J.D. McAllister, eds. ASTM STP 789:733748. Valore, R.C., 1988. Thermophysical properties of masonry and its constituents, parts I and II. International Masonry Institute, Washington, D.C. Van Geem, M.G. 1985. Thermal transmittance of concrete block walls with core insulation. ASHRAE Transactions 91(2). Verschoor, J.D. 1977. Effectiveness of building insulation applications. USN/CEL Report CR78.006—NTIS AD-AO53 452/9ST. Verschoor, J.D., and P. Greebler. 1952. Heat transfer by gas conductivity and radiation in fibrous insulations. ASME Transactions 74(6):961-968. Wilkes, K.E. 1979. Thermophysical properties data base activities at OwensCorning Fiberglas. Proceedings of the ASHRAE/DOE-ORNL Conference on Thermal Performance of the Exterior Envelopes of Buildings, ASHRAE SP 28, pp. 662-677.
26.23
Wilkes, K.E., and P.W. Childs. 1992. Thermal performance of fiberglass and cellulose attic insulations. Proceedings of the ASHRAE/DOE/BTECC/ CIBSE Conference on Thermal Performance of the Exterior Envelopes of Buildings V, pp. 357-367. Wilkes, K.E., and J.L. Rucker. 1983. Thermal performance of residential attic insulation. Energy and Buildings 5:263-277. Yarbrough, D.W. 1983. Assessment of reflective insulations for residential and commercial applications. ORNL/TM-8891. Oak Ridge National Laboratory, Oak Ridge, TN.
BIBLIOGRAPHY ASHRAE. 1996. Building insulation system thermal anomalies. ASHRAE Research Report RP-758. Enermodal Engineering, Ltd. ASTM. 1974. Heat transmission measurements in thermal insulations. Special Technical Publication STP 544. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1978. Thermal transmission measurements of insulation. Special Technical Publication STP 660. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1980. Thermal insulation performance. Special Technical Publication STP 718. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 1983. Thermal insulations, materials, and systems for energy conservation in the ’80s. Special Technical Publication STP 789. American Society for Testing and Materials, West Conshohocken, PA. Cardenes, T.J., and G.T. Bible. 1987. The thermal properties of wood—Data base. American Society of Testing and Materials, West Conshohocken, PA. CIMA. 2007. Measured thermal resistances for cellulose insulation products commercially available in 2007. Report prepared by R&D Services, Inc., for the Cellulose Insulation Manufacturers Association. Hedlin, C.P. 1988. Heat flow through a roof insulation having moisture contents between 0 and 1% by volume, in summer. ASHRAE Transactions 94(2):1579-1594. Pelanne, C.M. 1977. Heat flow principles in thermal insulation. Journal of Thermal Insulation 1:48. Raznjevic, K. 1976. Thermal conductivity tables. In Handbook of thermodynamic tables and charts. McGraw-Hill, New York. Rowley, F.B., and A.B. Algren. 1932. Heat transmission through building materials. University of Minnesota Bulletin 8, Minneapolis. Yarbrough, D.W., R.S. Graves, D.L. McElroy, A.O. Desjarlais, and R.P. Tye. 1987. The thermal resistance of spray-applied fiber insulations. Journal of Thermal Insulation 11(2):81-95.
Related Commercial Resources
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Related Commercial Resources CHAPTER 27
HEAT, AIR , AND MOISTURE CONTROL IN BUILDING ASSEMBLIES—EXAMPLES HEAT TRANSFER ................................................................... One-Dimensional Assembly U-Factor Calculation ................. Two-Dimensional Assembly U-Factor Calculation................. MOISTURE TRANSPORT .......................................................
27.1 27.1 27.3 27.7
Wall with Insulated Sheathing ................................................. 27.7 Vapor Pressure Profile (Glaser or Dew-Point) Analysis ......... 27.8 TRANSIENT HYGROTHERMAL MODELING ..................... 27.10 AIR MOVEMENT................................................................... 27.12
HERMAL and moisture design as well as long-term performance must be considered during the planning phase of buildings. Installing appropriate insulation layers and taking appropriate air and moisture control measures can be much more economical during construction than later. Design and material selection should be based on
toward nonlinearity, and both input and output values include great uncertainty. Air movement is even more difficult to characterize than moisture transport. Engineering makes use of the continuum in understanding from physical principles, to simple applications, to complex applications, to design guidance. Complex design applications can be handled by computers; however, this chapter begins by presenting simpler examples as a learning tool. Because complex applications are built up from simpler ones, understanding the simpler applications ensures that a critical engineering oversight of complex (computer) applications is retained. Computers have facilitated widespread use of twoand three-dimensional analysis as well as transient (time-dependent) calculations. As a consequence, steady-state calculations are less widely used. Design guidance, notably guidance regarding use of air and vapor barriers, faces changes in light of sophisticated transient calculations. ASHRAE Standard 160 creates a framework for using transient hygrothermal calculations in building envelope design. However, designers should recognize the limitations of these tools, as discussed in the following sections, and the need for continued advancements in the methods of analysis and understanding of heat and moisture migration in buildings. The following definitions pertain to heat transfer properties of envelope assemblies (see Chapter 25).
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T
• • • • • • • •
Building use Interior and exterior climate Space availability Thermal and moisture properties of materials Other properties required by location of materials Durability of materials Compatibility with adjacent materials Performance expectations of the assembly
Designers and builders often rely on generic guidelines and past building practice as the basis for system and material selection. Although this approach may provide insight for design decisions, selections and performance requirements should be set through engineering analysis of project-specific criteria. Recent developments have increased the capabilities of available tools and methods of thermal and moisture analysis. This chapter draws on Chapter 25’s fundamental information on heat, air and moisture transport in building assemblies, as well as Chapter 26’s material property data. Examples here demonstrate calculation of heat, moisture, and air transport in typical assemblies. For design guidance for common building envelope assemblies and conditions, see Chapter 44 of the 2015 ASHRAE Handbook—HVAC Applications. Insulation specifically for mechanical systems is discussed in Chapter 23. For specific industrial applications of insulated assemblies, see the appropriate chapter in other ASHRAE Handbook volumes. In the 2014 ASHRAE Handbook—Refrigeration, for refrigerators and freezers, see Chapters 15, 16, and 17; for insulation systems for refrigerant piping, see Chapter 10; for refrigerated-facility design, see Chapters 23 and 24; for trucks, trailers, rail cars, and containers, see Chapter 25; for marine refrigeration, see Chapter 26. For environmental test facilities, see Chapter 37 in the 2002 ASHRAE Handbook—Refrigeration. Engineering practice is predicated on the assumption that performance effects can be viewed in functional format, where discrete input values lead to discrete output values that may be assessed for acceptability. Heat transfer in solids lends itself to engineering analysis because material properties are relatively constant and easy to characterize, the transport equations are well established, analysis results tend toward linearity, and, for well-defined input values, output values are well defined. Airflow and moisture transport analysis, in contrast, is difficult: material properties are difficult to characterize, transport equations are not well defined, analysis results tend The preparation of this chapter is assigned to TC 4.4, Building Materials and Building Envelope Performance.
27.1 Copyright © 2017, ASHRAE
Symbol
Definition
RSystem, CSystem
System resistance (conductance); surface-to-surface resistance (conductance) for all materials in wall, including parallel paths for framing Assembly resistance (transmittance); air-to-air thermal resistance (transmittance), equal to system value plus film resistances (conductances) Assembly thermal transmittance, including thermal bridges (i.e., UAssembly plus bridge conductances)
RAssembly, UAssembly UWhole
Note: For all code applications that call for U, UWhole should be used.
1. 1.1
HEAT TRANSFER
ONE-DIMENSIONAL ASSEMBLY U-FACTOR CALCULATION
Wall Assembly U-Factor The assembly U-factor for a building envelope assembly determines the rate of steady-state heat conduction through the assembly. One-dimensional heat flow through building envelope assemblies is the starting point for determining whole-building heat transmittance. Example 1. Calculate the system R-value RSystem, assembly total resistance (RAssembly), and UAssembly-factor of the sandwich panel assembly shown in Figure 1; assume winter conditions when selecting values for air films from Table 3 in Chapter 26. Solution: Determine indoor and outdoor air film resistances from Table 3 in Chapter 26, and thermal resistance of all components from Table 1 in that chapter. If any elements are described by conductivity
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27.2
2017 ASHRAE Handbook—Fundamentals (SI)
(independent of thickness) rather than thermal resistance (thicknessdependent), then calculate the resistance. R, (m2 ·K)/W
Element 1. Outdoor air film 2. Vinyl siding (hollow backed) 3. Vapor-permeable felt 4. Oriented strand board (OSB), 11 mm 5. 150 mm expanded polystyrene, extruded (smooth skin) 6. 13 mm gypsum wallboard 7. Indoor air film Total
0.030 0.107 0.011 0.11 5.28 0.079 0.120 5.70
The conductivity k of expanded polystyrene is 0.035 W/(m·K). For 150 mm thickness, Rfoam = x/k = 0.150/0.035 = 4.29 (m2 ·K)/W To calculate the system’s R-value in the example, sum the R-values of the system components only, disregarding indoor and outdoor air films. RSystem = 0.107 + 0.011 + 0.07 + 4.29 + 0.079 = 4.56 (m2 ·K)/W The assembly R-value (RAssembly) consists of the system’s R-value plus the thermal resistance of the interior and exterior air films. RAssembly = Ro + RSystem+ Ri = 4.70 (m2 ·K)/W
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The wall’s UAssembly-factor is 1/RAssembly , or 0.21 W/(m2 ·K).
Roof Assembly U-Factor Example 2. Find the U-factor of the commercial roof assembly shown in Figure 2; assume summer conditions when selecting values for air films from Table 3 in Chapter 26. Solution: The calculation procedure is similar to that shown in Example 1. Note the U-factor of nonvertical assemblies depends on the direction of heat flow [i.e., whether the calculation is for winter (heat flow up) or summer (heat flow down)], because the resistances of indoor air films and plane air spaces in ceilings differ, based on the heat flow
Fig. 1 Structural Insulated Panel Assembly (Example 1)
Fig. 2 Roof Assembly (Example 2)
direction (see Table 3 in Chapter 26). The effects of mechanical fasteners are not addressed in this example. R, (m2 ·K)/W
Element 1. Indoor air film 2. 100 mm concrete, 1920 kg/m3,k = 1.1 3. 75 mm cellular polyisocyanurate (gasimpermeable facers) 4. 25 mm mineral fiberboard 5. 10 mm built-up roof membrane 6. Outdoor air film
0.16 0.09 4.97
Total
0.52 0.06 0.04 5.83
Using UAssembly = 1/RAssembly, the UAssembly-factor is 0.17 W/(m2 ·K).
Attics During sunny periods, unconditioned attics may be hotter than outdoor air. Peak attic temperatures on a hot, sunny day may be 10 to 45 K above outdoor air temperature, depending on factors such as shingle color, roof framing type, air exchange rate through vents, and use of radiant barriers. Therefore, simple one-dimensional solutions cannot be offered for attics. Energy efficiency estimates can be obtained using models such as Wilkes (1991).
Basement Walls and Floors Heat transfer through basement walls and floors to the ground depends on the following factors: (1) the difference between the air temperature in the room and that of the ground and outside air, (2) the material of the walls or floor, and (3) the thermal conductivity of surrounding earth. The latter varies with local conditions and is usually unknown. Because of the great thermal inertia of surrounding soil, ground temperature varies with depth, and there is a substantial time lag between changes in outdoor air temperatures and corresponding changes in ground temperatures. As a result, ground-coupled heat transfer is less amenable to steadystate representation than above-grade building elements. However, there are several simplified procedures for estimating groundcoupled heat transfer. These fall into two main categories: (1) those that reduce the ground heat transfer problem to a closed-form solution, and (2) those that use simple regression equations developed from statistically reduced multidimensional transient analyses. Closed-form solutions, including Latta and Boileau’s (1969) procedure discussed in Chapter 17, generally reduce the problem to onedimensional, steady-state heat transfer. These procedures use simple, “effective” U-factors or ground temperatures or both. Methods differ in the various parameters averaged or manipulated to obtain these effective values. Closed-form solutions provide acceptable results in climates that have a single dominant season, because the dominant season persists long enough to allow a reasonable approximation of steady-state conditions at shallow depths. The large errors (percentage) that are likely during transition seasons should not seriously affect building design decisions, because these heat flows are relatively insignificant compared to those of the principal season. The ASHRAE arc-length procedure (Latta and Boileau 1969) is a reliable method for wall heat losses in cold winter climates. Chapter 17 discusses a slab-on-grade floor model developed by one study. Although both procedures give results comparable to transient computer solutions for cold climates, their results for warmer U.S. climates differ substantially. Research conducted by Dill et al. (1945) and Hougten et al. (1942) indicates a heat flow of approximately 6.3 W/m2 through an uninsulated concrete basement floor with a temperature difference of 11 K between the basement floor and the air 150 mm above it. A U-factor of 5.7 W/(m2 ·K) is sometimes used for concrete basement floors on the ground. For basement walls below grade, the temperature difference for winter design conditions is greater than for the floor. Test results indicate that, at the mid-height of the below-grade
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Heat, Air , and Moisture Control in Building Assemblies—Examples portion of the basement wall, the unit area heat loss is approximately twice that of the floor. For small concrete slab floors (equal in area to a 7.5 by 7.5 m house) in contact with the ground at grade level, tests indicate that heat loss can be calculated as proportional to the length of exposed edge rather than total area. This amounts to 1.4 W per linear metre of exposed edge per degree temperature difference between indoor air and the average outdoor air temperature. This value can be reduced appreciably by installing insulation under the ground slab and along the edge between the floor and abutting walls. In most calculations, if the perimeter loss is calculated accurately, no other floor losses need to be considered. Chapters 17 and 18 contain heat transfer and load calculation guidance for floors on grade and at different depths below grade. The second category of simplified procedures uses transient twodimensional computer models to generate ground heat transfer data, which are then reduced to compact form by regression analysis (Mitalas 1982, 1983; Shipp 1983). These are the most accurate procedures available, but the database is very expensive to generate. In addition, these methods are limited to the range of climates and constructions specifically examined. Extrapolating beyond the outer bounds of the regression surfaces can produce significant errors. Guide details and recommendations related to application of concepts for basements are provided in Chapter 44 of the 2015 ASHRAE Handbook—HVAC Applications. Detailed analysis of heat transfer through foundation insulation may also be found in the Building Foundation Design Handbook (Labs et al. 1988).
1.2 TWO-DIMENSIONAL ASSEMBLY U-FACTOR CALCULATION The following examples show three methods of two-dimensional, steady-state conductive heat transfer analysis through wall assemblies. They offer approximations to overall rates of heat transfer (U-factor) when assemblies contain a layer composed of dissimilar materials. The methods are described in Chapter 25. The parallelpath method is used when the thermal conductivity of the dissimilar materials in the layer are rather close in value (within the same order of magnitude), as with wood-frame walls. The isothermal-planes method is appropriate for materials with conductivities moderately different from those of adjacent materials (e.g., masonry). The zone method and the modified zone method are appropriate for materials with a very high difference in conductivity (two orders of magnitude or more), such as with assemblies containing metal. Two-dimensional, steady-state heat transfer analysis is often conducted using computer-based finite difference methods. If the resolution of the analysis is sufficiently fine, computer methods provide better simulations than any of the methods described here, and the results typically show better agreement with measured values. The methods described here do not take into account heat storage in the materials, nor do they account for varying material properties (e.g., when thermal conductivity is affected by moisture content or temperature). Transient analysis is often used in such cases.
27.3
and the fraction of headers is 0.04. For studs 600 mm OC, the respective values are 0.78, 0.18, and 0.04. These fractions contain an allowance for multiple studs, plates, sills, extra framing around windows, headers, and band joists. These assumed framing fractions are used in Example 3, to illustrate the importance of including the effect of framing in determining a building’s overall thermal conductance. The actual framing fraction should be calculated for each specific construction. Example 3. Calculate the UAssembly-factor of the 38 by 90 mm stud wall shown in Figure 3. The studs are at 400 mm OC. There is 90 mm mineral fiber batt insulation (R-2.3) in the stud space. The inside finish is 13 mm gypsum wallboard, and the outside is finished with rigid foam insulating sheathing (R-0.7) and vinyl siding. The insulated cavity occupies approximately 75% of the transmission area; the studs, plates, and sills occupy 21%; and the headers occupy 4%. Solution: If the R-values of building elements are not already specified, obtain the R-values from Tables 1 and 3 of Chapter 26. Assume R = 7.0 (m2 ·K)/W for the wood framing. Also, assume the headers are solid wood, and group them with the studs, plates, and sills. Two simple methods may be used to determine the U-factor of wood frame walls: parallel path and isothermal planes. For highly conductive framing members such as metal studs, the modified zone method must be used. Parallel-Path Method: R (Insulated R (Studs, Plates, Cavity), and Headers), (m2 ·K)/W (m2 ·K)/W
Element 1. Outdoor air film, 6.7 m/s wind 2. Vinyl siding (hollow-backed) 3. Rigid foam insulating sheathing 4. Mineral fiber batt insulation, 90 mm 5. Wood stud, nominal 38 by 90 mm 6. Gypsum wallboard, 13 mm 7. Indoor air film, still air
0.03 0.11 0.70 2.30 — 0.08 0.12 R1 = 3.31
0.03 0.11 0.70 — 0.77 0.08 0.12 R2 = 1.81
Individual U-factors are reciprocals of the R-value, so U1 = 0.301 and U2 = 0.553 W/(m2 ·K). If the wood framing is accounted for using the parallel-path flow method, the wall’s U-factor is determined using Equation (15) from Chapter 25. The fractional area of insulated cavity is 0.75 and the fractional area of framing members is 0.25. UAssembly = (0.75 0.301) + (0.25 0.553) = 0.36 W/(m2 ·K) RAssembly =1/UAssembly = 2.78 (m2 ·K)/W With the isothermal-planes method, the fractional areas are applied only to the building layer that contains the studs and cavity-fill insulation. The average R-value for this layer (Ravs) is added to the R-values of the other components for a total R for the assembly.
Wood-Frame Walls The assembly R-values and U-factors of wood-frame walls can be calculated by assuming either parallel heat flow paths through areas with different thermal resistances or by assuming isothermal planes. Equation (15) in Chapter 25 provides the basis for the two methods. The framing factor expresses the fraction of the total building component (wall or roof) area that is framing. The value depends on the specific type of construction, and may vary based on local construction practices, even for the same type of construction. For stud walls 400 mm on center (OC), the fraction of insulated cavity may be as low as 0.75, where the fraction of studs, plates, and sills is 0.21
Fig. 3 (A) Wall Assembly for Example 3, with Equivalent Electrical Circuits: (B) Parallel Path and (C) Isothermal Planes
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27.4
2017 ASHRAE Handbook—Fundamentals (SI)
Isothermal-Planes Method: R (Studs, Plates, R (Stud Cavity Average Cavity, Elements), and Headers), (m2 ·K)/W (m2 ·K)/W
Element 1. Outdoor air film, 6.7 m/s wind 2. Vinyl siding (hollow-backed) 3. Rigid foam insulating sheathing 4. Mineral fiber batt insulation, 90 mm 5. Wood stud, nominal 38 by 90 mm 6. Gypsum wallboard, 13 mm 7. Indoor air film, still air
2.30 0.77
0.03 0.11 0.70 1.54 (Ravs)
Fig. 4 0.08 0.12 RT = 2.58
The average R-value Ravs of the stud cavity is calculated using the fractional area of stud and insulation using Equation (15) from Chapter 25. Uavs = 0.75(1/2.3) + 0.25(1/0.77) = 0.651 Ravs = 1/Uavs = 1.54 (m2 ·K)/W
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If the wood framing is included using the isothermal-planes method, the U-factor of the wall is determined using Equations (10) and (11) from Chapter 25 as follows: RAssembly = 2.58 (m2 ·K)/W UAssembly = 1/RAssembly = 0.388 W/(m2 ·K) For a frame wall with a 600 mm OC stud space, the assembly R-value is 2.76 (m2 ·K)/W. Similar calculation procedures may be used to evaluate other wall designs, except those with thermal bridges.
Masonry Walls The average overall R-values of masonry walls can be estimated by assuming a combination of layers in series, one or more of which provides parallel paths. This method is used because heat flows laterally through block face shells so that transverse isothermal planes result. Average total resistance RAssembly is the sum of the resistances of the layers between such planes, each layer calculated as shown in Example 4. Example 4. Calculate the overall thermal resistance and average U-factor of the 194 mm thick insulated concrete block wall shown in Figure 4. The two-core block has an average web thickness of 25 mm and a face shell thickness of 32 mm. Overall block dimensions are 194 by 194 by 395 mm. Measured thermal resistances of 1700 kg/m3 concrete and 110 kg/m3 expanded perlite insulation are 0.70 and 20 (m2 ·K)/W, respectively. (Note: This type of insulation is no longer commonly used with concrete block walls. This historical example has been retained because it is a simplified case with supporting experimental data.) Solution: The equation used to determine the overall thermal resistance of the insulated concrete block wall is derived from Equations (7) and (15) from Chapter 25 and is given below: aw ac U Avg = ------ + ----Rw Rc
1 R Avg = -----------U Avg
RAssembly = Ri + Rf + Ravg + Ro where RT(av) = overall thermal resistance based on assumption of isothermal planes Ri = thermal resistance of inside air surface film (still air) Ro = thermal resistance of outside air surface film (6.7 m/s wind) Rf = total thermal resistance of face shells Rc = thermal resistance of cores between face shells Rw = thermal resistance of webs between face shells aw = fraction of total area transverse to heat flow represented by webs of blocks ac = fraction of total area transverse to heat flow represented by cores of blocks
Insulated Concrete Block Wall (Example 4)
From the information given and the data in Tables 1 and 3 in Chapter 26, determine the values needed to compute the overall thermal resistance. Ri Ro Rf Rc Rw aw ac
= = = = = = =
0.12 0.03 2 0.032 0.70 = 0.045 (0.194 2 0.032) 20 = 2.60 (0.194 2 0.032) 0.70 = 0.091 3 25/395 = 0.190 1 0.190 = 0.810
Using the equation given, the overall thermal resistance and average Ufactor are calculated as follows: Uavs = 0.190/0.09 + 0.810/2.60 = 2.42 Ravs = 1/Uavs = 0.413 (m2 ·K)/W RAssembly = 0.12 + 0.045 + 0.413 + 0.03 = 0.608 (m2 ·K)/W UAssembly = 1/RAssembly = 1.64 W/(m2 ·K) Based on guarded hot-box tests of this wall without mortar joints, Tye and Spinney (1980) measured the assembly R-value for this insulated concrete block wall as 0.551 (m2 ·K)/W.
Assuming parallel heat flow only, the calculated resistance is higher than that calculated on the assumption of isothermal planes. The actual resistance generally is some value between the two calculated values. In the absence of test values, examination of the construction usually reveals whether a value closer to the higher or lower calculated R-value should be used. Generally, if the construction contains a layer in which lateral conduction is high compared with transmittance through the construction, the calculation with isothermal planes should be used. If the construction has no layer of high lateral conductance, the parallel heat flow calculation should be used. Hot-box tests of insulated and uninsulated masonry walls constructed with block of conventional configuration show that thermal resistances calculated using the isothermal planes heat flow method agree well with measured values (Shu et al. 1979; Valore 1980; Van Geem 1985). Neglecting horizontal mortar joints in conventional block can result in thermal transmittance values up to 16% lower than actual, depending on the masonry’s density and thermal properties, and 1 to 6% lower, depending on the core insulation material (McIntyre 1984; Van Geem 1985). For aerated concrete block walls, other solid masonry, and multicore block walls with full mortar joints, neglecting mortar joints can cause errors in R-values up to 40% (Valore 1988). Horizontal mortar joints, usually found in concrete block wall construction, are neglected in Example 4.
Constructions Containing Metal Curtain and metal stud-wall constructions often include metallic and other thermal bridges, which can significantly reduce the thermal resistance. However, the capacity of the adjacent facing materials to transmit heat transversely to the metal is limited, and some contact resistance between all materials in contact limits the reduction. Contact resistances in building structures are only 0.01 to 0.1 (m2 ·K)/W, too small to be of concern in many cases. For these
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Heat, Air , and Moisture Control in Building Assemblies—Examples metal stud/wall constructions, the recommended approach is the modified zone method (see Example 7) or two-dimensional analysis software such as THERM. Thermal characteristics for panels of sandwich construction can be computed by combining the thermal resistances of the layers. R-values for assembled sections should be determined on a representative sample by using a hot-box method. If the sample is a wall section with air cavities on both sides of fibrous insulation, the sample must be of representative height because convective airflow can contribute significantly to heat flow through the test section. Computer modeling can also be useful, but all heat transfer mechanisms must be considered. The metal studs in Example 5 are 90 mm deep and placed at 400 mm on center. The metal member is only 0.5 mm thick, but it is in contact with adjacent facings over a 32 mm wide area. The steel member is 32 mm deep, has a thermal resistance of approximately 0.0019 (m2 ·K)/W, and is virtually isothermal. For this insulated steel frame wall, Farouk and Larson (1983) measured an assembly R-value of 1.16 (m2 ·K)/W. For the same assembly, the recommended modified zone method (see Example 5) gives an assembly R-value of 1.18 (m2 ·K)/W. Two-dimensional analysis (THERM) yields Uav = 1.0 or R = 1.00 (m2 ·K)/W. ASHRAE/IES Standard 90.1 describes how to determine the thermal resistance of wall assemblies containing metal framing by using insulation/framing adjustment factors in Table A9.2B of the standard. For 38 by 90 mm steel framing, 400 mm OC, Fc = 0.50. Using the correction factor method, an assembly R-value of 1.13 (m2 ·K)/ W [0.08 + 1.94(0.50) + 0.08] is obtained for the wall described here.
Zone Method of Calculation For structures with widely spaced metal members of substantial cross-sectional area, the isothermal planes method can give thermal resistance values that are too low. For these constructions, the zone method can be used. This method involves two separate computations: one for a chosen limited portion, zone A, containing the highly conductive element; the other for the remaining portion of simpler construction, zone B. The two computations are then combined using the parallel-flow method, and the average transmittance per unit overall area is calculated. The basic laws of heat transfer are applied by adding the area conductances CA of elements in parallel, and adding area resistances R/A of elements in series. The modified zone method improves on the zone method with better guidance for defining the widths of zones A and B.
Modified Zone Method for Metal Stud Walls with Insulated Cavities The modified zone method is similar to the parallel-path and zone methods; all three are based on parallel-path calculations. Figure 5 shows the width w of the zone of thermal anomalies around a metal stud. This zone can be assumed to equal the length of the stud flange L (parallel-path method), or can be calculated as a sum of the length of stud flange and a distance double that from wall
27.5
surface to metal di (zone method). In the modified zone method, the width of the zone depends on three parameters: • Ratio between thermal resistivity of sheathing material and cavity insulation • Size (depth) of stud • Thickness of sheathing material Example 5. Calculate the U-factor of the wall section shown in Figure 5 using the modified zone method. Solution: The wall cross section is divided into two zones: the zone of thermal anomalies around the metal stud (zone W), and the cavity zone (zone cav). Wall material layers are grouped into exterior and interior surface sections A (sheathing, siding) and B (wallboard), and interstitial sections I and II (cavity insulation, metal stud flange). Assuming that the wall materials in section A are thicker than those in section B, as shown, they can be described as follows:
di dj n
m
i=1
j=1
where n = number of material layers (of thickness di) between metal stud flange and wall surface for section A m = number of material layers (of thickness dj) for section B Then, the width W of zone W can be estimated by W = L + zf d i n
i=1
where L = stud flange size di = thickness of material layers in section A zf = zone factor, shown in Figure 6 (zf = 2 for zone method) Kosny and Christian (1995) developed the modified zone method and verified its accuracy for over 200 simulated cases of metal frame walls with insulated cavities. For all configurations considered, the discrepancy between results were within 2%. Hot-box-measured R-values for 15 metal stud walls tested by Barbour et al. (1994) were compared with results obtained by Kosny and Christian (1995) and McGowan and Desjarlais (1997). The modified zone method was found to be the most accurate simple method for estimating the clear-wall Rvalue of light-gage steel stud walls with insulated cavities. However, this analysis does not apply to construction with metal sheathing. Also, ASHRAE Standard 90.1 may require a different method of analysis. Step 1. Determine zone factor zf , and the ratio of the exterior sheathing material’s resistivity to the cavity material’s resistivity. Resistivity r is the reciprocal of conductivity; Table 1 in Chapter 26 lists conductivities of various materials. Element Stud spacing Resistivity of sheathing material Resistivity of cavity insulation Ratio ri /rins Zone factor from chart
Symbol
Value
Units
s ri rins
0.406 34.67 23.92 1.449 1.71
m (m·K)/W (m·K)/W (no units) (no units)
zf
Step 2. Calculate width W of affected zone W:
Fig. 5 Wall Section and Equivalent Electrical Circuit (Example 5)
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2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 7
Corner Composed of Homogeneous Material Showing Locations of Isotherms
The particular thermal resistances of the zone elements are then calculated. Zone W is the zone at the web of the metal member. For width W, the thermal conductance C is calculated as the sum of the contributory areas. Because the thickness of the web of the metal member is dI and the length along the flange section is L, dI W – dI L W–L C I = --------------- C ins + ----- C met and C II = -------------- C ins + ----- C met W W W W
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Using resistance rather than conductance, the contributing R-values are calculated as I
II
I
II
R met R ins W R met R ins W R I = ----------------------------------------------------------- and R II = --------------------------------------------------------II II II I I I L R ins – R met + WR met d I R ins – R met + WR met At the cavity, the sum of the series R-values is
Fig. 6
Rcav
Modified Zone Factor for Calculating R-Value of Metal Stud Walls with Cavity Insulation W = L + zf
RW
= R A + R B + R I + 2R II
The total conductivity across the length s is proportional to the contributing lengths of zone W and the cavity:
Element Cavity thickness Thickness of metal Interior dimension between flanges Thickness of exterior insulating materials Flange length Affected zone thickness
Symbol
Value
Units
ds dII dI
0.089 0.0010 0.088 0.051 0.038 0.125
m m m m m m
L W
Step 3. Calculate the exterior and interior thermal resistances, using conductivity or thermal resistance values from step 1. Symbol Value
Exterior materials Thickness of first exterior material Resistivity of first exterior material Resistance of first material Resistances of other materials Sum of resistances of exterior materials Interior materials Thickness of interior material Resistivity of interior material Resistance of interior material
II
In zone W, the sum of the R-values is
di
Element
I
= R A + R B + R ins + 2R ins
Units m (m·K)/W (m2 ·K)/W
RA
0.0381 34.7 1.32 0.145 1.47
dj rj RB
0.00127 6.24 0.043
m (m·K)/W (m2 ·K)/W
de re
(m2 ·K)/W
Step 4. Calculate the thermal resistance of the sections in zone around the metal element. The building elements in series from outside to inside are shown in Figure 7. Element
Symbol
Value
Units
Resistivity of steel I Rins II R ins I Rmet II R met
rmet I dxri II dxri I d xrmet II dxrmet
0.0208 2.10 0.024 0.00183 0.00002
(m·K)/W (m2 ·K)/W (m2 ·K)/W (m2 ·K)/W (m2 ·K)/W
cav W C tot = ----- C W + --------- C cav s s or
RW Rcav s R tot = -------------------------------------------------------------------------W R cav – R W + s R W
Element
Symbol
Value
Resistance at web Resistance at flange Sum of resistances at cavity Sum of resistances at zone W Assembly R Assembly U
RI RII Rcav RW RAssembly UAssembly
0.203 0.00007 3.66 1.713 2.71 0.369
Units (m2 ·K)/W (m2 ·K)/W (m2 ·K)/W (m2 ·K)/W (m2 ·K)/W W/(m2 ·K)
In this example, the calculated total R-value for the wall is 2.71 (m2 ·K)/W, and the wall’s U-factor is 0.369 W/(m2 ·K).
Complex Assemblies Building enclosure geometry of two- and three-dimensional assemblies may be complex, including corners, terminations of materials, and junctures of different materials. Such assemblies cannot be analyzed effectively with explicit calculations; rather, they require iterative calculations using computers. These calculations have been made for a number of common assemblies (Hershfield 2011), and the results can be applied within a simpler estimation method, as shown in Example 6. Buildings with thermal mass or other forms of thermal storage require dynamic models to evaluate their performance in the environmental climate of interest.
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Heat, Air , and Moisture Control in Building Assemblies—Examples
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Figure 7 shows a corner composed of homogeneous material. Surface temperatures can be estimated from the intersections of isotherms and the surface. If, in this figure, the interior were warm with respect to outside, then the line at the corner would be colder than the remainder of the interior surface. This effect may be exacerbated by the air film at the corner, which has a greater effective thickness than on the plane of the wall, and therefore offers greater thermal resistance, further lowering the corner temperature. Figure 8 shows an insulating material applied to a conductive material. Insulation is placed at the inside, during a period of cold outdoor temperatures. A computer program may be used to trace the isotherms. The interior isotherm is cut where the insulating material is interrupted, indicating lowered temperature at that location (point A). In fact, the temperature at the edge of the interrupted insulation is even lower than that at the surface of the uninsulated wall. Interruptions in insulation can lead to thermal bridges. For this reason, insulating conductive assemblies such as masonry or concrete is often more successful when applied to the outside rather than to the inside of the building. Example 6. Comparing Linear Transmittance Method to Area-Weighted Method for Brick Veneer Shelf Angle Anomaly. Calculate the overall U-value of the steel-stud brick veneer assembly with a slab and shelf angle with exterior insulation R-15 (2.6 RSI) and interior stud cavity insulation R-12 (2.1 RSI) (see Figure 9), using the following information and method: Gross wall height = 2.7 m Gross wall length = 15.2 m
Gross wall area
27.7 = 41.0 m
Area-Weighted Method. The thermal transmittance Ub for the area around the shelf angle is 1.162 W/(m2·K) for the effective lengths L1 = 205 mm, L2 = 619 mm, and Uo = 0.287 W/(m2·K). [Area thermal transmittance depends on the definition of the effective lengths (Hershfield 2011). This calculation is not shown here.] Calculate area for the thermal anomaly and area for the clear field: Ab =0.619 m × 15.2 m = 9.4 m2
Ao = 41.0 – 9.4 = 31.6 m2
Calculate the overall U-value: Ub Ab + Uo Ao 1.162 9.4 + 0.287 31.6 U = -------------------------------------- = -------------------------------------------------------------------- = 0.49 W/m2·K) A total 41 Linear Transmittance Method. From Appendix F of RP-1365, for this brick veneer assembly, Uo = 0.287 W/(m2 · K) and the linear transmittance of a slab with a shelf angle for this assembly is 0.544 W/(m · K). Calculate the overall U-value L L 15.2 ------------- U = ------------- + Uo = 0.544 × ---------- + 0.29 = 0.49 W/(m2 ·K) A total A total 41
This example illustrates the simplicity of the linear transmittance method compared to the area-weighted method for thermal anomalies in opaque building envelope assemblies. The amount of information that must be provided is reduced and the calculation is simpler. The weighted average method is further complicated for a whole-building elevation when accounting for 3D intersections. Some [e.g., Kemp (1997)] suggest that the overlapping effects be combined using mitered corners.
Windows and Doors
Fig. 8 Insulating Material Installed on Conductive Material, Showing Temperature Anomaly (Point A) at Insulation Edge
Table 4 of Chapter 15 lists U-factors for various fenestration products. For heat transmission coefficients for wood and steel doors, see Table 6 in Chapter 15. All U-factors are approximate, because a significant portion of the resistance of a window or door is contained in the air film resistances, and some parameters that may have important effects are not considered. For example, the listed U-factors assume the surface temperatures of surrounding bodies are equal to the ambient air temperature. However, the indoor surface of a window or door in an actual installation may be exposed to nearby radiating surfaces, such as radiant heating panels, or opposite walls with much higher or lower temperatures than the indoor air. Air movement across the surface of a window or door, such as that caused by nearby heating and cooling outlet grilles or by wind outdoors, increases the U-factor.
2. MOISTURE TRANSPORT The following examples build on the previous sections by discussing methods that combine heat and moisture transport analysis. The theory of hygrothermal analysis is described in Chapter 25. The methods include fundamental calculations that can be performed by hand as well as more advanced transient calculations that require computer modeling. A few simplified examples are presented here to aid in understanding these multimode, dynamic transport cases. These examples are simplified so that they can be explicitly calculated by assuming steady-state conditions, thus neglecting all storage phenomena. For other examples of explicit methods, see the Bibliography.
2.1
Fig. 9 Brick Veneer Shelf for Example 6
WALL WITH INSULATED SHEATHING
For an initial assessment of the impact of including any materials with low water vapor permeability in a building assembly, estimate the condensation resistance using a simplified moisture analysis. This simplified analysis assumes that interior surfaces may have little or no resistance to air or vapor flow, assumes steady-state conditions, and neglects the effect of radiative heat transfer on the
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2017 ASHRAE Handbook—Fundamentals (SI)
exposed surfaces. The section on Surface Condensation in Chapter 25 describes how to determine the risk of condensation on lowpermeability surfaces. Example 7. For the assembly shown in Figure 3 (but assume the wall has a thicker layer of mineral wool insulation, per following table), determine the range of indoor relative humidity for which condensation does not occur on the inside of the insulating sheathing. Assume a design outdoor temperature of –1°C, and indoor temperature of 21°C. Assume that the cavity air is at the same vapor pressure as the indoor air, which can occur with openings through the wallboard. Ignore radiant effects on the wall exterior. Assume the rigid insulation is vapor impermeable, and interior materials have little resistance to air or vapor flow. Air Film or Material
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1. 2. 3. 4. 5. 6. 7.
Indoor air film coefficient Gypsum wallboard Mineral fiber insulation 25 mm extruded polystyrene OSB sheathing, 13 mm Vinyl siding (hollow backed) Exterior air film coefficient Assembly R-value
Thermal Resistance, (m2 ·K)/W 0.12 0.08 3.34 0.88 0.12 0.107 0.03 4.68
Solution: The temperature difference is 22 K. The sum of the R-values from the foam/mineral fiber interface inward is 3.54 (m2 ·K)/W. The sum of the R-values from that interface outward is 1.14 (m2 ·K)/W. The temperature difference ratio [Equation (14), Chapter 25] is 1.14/4.68, or –0.24. The interface temperature is 4.3°C. The saturation vapor pressure of indoor air is 2476 Pa, and the saturation vapor pressure at the interface is 831 Pa (see Chapter 1). The upper bound for indoor relative humidity is therefore 831/2476, or 34% rh. (Another way to approach this problem would be to assume a given relative humidity within the indoor space and determine the range of outdoor temperatures for which condensation would not occur at this location.)
Many factors influence the likelihood (or not) of damage in an assembly such as this with exterior rigid insulation. Solar effects generally ensure a period of high temperatures that allows drying, so a design based on this result alone may be overly conservative. On the other hand, cold sky temperatures may increase heat loss from the assembly, which lowers the surface temperature below ambient, so a design based on this result alone may be subject to moisture damage. Even when indoor humidity is high enough that condensation at the interface is indicated, the rate of water formation may be slowed by airtightness at the wallboard and by vapor diffusion protection. In light of these dynamic effects, the analyst must consider the number of hours at which condensation could occur without damaging the assembly. The varying nature of the boundary conditions, as well as the storage capability of the assembly materials, lead to the need for more comprehensive dynamic evaluations (ASHRAE Standard 160).
2.2 VAPOR PRESSURE PROFILE (GLASER OR DEW-POINT) ANALYSIS The historical steady-state one-dimension tool for evaluating moisture accumulation and drying within exterior envelopes (walls, roofs, and ceilings) is the dew-point or Glaser method. With the increasing prominence of transient modeling tools, much more accurate estimates of temperature and humidity can be achieved than were possible using steady-state analysis. Users should recognize the limitations of the steady-state approach, which include the following: • Condensation is a phase change from vapor to liquid. As long as relative humidity in the pores of hygroscopic, capillary-porous building materials stays below 100%, vapor does not condense but is adsorbed as hygroscopic moisture or absorbed as liquid.
Only once moisture content at the surface touches the capillary maximum will vapor condense on that surface. The dew-point method results have often been interpreted to indicate condensation, when, in fact, increases in moisture content were through sorption or absorption, not visible condensation at the surface. • Heat and moisture storage effects are not included in a dew-point analysis. Experience shows they play a significant role in heat and moisture performance of assemblies. To account for storage, it is recommended to use average values (e.g., monthly average temperature) rather than more extreme design temperatures. • Diffusion is the only moisture transport mechanism considered. Airflow, capillary transport, rain wetting, initial conditions, latent effects, solar effects, and ventilation cannot or only approximately can be included in the method. They may have a dominant effect on building assembly performance. • The dew-point method allows calculation of a rate of moisture accumulation or rate of drying from a critical location within the assembly. However, the method does not allow estimating damage associated with any rate of accumulation or drying. The method is presented here for reasons of historical continuity, and because it serves as an illustration of the principles of heat conduction and vapor diffusion. However, the dew-point method is not recommended as a sole basis for hygrothermal design of building envelope assemblies. ASHRAE Standard 160 is recommended to assist in hygrothermal analysis for design purposes.
Winter Wall Wetting Examples Example 8. For a wood-framed wall, assume monthly mean conditions of 21°C, 40% rh indoors and –6.6°C, 50% rh outdoors. Indoor and outdoor vapor pressures are 995 and 175 Pa, respectively. Solution: Step 1. List the components in the building assembly, with their R-values and permeances.
Air Film or Material
Vapor ProporPropor- Vapor Diffusion tional tional PerThermal Temper- meance, Resistance, Vapor Resistance, ature (Pa·s·m2)/ Pressure ng/ (m2 ·K)/W Drop (s·m2 ·Pa) ng Drop
1 Surface film coefficient 2 Gypsum board, painted, cracked joints 3 Insulation, mineral fiber 4 Plywood sheathing 5 Wood siding 6 Surface film coefficient
0.12
0.050
9200
0.000109
0.003
0.08
0.033
290
0.00345
0.088
1.90
0.785
1700.0
0.00059
0.015
0.11
0.045
29
0.03448
0.881
0.18 0.03
0.074 0.012
2010 57000
0.00050 0.00002
0.013 0.000
Total
2.42
1.000
0.03914
1.000
Step 2. List the indoor and outdoor temperature and relative humidity. Vapor pressure at indoor and outdoor locations is determined by multiplying the saturation vapor pressure at that temperature by the relative humidity. Step 3. Calculate the proportional temperature drop across each layer. The temperature drop is proportional to the R-value: t layer R layer ----------------- = -------------ti – to RT The table in step 1 lists the resulting proportional temperature drops. Calculate the proportional water vapor pressure drops across each layer. These are calculated the same way as the proportional temperature drops in step 1:
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Heat, Air , and Moisture Control in Building Assemblies—Examples p layer Z layer ------------------ = -------------pi – po ZT where ZT = total water vapor diffusion resistance of wall (sum of diffusion resistances of all layers), (Pa·s·m2)/ng p = partial water vapor pressure, Pa Boundary or Interface Between Materials Indoor air 5-6 interface 4-5 interface 3-4 interface 2-3 interface 1-2 interface Outdoor air
Saturation Corrected Initial Vapor Vapor Vapor Rel. Tempera Pressure, Hum., Pressure, Pressure, Pa ture, °C % Pa Pa 21 19.6 18.7 –2.9 –4.2 –6.3 –6.6
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Difference 27.6
2488 2286 2161 478 430 361 350
40
50
Difference
995 993 921 908 186 175 175
995 760 756 478 430 183 175
820
Step 4. Determine the temperature at each interface, using the temperature difference from indoors to outdoors, and the proportional temperature drop. Find the saturation water vapor pressure corresponding to the interface temperatures from step 1. These values can be found in Table 2 in Chapter 1. Step 5. From step 1, the total water vapor diffusion resistance of the wall without the vapor retarder is Zwall = 1/9200 + 1/290 + 1/1700 + 1/29 + 1/2010 + 1/57 000 = 0.04 (Pa·s·m2)/ng The partial water vapor pressure drop across the whole wall is calculated from the indoor and outdoor saturation water vapor pressures and relative humidities (see the table in step 3). pwall = pi – po = (40/100)2488 – (50/100)350 = 820 Pa Step 6. Figure 10 shows the calculated saturation and partial water vapor pressures. Comparison reveals that the calculated partial water vapor pressure on the interior surface of the sheathing is well above saturation. This indicates incipient accumulation of water (condensation or sorption), probably on the surface of the sheathing, not within the insulation. If the accumulation rate is of interest, two additional steps are necessary. Step 7. The calculated water vapor pressure exceeds the saturation water vapor pressure by the greatest amount at the back side of the sheathing (Figure 12). Therefore, this is the most likely location for accumulation. Under conditions of phase change (condensation or sorption), the water vapor pressure should equal the saturation water vapor
Fig. 10 Dew-Point Calculation in Wood-Framed Wall (Example 8)
27.9
pressure at that interface (see the corrected vapor pressure column in step 3). Step 8. The change of water vapor pressure on the OSB sheathing alters all other partial water vapor pressures as well as water vapor flux through the wall. Calculating partial water vapor pressures is similar to the calculation in step 3, but the wall is now divided in two parts: one on the interior of the condensation interface (i.e., gypsum board and insulation) and the other on the exterior (OSB sheathing and wood siding). Water vapor pressure drop over the first (interior) part of the wall is p1 = 995 – 478 = 517 Pa and over the second (exterior) part is p2 = 478 – 175 = 303 Pa The diffusion resistances of both parts of the wall are Z1 = 1/9200 + 1/290 + 1/1700 = 0.004 (Pa·s·m2)/ng Z2 = 1/29 + 1/2010 + 1/57 000 = 0.035 (Pa·s·m2)/ng The water vapor pressure drops across each material can be calculated from the part between the inside and sheathing p layer Z layer ---------------------------------- = ----------------------sheathing p i – psheathing Zi and the part between the sheathing and outside p layer Z layer ----------------------------------- = ----------------------o psheathing – p o Z sheathing Vapor Pressure Vapor Ztot , Difference, Flow, Pa ng/(s·m2) (Pa·s·m2)/ng Indoor air to critical interface Critical interface to outdoor air
0.004 0.035
517 303
124 746 8 661
As shown in Figure 10, final calculations of water vapor pressure no longer exceed saturation, which means that the condensation plane was chosen correctly. However, vapor flux is no longer the same throughout the wall. The flux from inside increases; to the outside it decreases. The difference between both is the rate of moisture accumulation by interstitial condensation at the back side of the sheathing: p i – psheathing psheathing – p o - – ---------------------------------mc = --------------------------------sheathing o Zi Z sheathing In this case mc = 116 000 ng/(s·m2). Assume the12 mm OSB sheathing (density of 500 kg/m3) begins with moisture content of 10%. The mass of dry OSB at that thickness is 6.5 kg/m2, so the mass of water is 0.65 kg. If these conditions persist for 30 days (720 h), then the amount of accumulated water is 0.30 kg. This raises the moisture content of the wood to 12.5%. It is evident that wetting by diffusion is very slow. The Glaser method should not be used to show simply that calculated vapor pressure at one location exceeds saturation vapor pressure at that location. If that condition is detected, then the rate of accumulation must be calculated and the results compared to the affected material’s estimated storage potential. Unfortunately, guidance on interpretation of accumulated water with the Glaser method is not available. Considerations of moisture storage potential in building materials can be addressed only with transient modeling, not with steady-state methods. Example 9. A wood-framed construction has a wet layer inside. The wall layers include a 0.2 mm polyethylene membrane between insulation and gypsum board, and an exterior insulation and finish system (EIFS) with 40 mm of expanded polystyrene as substrate and a spun-glass reinforced stucco finish. If the OSB sheathing became soaked because of rain infiltration at the windows, how long before the OSB reaches hygroscopic equilibrium after leaks are sealed? Solve for two monthly mean conditions: winter, with 21°C, 40% rh indoors and –6.6°C, 50%
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2017 ASHRAE Handbook—Fundamentals (SI)
rh outdoors; and summer, with 25°C, 70% rh indoors and 23°C, 70% rh outdoors. Solution: Step 1. List the components in the building assembly, with their Rvalues and permeances. ProporProporVapor Thermal tional tional Vapor Diffusion Vapor Resis- Temper- Permetance R, ature ance, ng/ Resistance, Pressure (m2 ·K)/W Drop (Pa·s·m2) (Pa·s·m2)/ng Drop
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Air Film or Material 1.Air film coefficient 2.Gypsum board, painted 3.Polyethylene foil 4.Insulation, mineral fiber 5.OSB sheathing 6.EPS 7.EIFS stucco lamina and finish 8.Air film coefficient Total
Step 4. Calculate the total vapor resistance on either side of the critical OSB layer. Between inside and sheathing, p layer Z layer ---------------------------------- = ----------------------sheathing p i – psheathing Zi Between sheathing and outside, p layer Z layer ----------------------------------- = ----------------------o psheathing – p o Z sheathing From the summer and winter condition tables, the diffusion resistances of both parts of the wall are
0.120
0.038
9200
0.0001
0.000
0.079
0.025
290
0.003
0.001
0.000
0.000
0.435
2.299
0.975
1.900
0.595
1700
0.001
0.000
0.055
0.017
29
0.034
0.015
1.000 0.010
0.313 0.003
72.5 185
0.014 0.005
0.006 0.002
p i – psheathing 2 - = 106 ng/ s· m = 106 ng/(s·m2) mshething, i = --------------------------------i Z sheathing
0.030
0.009
57 000
0.00002
0.000
3.190
1.000
2.36
1.000
psheathing – p o mshething, o = ---------------------------------- = 28 659 ng/(s·m2) sheathing Zo
Step 2. List the indoor and outdoor temperature and relative humidity. As in Example 8, indoor and outdoor vapor pressure is determined by multiplying the saturation vapor pressure at that temperature by the relative humidity. Winter conditions: Saturated Relative Temperature, Vapor Humidity, °C Pressure, Pa % Indoors 1 and 2 2 and 3 3 and 4 4 and 5 5 and 6 6 and 7 7 and 8 Outdoors
21 20.0 19.3 19.3 2.9 2.4 –6.3 –6.3 –6.6
Difference
27.6
2488 2334 2237 2237 751 726 361 358 350
40
50
Z1 = 0.0001 + 0.003 + 2.299 + 0.001 = 2.303 (Pa·s·m2)/ng Z2 = 0.014 + 0.005 + 0.00002 = 0.019 (Pa·s·m2)/ng Step 5. Calculate the vapor pressure difference on either side of the critical OSB layer (the last column in the summer and winter condition tables). From the vapor resistance on each side and the vapor pressure difference on each side, vapor flow in each direction can be calculated. Winter
Vapor Pressure Vapor Difference, Flow, Ztot , Pa ng/(s·m2) (Pa·s·m2)/ng Indoor air to critical interface Critical interface to outdoor air
Vapor Pressure, Pa
Saturated Vapor Pressure, Pa
Relative Humidity, %
Vapor Pressure, Pa
Indoors 1 and 2 2 and 3 3 and 4 4 and 5 5 and 6 6 and 7 7 and 8 Outdoors
25 24.9 24.9 24.9 23.7 23.7 23.0 23.0 23
3169 3155 3146 3146 2929 2923 2815 2814 2811
70
2219 2219 2220 2929 2929 2923 2237 1968 1967
Difference
2.0
70
Step 3. Indoor and outdoor vapor pressures are calculated from the given conditions of temperature and relative humidity. The vapor pressure at each side of the OSB sheathing is assigned the value of the saturation vapor pressure at that temperature (see bold values in the summer and winter condition tables).
106 28 659 28 553
p i – psheathing mshething, i = --------------------------------- = –309 ng/(s·m2) i Z sheathing psheathing – p o mshething, o = ---------------------------------- = 49 744 ng/(s·m2) sheathing Zo Vapor Ztot, Pressure Vapor (Pa·s·m 2)/ Difference, Flow, ng/ Pa ng (s·m2) Indoor air to critical interface Critical interface to outdoor air
Temperature, °C
244.4 550.7 Net drying, ng/s
Summer
995 995 995 751 751 726 330 176 175
Summer conditions:
2.303 0.019
2.303 0.019
–710.8 955.9 Net drying, ng/s
–309 49 744 50 052
Drying consequently amounts to 74 g/m2 per month in winter and 130 g/m2 per month in summer. OSB soaked with water can have excess moisture content of up to 4.8 kg/m2. Drying only by onedimensional diffusion would appear to take several years at this rate. Radiation, air movement, and two- and three-dimensional effects can change the rate of drying. Figures 11 and 12 show the calculated water vapor pressure in winter and summer. The OSB is at water vapor saturation pressure; saturation is not reached at any other interface.
3.
TRANSIENT HYGROTHERMAL MODELING
Fundamentals of hygrothermal modeling tools, including modeling criteria and method of reporting, are discussed in Chapter 25. Although this chapter does not provide a complete example, it introduces input data commonly required by these programs and discusses considerations for analyzing output when using these tools.
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Heat, Air , and Moisture Control in Building Assemblies—Examples
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Fig. 11
Fig. 12
Drying Wet Sheathing, Winter (Example 9)
Drying Wet Sheathing, Summer (Example 9)
For many applications and for design guide development, actual behavior of an assembly under transient climatic conditions may be simulated to account for short-term processes such as driving rain absorption, summer condensation, and phase changes. Computer simulations allow designers to model these conditions over time. It is important, however, to understand the model’s application limits. Applying one of these models requires at least the following information: exterior climate conditions, indoor temperature and humidity, and building assembly materials and sizes. Many programs include a material property database and exterior climate and indoor condition data, allowing simple modeling to be performed without customization. However, using generic material property and weather data may not accurately recreate actual target conditions. Features of a complete moisture analysis model include • Transient heat, air, and moisture transport formulation, incorporating the physics of – Airflow – Water vapor transport by advection (combination of water vapor diffusion and air-driven vapor flow) – Liquid transport by capillary action, gravity, and pressure differences – Heat flow by apparent conduction, convection, and radiation – Heat and moisture storage/capacity of materials – Condensation and evaporation processes with linked latent-tosensible heat transformation – Freezing and thawing processes with linked latent-to-sensible heat transformation and based on laws of conservation of heat, mass, and momentum
27.11
• Material properties as functions of moisture content, relative humidity, and temperature, such as – Density – Air properties: permeability and permeance – Thermal properties: specific heat capacity, apparent thermal conductivity, Nusselt numbers, and long-wave emittance (for cavities and air spaces) – Moisture properties: porosity, sorption curve, water retention curve, vapor permeability, water permeability, or liquid diffusivity • Boundary conditions (generally on an hourly basis) – Outside temperature and relative humidity – Incident short-wave solar and long-wave sky radiation (depending on inclination and orientation) – Wind speed, orientation, and pressures – Wind-driven rain at exterior surfaces (depending on location and aerodynamics) – Interior temperature, air pressure excess, relative humidity (or interior moisture sources and ventilation flows), and air stratification – Surface conditions – Heat transfer film coefficients (combined convection and radiation, separate for convection and radiation) – Mass transfer film coefficients – Short-wave absorptance of exterior surfaces – Long-wave emittance of exterior surfaces – Contact conditions between layers and materials. Interfaces may be bridgeable for vapor diffusion, airflow, and gravity or pressure liquid flow only. They may be ideally capillary, introduce additional capillary resistance, or behave as real contact. Not all these features are required for every analysis, and some applications may need additional information (e.g., moisture flow through unintentional cracks and intentional openings, rain penetration through veneer walls and exterior cladding). To model these phenomena accurately, experiments may be needed to define systems and subsystems in field situation, because only then are all exterior loads and influences captured. It is important to recognize that simulation results are based on input data. Therefore, the more accurate the input data used, the closer the results will match real-world conditions. Exterior weather conditions, material properties, and interior operating conditions all vary widely, so it is important to get the best data available on materials being modeled. For most users, this is a difficult task. Many product manufacturers do not provide the material property data needed for the simulations. Developing weather data for a particular site is also beyond the expertise of many. Combined heat, air, and moisture models also have limitations. Users should be aware which transport phenomena and types of boundary conditions are included and which are not. For instance, some models cannot handle air transport or rain wetting of the exterior. Even an apparently simple problem, such as predicting rain leakage through a brick veneer, is beyond existing tools’ capabilities. In such cases, simple qualitative schemes and field tests still are the way to proceed. In addition, results also tend to be very sensitive to the choice of indoor and outdoor conditions. Usually, exact conditions are not known. Outputs from these programs typically include the moisture content of materials as well as relative humidity within the assembly. Interpretation of results is not easy: accurate data on moisture and temperature conditions that materials can tolerate are often not available. Although moisture accumulation may result in indoor air quality issues or material degradation, the effect of moisture accumulation in building assemblies depends on many factors, including choice of construction materials, and varies by building, making interpretation of results even more difficult.
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27.12
2017 ASHRAE Handbook—Fundamentals (SI)
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4.
AIR MOVEMENT
Moisture movement through building envelope assemblies is more strongly affected by air movement than by diffusion. To minimize moisture penetration by air leakages, the building envelope should be as airtight as possible. The airflow retarder must also be sufficiently strong and well supported to resist wind loads. In older residential buildings, air leakage provided sufficient ventilation and rarely led to interstitial condensation. However, in airtight buildings, mechanical ventilation is needed to ensure acceptable air quality and prevent moisture and health problems caused by excessive indoor humidity. Ventilation of the wall assembly and/or drainage must go to the outside of the airtight layer of construction, or it will increase building air leakage. To avoid moisture problems at the airtight layer, either the layer temperature must be kept above the dew point by locating it on the warm side of the insulation, or the layer permeance must allow vapor transmission. As described in detail in the section on Leakage Distribution in Chapter 16, air leakage through building envelopes is not confined to doors and windows. Although 6 to 22% of air leakage occurs there, 18 to 50% typically takes place through walls, and 3 to 30% through the ceiling. Leakage often occurs between sill plate and foundation, through interior walls, electrical outlets, plumbing penetrations, and cracks at top and bottom of exterior walls. Not all cracks and openings can be sealed in existing buildings, nor can absolutely tight construction be achieved in new buildings. Provide as tight an enclosure as possible to reduce leakage, minimize potential condensation within the envelope, and reduce energy loss. However, the project team must also recognize the impact of airtightness. Moisture accumulation in building envelopes can also be minimized by controlling the dominant direction of airflow by operating the building at a small negative or positive air pressure, depending on climate. In cooling climates, pressure should be positive to keep out humid outside air. In heating climates, pressure should be neither strongly negative, which could risk drawing soil gas or combustion products indoors and affect occupant comfort, nor strongly positive, which could risk driving moisture into building envelope cavities.
Equivalent Permeance Dew-point analysis allows simple estimation of the effect of wall and roof cavity ventilation on heat and vapor transport by using parallel thermal and vapor diffusion resistances (TenWolde and Carll 1992; Trethowen 1979). These parallel resistances account for heat and vapor that bypass exterior material layers with ventilation air from outside. Equivalent thermal and water vapor diffusion resistances are approximated from the following equations: 1 Rpar = --------ma c 1 Zpar = ---------ma where Rpar = parallel equivalent thermal resistance, (m2 ·K)/W Zpar = parallel equivalent water vapor diffusion vapor flow resistance, (Pa·s·m2)/ng ma = ventilation flux, kg/(m2 ·s) = 0.62/Pa, where Pa is atmospheric pressure, Pa
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore.
ASHRAE. 2009. Criteria for moisture-control design analysis in buildings. ANSI/ASHRAE Standard 160-2009. ASHRAE. 2010. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2010. Barbour, E., J. Goodrow, J. Kosny, and J.E. Christian. 1994. Thermal performance of steel-framed walls. Prepared for American Iron and Steel Institute by NAHB Research Center. Dill, R.S., W.C. Robinson, and H.E. Robinson. 1945. Measurements of heat losses from slab floors. National Bureau of Standards. Building Materials and Structures Report BMS 103. Farouk, B., and D.C. Larson. 1983. Thermal performance of insulated wall systems with metal studs. Proceedings of the 18th Intersociety Energy Conversion Engineering Conference, Orlando, FL. Hershfield, M. 2011. Thermal performance of building envelope details for mid- and high-rise buildings. Final Report, ASHRAE Research Project RP-1365. Hougten, F.C., S.I. Taimuty, C. Gutberlet, and C.J. Brown. 1942. Heat loss through basement walls and floors. ASHVE Transactions 48:369. Kemp, S. 1997. Modeling two- and three-dimensional heat transfer through composite wall and roof assemblies in transient energy simulation programs. Final Report, ASHRAE Research Project RP-1145. Kosny, J., and J.E. Christian. 1995. Reducing the uncertainties associated with using the ASHRAE zone method for R-value calculations of metal frame walls. ASHRAE Transactions 101(2):779-788. Labs, K., J. Carmody, R. Sterling, L. Shen, Y.J. Huang, and D. Parker. 1988. Building foundation design handbook. Report ORNL/SUB/86-72143/1. Oak Ridge National Laboratory, Oak Ridge, TN. Latta, J.K., and G.G. Boileau. 1969. Heat losses from house basements. Canadian Building 19(10). McGowan, A., and A.O. Desjarlais. 1997. An investigation of common thermal bridges in walls. ASHRAE Transactions 103(1):509-517. McIntyre, D.A. 1984. The increase in U-value of a wall caused by mortar joints. ECRC/M1843. The Electricity Council Research Centre, Copenhurst, U.K. Mitalas, G.P. 1982. Basement heat loss studies at DBR/NRC. NRCC 20416. Division of Building Research, National Research Council of Canada, September. Mitalas, G.P. 1983. Calculation of basement heat loss. ASHRAE Transactions 89(1B):420. Shipp, P.H. 1983. Basement, crawlspace and slab-on-grade thermal performance. Proceedings of the ASHRAE/DOE Conference, Thermal Performance of the Exterior Envelopes of Buildings II, ASHRAE SP 38: 160-179. Shu, L.S., A.E. Fiorato, and J.W. Howanski. 1979. Heat transmission coefficients of concrete block walls with core insulation. Proceedings of the ASHRAE/DOE-ORNL Conference, Thermal Performance of the Exterior Envelopes of Buildings, ASHRAE SP 28:421-435. TenWolde A. and C. Carll. 1992. Effect of cavity ventilation on moisture in walls and roofs. Proceedings of ASHRAE Conference, Thermal Performance of the Exterior Envelopes of Buildings V, pp. 555-562. THERM. 2012. THERM. windows.lbl.gov/software/therm/therm.html. Trethowen, H.A. 1979. The Kieper method for building moisture design. BRANZ Reprint 12, Building Research Association of New Zealand. Tye, R.P., and S.C. Spinney. 1980. A study of various factors affecting the thermal performance of perlite insulated masonry construction. Dynatech Report PII-2. Holometrix, Inc. (formerly Dynatech R/D Company), Cambridge, MA. Valore, R.C. 1980. Calculation of U-values of hollow concrete masonry. American Concrete Institute, Concrete International 2(2):40-62. Valore, R.C. 1988. Thermophysical properties of masonry and its constituents, parts I and II. International Masonry Institute, Washington, D.C. Van Geem, M.G. 1985. Thermal transmittance of concrete block walls with core insulation. ASHRAE Transactions 91(2). Wilkes, K.E. 1991. Thermal model of attic systems with radiant barriers. Report ORNL/CON-262. Oak Ridge National Laboratory, TN.
BIBLIOGRAPHY Hens, H. 1978. Condensation in concrete flat roofs. Building Research and Practice Sept./Oct.:292-309. TenWolde, A. 1994. Design tools. Chapter 11 in Moisture control in buildings. ASTM Manual MNL 18. American Society for Testing and Materials, West Conshohocken, PA. Vos, B.H., and E.J.W. Coelman. 1967. Condensation in structures. Report B67-33/23, TNO-IBBC, Rijswijk, the Netherlands.
Related Commercial Resources
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Related Commercial Resources CHAPTER 28
COMBUSTION AND FUELS Principles of Combustion......................................................... Fuel Classification ................................................................... Gaseous Fuels.......................................................................... Liquid Fuels .............................................................................
1.
28.1 28.5 28.5 28.8
PRINCIPLES OF COMBUSTION
OMBUSTION is a chemical reaction in which an oxidant reacts rapidly with a fuel to liberate stored energy as thermal energy, generally in the form of high-temperature gases. Small amounts of electromagnetic energy (light), electric energy (free ions and electrons), and mechanical energy (noise) are also produced during combustion. Except in special applications, the oxidant for combustion is oxygen in the air. The oxidation normally occurs with the fuel in vapor form. One notable exception is oxidation of solid carbon, which occurs directly with the solid phase. Conventional fuels contain primarily hydrogen and carbon, in elemental form or in various compounds (hydrocarbons). Their complete combustion produces mainly carbon dioxide (CO2) and water (H2O); however, small quantities of carbon monoxide (CO) and partially reacted flue gas constituents (gases and liquid or solid aerosols) may form. Most conventional fuels also contain small amounts of sulfur, which is oxidized to sulfur dioxide (SO2) or sulfur trioxide (SO3) during combustion, and noncombustible substances such as mineral matter (ash), water, and inert gases. Flue gas is the product of complete or incomplete combustion and includes excess air (if present), but not dilution air (air added to flue gas downstream of the combustion process, such as through the relief opening of a draft hood). Fuel combustion rate depends on the (1) rate of chemical reaction of combustible fuel constituents with oxygen, (2) rate at which oxygen is supplied to the fuel (mixing of air and fuel), and (3) temperature in the combustion region. The reaction rate is fixed by fuel selection. Increasing the mixing rate or temperature increases the combustion rate. With complete combustion of hydrocarbon fuels, all hydrogen and carbon in the fuel are oxidized to H2O and CO2. Generally, complete combustion requires excess oxygen or excess air beyond the amount theoretically required to oxidize the fuel. Excess air is usually expressed as a percentage of the air required to completely oxidize the fuel. In stoichiometric combustion of a hydrocarbon fuel, fuel is reacted with the exact amount of oxygen required to oxidize all carbon, hydrogen, and sulfur in the fuel to CO2, H2O, and SO2. Therefore, exhaust gas from stoichiometric combustion theoretically contains no incompletely oxidized fuel constituents and no unreacted oxygen (i.e., no carbon monoxide and no excess air or oxygen). The percentage of CO2 contained in products of stoichiometric combustion is the maximum attainable and is referred to as the stoichiometric CO2, ultimate CO2, or maximum theoretical percentage of CO2. Stoichiometric combustion is seldom realized in practice because of imperfect mixing and finite reaction rates. For economy and safety, most combustion equipment should operate with some excess air. This ensures that fuel is not wasted and that combustion is complete despite variations in fuel properties and supply rates of fuel and air. The amount of excess air to be supplied to any combustion
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C
The preparation of this chapter is assigned to TC 6.10, Fuels and Combustion.
28.1 Copyright © 2017, ASHRAE
Solid Fuels.............................................................................. Combustion Calculations ....................................................... Efficiency Calculations .......................................................... Combustion Considerations ...................................................
28.10 28.11 28.15 28.16
equipment depends on (1) expected variations in fuel properties and in fuel and air supply rates, (2) equipment application, (3) degree of operator supervision required or available, and (4) control requirements. For maximum efficiency, combustion at low excess air is desirable. Incomplete combustion occurs when a fuel element is not completely oxidized during combustion. For example, a hydrocarbon may not completely oxidize to carbon dioxide and water, but may form partially oxidized compounds, such as carbon monoxide, aldehydes, and ketones. Conditions that promote incomplete combustion include (1) insufficient air and fuel mixing (causing local fuel-rich and fuel-lean zones), (2) insufficient air supply to the flame (providing less than the required amount of oxygen), (3) insufficient reactant residence time in the flame (preventing completion of combustion reactions), (4) flame impingement on a cold surface (quenching combustion reactions), or (5) flame temperature that is too low (slowing combustion reactions). Incomplete combustion uses fuel inefficiently, can be hazardous because of carbon monoxide production, and contributes to air pollution.
Combustion Reactions The reaction of oxygen with combustible elements and compounds in fuels occurs according to fixed chemical principles, including • Chemical reaction equations • Law of matter conservation: the mass of each element in the reaction products must equal the mass of that element in the reactants • Law of combining masses: chemical compounds are formed by elements combining in fixed mass relationships • Chemical reaction rates Oxygen for combustion is normally obtained from air, which is a mixture of nitrogen, oxygen, small amounts of water vapor, carbon dioxide, and inert gases. For practical combustion calculations, dry air consists of 20.95% oxygen and 79.05% inert gases (nitrogen, argon, etc.) by volume, or 23.15% oxygen and 76.85% inert gases by mass. For calculation purposes, nitrogen is assumed to pass through the combustion process unchanged (although small quantities of nitrogen oxides form). Table 1 lists oxygen and air requirements for stoichiometric combustion and the products of stoichiometric combustion of some pure combustible materials (or constituents) found in common fuels.
Flammability Limits Fuel burns in a self-sustained reaction only when the volume percentages of fuel and air in a mixture at standard temperature and pressure are within the upper and lower flammability limits (UFL and LFL), also called explosive limits (UEL and LEL; see Table 2). Both temperature and pressure affect these limits. As mixture temperature increases, the upper limit increases and the lower limit decreases. As the pressure of the mixture decreases below atmospheric pressure, the upper limit decreases and the lower limit increases. However, as pressure increases above atmospheric, the upper limit increases and the lower limit is relatively constant.
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28.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 1
Combustion Reactions of Common Fuel Constituents Stoichiometric Oxygen and Air Requirements kg/kg Fuela
Molecular Formula
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Constituent Carbon (to CO) Carbon (to CO2) Carbon monoxide Hydrogen Methane Ethane Propane Butane Alkanes
C C CO H2 CH4 C2H6 C3H8 C4H10 CnH2n+2
Ethylene Propylene Alkenes Acetylene Alkynes
C2H4 C3H6 CnH2n C2H2 CnH2m
Sulfur (to SO2) S Sulfur (to SO3) S Hydrogen sulfide H2S
m3/m3 Fuel
Combustion Reactions
O2
Air
O2
Air
C + 0.5O2 CO C + O2CO2 CO + 0.5O2 CO2 H2 + 0.5O2H2O CH4 + 2O2 CO2 + 2H2O C2H6 + 3.5O2 2CO2 + 3H2O C3H8 + 5O2 3CO2 + 4H2O C4H10 + 6.5O2 4CO2 + 5H2O CnH2n + 2 + (1.5n + 0.5)O2 nCO2 + (n + 1)H2O C2H4 + 3O2 2CO2 + 2H2O C3H6 + 4.5O2 3CO2 + 3H2O CnH2n + 1.5nO2 nCO2 + nH2O C2H2 + 2.5O2 2CO2 + H2O CnH2m + (n + 0.5m)O2 nCO2 + mH2O
1.33 2.66 0.57 7.94 3.99 3.72 3.63 3.58 —
5.75 11.51 2.47 34.28 17.24 16.09 15.68 15.47 —
b
b
3.42 3.42 3.42 3.07 —
14.78 14.78 14.78 13.27 —
S + O2 SO2 S + 1.5O2 SO3 H2S + 1.5O2 SO2 + H2O
1.00 1.50 1.41
4.31 6.47 6.08
Substance Carbon Carbon monoxide Hydrogen Methane Ethane Propane n-Butane Ethylene Propylene Acetylene Sulfur Hydrogen sulfide
b
2.39 2.39 9.57 16.75 23.95 31.14 7.18n + 2.39
3.00 14.38 4.50 21.53 1.50n 7.18n 2.50 11.96 n+ 4.78n 0.5m + 2.39m b
b
b
b
1.50
7.18
kg/kg Fuel CO2
H2O
— 29.30 34.70 — 11.73 13.18 13.75 14.05 —
— — — 72 59 57 55 54 53
— — 1.0 — 1.0 2.0 3.0 4.0 n
— — — 1.0 2.0 3.0 4.0 5.0 n+1
— — 3.664 — 1.571 — — 8.937 2.744 2.246 2.927 1.798 2.994 1.634 3.029 1.550 44.01n 18.01(n + 1) 14.026n + 2.016 14.026n + 2.016 3.138 1.285 3.138 1.285 3.138 1.285 3.834 0.692 22.005n 9.008m 6.005n + 1.008m 6.005n + 1.008m
15.05 15.05 15.05 17.53 —
52 52 52 39 —
2.0 3.0 n 2.0 n
2.0 3.0 n 1.0 m
SOx
H2O
SOx
H2O
— — —
— — 52
1.0SO2 1.0SO3 1.0SO2
— — 1.0
1.998 (SO2) 2.497 (SO3) 1.880 (SO2)
— — 0.528
bVolume ratios are not given for fuels that do not exist in vapor form at reasonable temperatures or pressure.
Adapted, in part, from Gas Engineers Handbook (1965). aAtomic masses: H = 1.008, C = 12.01, O = 16.00, S = 32.06.
Table 2
b
0.50 0.50 2.00 3.50 5.00 6.50 1.5n + 0.5
Flue Gas from Stoichiometric Combustion with Air m3/m3 Fuel Ultimate Dew CO2, Point,c CO2 H2O % °C
Flammability Limits and Ignition Temperatures of Common Fuels in Fuel/Air Mixtures Molecular Lower Flammability Upper Flammability Formula Limit, % Limit, % C CO H2 CH4 C2H6 C3H8 C4H10 C2H4 C3H6 C2H2 S H2S
— 12.5 4.0 5.0 3.0 2.1 1.86 2.75 2.00 2.50 — 4.3
— 74 75.0 15.0 12.5 10.1 8.41 28.6 11.1 81 — 45.50
Ignition Temperature, °C 660 609 520 705 520 to 630 466 405 490 450 406 to 440 190 292
References Hartman (1958) Scott et al. (1948) Zabetakis (1956) Gas Engineers Handbook (1965) Trinks (1947) NFPA (1962) NFPA (1962) Scott et al. (1948) Scott et al. (1948) Trinks (1947) Hartman (1958) Scott et al. (1948)
Flammability limits adapted from Coward and Jones (1952). All values corrected to 15.6°C, 104 kPa, dry.
Ignition Temperature Ignition temperature is the lowest temperature at which heat is generated by combustion faster than it is lost to the surroundings and combustion becomes self propagating (see Table 2). The fuel/ air mixture will not burn freely and continuously below the ignition temperature unless heat is supplied, but chemical reaction between the fuel and air may occur. Ignition temperature is affected by a large number of factors. The ignition temperature and flammability limits of a fuel/air mixture, together, are a measure of the potential for ignition (Gas Engineers Handbook 1965).
Combustion Modes Combustion reactions occur in either continuous or pulse flame modes. Continuous combustion burns fuel in a sustained manner as long as fuel and air are continuously fed to the combustion zone and the fuel/air mixture is within the flammability limits. Continuous combustion is more common than pulse combustion and is used in most fuel-burning equipment.
Pulse combustion is an acoustically resonant process that burns various fuels in small, discrete fuel/air mixture volumes in a very rapid series of combustions. The introduction of fuel and air into the pulse combustor is controlled by mechanical or aerodynamic valves. Typical combustors consist of one or more valves, a combustion chamber, an exit pipe, and a control system (ignition means, fuel-metering devices, etc.). Typically, combustors for warm-air furnaces, hot-water boilers, and commercial cooking equipment use mechanical valves. Aerodynamic valves are usually used in higher-pressure applications, such as thrust engines. Separate valves for air and fuel, a single valve for premixed air and fuel, or multiple valves of either type can be used. Premix valve systems may require a flame trap at the combustion chamber entrance to prevent flashback. In a mechanically valved pulse combustor, air and fuel are forced into the combustion chamber through the valves under pressures less than 3.5 kPa. An ignition source, such as a spark, ignites the fuel/air mixture, causing a positive pressure build-up in the combustion
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Combustion and Fuels chamber. The positive pressure causes the valves to close, leaving only the exit pipe of the combustion chamber as a pressure relief opening. Combustion chamber and exit pipe geometry determine the resonant frequency of the combustor. The pressure wave from initial combustion travels down the exit pipe at sonic velocity. As this wave exits the combustion chamber, most of the flue gases present in the chamber are carried with it into the exit pipe. Flue gases remaining in the combustion chamber begin to cool immediately. Contraction of cooling gases and momentum of gases in the exit pipe create a vacuum inside the chamber that opens the valves and allows more fuel and air into the chamber. While the fresh charge of fuel/air enters the chamber, the pressure wave reaches the end of the exit pipe and is partially reflected from the open end of the pipe. The fresh fuel/air charge is ignited by residual combustion and/or heat. The resulting combustion starts another cycle. Typical pulse combustors operate at 30 to 100 cycles per second and emit resonant sound, which must be considered in their application. The pulses produce high convective heat transfer rates.
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Heating Value Combustion produces thermal energy (heat). The quantity of heat generated by complete combustion of a unit of specific fuel is constant and is called the heating value, heat of combustion, or caloric value of that fuel. A fuel’s heating value can be determined by measuring the heat evolved during combustion of a known quantity of the fuel in a calorimeter, or it can be estimated from quantitative chemical analysis of the fuel and the heating values of the various chemical elements in the fuel. For information on calculating heating values, see the sections on Characteristics of Fuel Oils and Characteristics of Coal. Higher heating value (HHV), gross heating value, or total heating value includes the latent heat of vaporization and is determined when water vapor in the fuel combustion products is cooled and condensed at standard temperature and pressure. Conversely, lower heating value (LHV) or net heating value does not include latent heat of vaporization. In the United States, when the heating value of a fuel is specified without designating higher or lower, it generally means the higher heating value. (LHV is mainly used for internal combustion engine fuels.) Heating values are usually expressed in kJ/L or MJ/m3 for gaseous fuels, MJ/L for liquid fuels, and MJ/kg for solid fuels. Heating values are always given in relation to standard temperature and pressure, usually 16, 20, or 25°C and 101.325 kPa, depending on the particular industry practice. Heating values in the United States and Canada are based on standard conditions of 15.6°C and 101.4 kPa, dry. Heating values of several substances in common fuels are listed in Table 3. With incomplete combustion, not all fuel is completely oxidized, and the heat produced is less than the heating value of the fuel. Therefore, the quantity of heat produced per unit of fuel consumed decreases (lower combustion efficiency). Not all heat produced during combustion can be used effectively. The greatest heat loss is the thermal energy of the increased temperature of hot exhaust gases above the temperature of incoming air and fuel. Other heat losses include radiation and convection heat transfer from the outer walls of combustion equipment to the environment.
Altitude Compensation Air at altitudes above sea level is less dense and has less mass of oxygen per unit volume. The volume concentration of oxygen, however, remains the same as sea level. Therefore, combustion at altitudes above sea level has less available oxygen to burn with the fuel unless compensation is made for the altitude. Combustion occurs, but the amount of excess air is reduced. If excess air is
28.3 Table 3 Heating Values of Substances Occurring in Common Fuels
Substance
Mole- Higher cular Heating For- Values,a mula MJ/m3
Higher Heating Values,a MJ/kg
Lower Heating Specific Values,a Density,b MJ/kg kg/m3
Carbon (to CO) Carbon (to CO2) Carbon monoxide Hydrogen Methane Ethane Propane Butane Ethylene Propylene Acetylene Sulfur (to SO2) Sulfur (to SO3) Hydrogen sulfide
C C CO H2 CH4 C2H6 C3H8 C4H10 C2H4 C3H6 C2H2 S S H2S
9.188 32.780 10.111 142.107 55.533 51.923 50.402 49.593 50.325 48.958 50.014 9.257 13.816 16.508
9.188 32.780 10.111 120.075 49.997 47.492 46.373 45.771 47.160 45.792 48.309 9.257 13.816 15.205
— — 12.0 12.1 37.7 66.1 94.0 128.9 59.8c 87.2 c 55.0 — — 24.1
— — 1.187 0.085 0.679 1.28 1.92 2.53 — 1.78 1.120 — — 1.456
Adapted from Gas Engineers Handbook (1965).
aAll values corrected to 15.6°C, 101.4 kPa, dry. For gases saturated with water vapor at
16°C, deduct 1.74% of value to adjust for gas volume displaced by water vapor. 0°C and 101.3 kPa. cNorth American Combustion Handbook (1986). bAt
reduced enough by an increase in altitude, combustion is incomplete or ceases. When gas-fired appliances operate at altitudes substantially above sea level, three notable effects occur (see Chapter 31 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment): • Oxygen available for combustion is reduced in proportion to the atmospheric pressure reduction. • With gaseous fuels, the heat of combustion per unit volume of fuel gas (gas heat content) is reduced because of reduced fuel gas density in proportion to the atmospheric pressure reduction. • Reduced air density affects the performance and operating temperature of heat exchangers and appliance cooling mechanisms. Altitude compensation matches fuel and air supply rates to attain complete combustion without too much excess air or too much fuel. This can be done at increased altitude by increasing the air supply amount to the combustion zone with a combustion air blower (air compensation), or by decreasing the fuel supply rate to the combustion zone by decreasing the fuel input (derating). Power burners use combustion air blowers and can increase the air supply rate to compensate for altitude. The combustion zone can be pressurized to attain the same air density in the combustion chamber as that at sea level. Derating can be used as an alternative to power combustion. U.S. fuel gas codes generally do not require derating of nonpower burners at altitudes up to 600 m. At altitudes above 600 m, many fuel gas codes require that burners be derated 4% for each 300 m above sea level (NFPA/AGA National Fuel Gas Code). Chimney or vent operation also must be considered at high altitudes (see Chapter 35 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment). In addition to reducing the gas heat content of fuel gases, reduced fuel gas density also causes increased gas velocity through flow metering orifices. The net effect is for gas input rate to decrease naturally with increases in altitude, but at less than the rate at which atmospheric oxygen decreases. This effect is one reason that derating is required when appliances are operated at altitudes significantly above sea level. Early research with draft hood-equipped appliances established that appliance input rates should be reduced at the rate of 4% per 300 m above sea level, for altitudes higher than 600 m above sea level (Figure 1).
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28.4
2017 ASHRAE Handbook—Fundamentals (SI) HHV = gas higher heating value at standard temperature and pressure of 15.6°C and 101.4 kPa, respectively, MJ/m3 B = local barometric pressure, kPa (not corrected to sea level: do not use barometric pressure as reported by weather forecasters, because it is corrected to sea level) Ps = standard pressure = 101.4 kPa
For example, at 1524 m, the barometric pressure is 84.316 kPa. If the HHV of a fuel gas sample is 37.5 MJ/m3 (at standard temperature and pressure), local gas heat content is 31.21 MJ/m3 at 84.316 kPa barometric pressure 1524 m above sea level. HC = 37.5 MJ/m3 × 84.316 kPa/101.4 kPa = 31.18 MJ/m3
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Fig. 1 Altitude Effects on Gas Combustion Appliances Experience with recently developed appliances with fan-assisted combustion systems demonstrates that the 4% rule may not be required in all cases. It is therefore important to consult the manufacturer’s listed appliance installation instructions, which are based on how the combustion system operates, and other factors (e.g., impaired heat transfer). ASHRAE research projects RP-1182 (Fleck et al. 2007) and RP-1388 (Suchovsky et al. 2011) concluded that the tested induced-draft combustion systems experienced a natural derate of 1.8% (the dotted curve in Figure 1) in gas input rate per 305 m increase in altitude above sea level. This gas input derate for altitude may provide safe combustion operation (less than 400 ppm of CO concentration in air-free flue gas) for induced-draft combustion systems where the combustion air drawn in by the inducer does not participate in gas entrainment or gas orifice outlet pressure that affects the gas rate. Additionally, the gas regulation reference pressure and the gas orifice outlet pressure must be at the same fluidic potential during operation. Alternatively, gas-fired systems where the gas orifice outlet pressure is affected by mechanical draft and not equal to the gas regulation pressure follow a different natural derate. Gas control systems such as the 100% premix systems using a zero-pressure (or near-zero) regulator gas delivery system and the Bernoulli principle for gas entrainment (often called constant-gas/air-ratio or tracking systems) have a natural derate of 3.1% per 305 m of elevation. These combustion systems can follow a 1:1 ratio with barometric pressure changes (the solid curve in Figure 1). These ASHRAE research projects showed that some gas-fired products as currently designed and constructed can be installed and operated safely and acceptably at some higher altitudes with no modifications to the sea-level gas orifices, gas manifold pressure, etc. It is important for appliance specifiers to be aware that the heating capacity of appliances is substantially reduced at altitudes significantly above sea level. To ensure adequate delivery of heat, derating of heating capacity must also be considered and quantified. By definition, fuel gas HHV value remains constant for all altitudes because (in North America) it is based on standard conditions of 101.4 kPa, dry, and 15.6°C. Some fuel gas suppliers at high altitudes (e.g., at Denver, Colorado, at 1524 m) may report fuel gas heat content at local barometric pressure instead of standard pressure. This can be calculated using the following equation: B HC = HHV ----Ps
(1)
Therefore, local gas heat content of a sample of fuel gas can be expressed as 31.18 MJ/m3 at local barometric pressure of 84.316 kPa and standard temperature, or as 37.5 MJ/m3 (HHV). Both gas heat contents are correct, but the application engineer must understand the difference to use each one correctly. As described earlier, local heat content HC can be used to determine appliance input rate. When gas heat value (either HHV or HC) is used to determine gas input rate, the gas pressure and temperature in the meter must also be considered. Add the gage pressure of gas in the meter to the local barometric pressure to calculate the heat content of the gas at the pressure in the meter. Gas temperature in the meter also affects the heat content of the gas in the meter. Gas heat value is directly proportional to gas pressure and inversely proportional to its absolute temperature in accordance with the perfect gas laws, as shown in the following example calculations for gas input rate with either the HHV or local heat content method. Example 1. Calculate the gas input rate for 37.5 MJ/m3 HHV fuel gas, 2.800 m3/h volumetric flow rate of 24°C fuel gas at 84.316 kPa barometer pressure (1525 m altitude) with 1.700 kPa fuel gas pressure in the gas meter. HHV Method: Q = 0.2778HHV × VFRs where Q = fuel gas input rate, kW 0.2778 = conversion factor, MJ/h to kW HHV = fuel gas higher heating value at standard temperature and pressure, MJ/m3 VFRs = fuel gas volumetric flow rate adjusted to standard temperature and pressure, m3/h = VFR(Ts × P)/(T × Ps) VFR = fuel gas volumetric flow rate at local temperature and pressure conditions, m3/h Ts = standard temperature, 288.75 K (15.6°C + 273.15) P = gas meter absolute pressure, kPa (local barometer pressure + gas pressure in meter relative to barometric pressure) = 84.316 kPa + 1.700 kPa = 86.016 kPa gas meter absolute pressure T = absolute temperature of fuel gas, K (fuel gas temperature in °C + 273.15 K) Ps = standard pressure, 101.4 kPa Substituting given values into the equation for VFRs gives 3
2.800 m /h 288.75 K 86.016 kPa VFRs = ------------------------------------------------------------------------------------------- = 2.308 m3/h 24.00C + 273.15 K 101.4 kPa Then, Q = 0.2778 37.5 MJ/m3 2.308 m3/h = 24.061 kW Local Gas Heat Content Method: The local gas heat content is simply the HHV adjusted to local gas meter pressure and temperature conditions. The gas input rate is simply the observed volumetric gas flow rate times the local gas heat content. Q = 0.2778HC × VFR
where HC = local gas heat content at local barometric pressure and standard temperature conditions, MJ/m3
where Q = fuel gas input rate, kW
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Combustion and Fuels 0.2778 = conversion factor, MJ/h to kW HC = fuel gas heat content at local gas meter pressure and temperature conditions, MJ/m3 VFR = fuel gas volumetric flow rate, referenced to local gas meter pressure and temperature conditions, m3/h HC = HHV(Ts × P)/(T × Ps ) Ts = standard temperature, 288.75 K (15.6°C + 273.15 K) P = gas meter absolute pressure, kPa (local barometer pressure + gas pressure in gas meter relative to barometric pressure) Ps = standard pressure = 101.4 kPa Pl = local barometric pressure = 84.316 kPa T = absolute temperature of fuel gas, 297.15 K (24.00°C fuel gas temperature + 273.15 K) Substituting given values into the equation for HC gives 37.5 288.75 84.316 + 1.700 HC = --------------------------------------------------------------------------- = 30.91 MJ/m3 297.15 101.4 Then, Q = 0.2778 × 30.91 MJ/m3 × 2.800 m3/h = 24.044 kW
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The gas input rate is about the same for both calculation methods.
2.
FUEL CLASSIFICATION
Generally, hydrocarbon fuels are classified according to physical state (gas, liquid, or solid). Different types of combustion equipment are usually needed to burn fuels in the different physical states. Gaseous fuels can be burned in premix or diffusion burners. Liquid fuel burners must include a means for atomizing or vaporizing fuel and must provide adequate mixing of fuel and air. Solid fuel combustion equipment must (1) heat fuel to vaporize sufficient volatiles to initiate and sustain combustion, (2) provide residence time to complete combustion, and (3) provide space for ash containment. Principal fuel applications include space heating and cooling of residential, commercial, industrial, and institutional buildings; service water heating; steam generation; and refrigeration. Major fuels for these applications are natural and liquefied petroleum gases (LPG), fuel oils, diesel and gas turbine fuels (for on-site energy applications), and coal. Fuels of limited use, such as manufactured gases, kerosene, liquid fuels derived from biological materials (wood, vegetable oils, and animal fat products), briquettes, wood, and coke, are not discussed here. Fuel choice is based on one or more of the following: Fuel factors • Availability, including dependability of supply • Convenience of use and storage • Economy • Cleanliness, including amount of contamination in unburned fuel [affecting (1) usability in fuel-burning equipment and (2) environmental impact] Combustion equipment factors • Operating requirements • Cost • Service requirements • Ease of control
3.
GASEOUS FUELS
Although various gaseous fuels have been used as energy sources in the past, heating and cooling applications are presently limited to natural gas and liquefied petroleum gases.
Types and Properties Natural Gas. This is a nearly odorless, colorless gas that accumulates in the upper parts of oil and gas reservoirs. Raw natural gas
28.5 is a mixture of methane (55 to 98%), higher hydrocarbons (primarily ethane), and noncombustible gases. Some constituents, principally water vapor, hydrogen sulfide, helium, liquefied petroleum gases, and gasoline, are removed before distribution. Natural gas used as fuel typically contains methane, CH4 (70 to 96%); ethane, C2H6 (1 to 14%); propane, C3H8 (0 to 4%); butane, C4H10 (0 to 2%); pentane, C5H12 (0 to 0.5%); hexane, C6H14 (0 to 2%); carbon dioxide, CO2 (0 to 2%); oxygen, O2 (0 to 1.2%); and nitrogen, N2 (0.4 to 17%). The composition of natural gas depends on its geographical source. Because the gas is drawn from various sources, the composition of gas distributed in a given location can vary slightly, but a fairly constant heating value is usually maintained for control and safety. Local gas utilities are the best sources of current gas composition data for a particular area. Heating values of natural gases vary from 34 to 45 MJ/m3; the usual range is 37.3 to 39.1 MJ/m3 at sea level. The heating value for a particular gas can be calculated from the composition data and values in Table 3. For safety purposes, odorants (e.g., mercaptans) are added to natural gas and LPG to give them noticeable odors. Liquefied Petroleum Gases (LPG). These gases consist primarily of propane and butane, and are usually obtained as a by-product of oil refinery operations or by stripping liquefied petroleum gases from the natural gas stream. Propane and butane are gaseous under usual atmospheric conditions, but can be liquefied under moderate pressures at normal temperatures. Commercial propane consists primarily of propane but generally contains about 5 to 10% propylene. Its heating value is about 50.15 MJ/kg, about 93 MJ/m3 of gas, or about 25.4 GJ/m3 of liquid propane. At atmospheric pressure, commercial propane has a boiling point of about –42°C. The low boiling point of propane allows it to be used during winter in the northern United States and southern Canada. Tank heaters and vaporizers allow its use also in colder climates and where high fuel flow rates are required. American Society for Testing and Materials (ASTM) Standard D1835 and Gas Processors Association (GPA) Standard 2140, which are similar, provide formulating specifications for required properties of liquefied petroleum gases at the time of delivery. Propane is shipped in cargo tank vehicles, rail cars, and barges. It is stored at consumer sites in tanks that comply with requirements of the ASME Boiler and Pressure Vessel Code or transportable cylinders that comply with requirements of the U.S. Department of Transportation. HD-5 propane is a special LPG product for use in internal combustion engines under moderate to high severity. Its specifications are included in ASTM Standard D1835 and GPA Standard 2140. Propane/air mixtures are used in place of natural gas in small communities and by natural gas companies to supplement normal supplies at peak loads. Table 4 lists heating values and densities for various fuel/air ratios. Commercial butane consists primarily of butane but may contain up to 5% butylene. It has a heating value of about 49.3 MJ/kg, about 120 MJ/m3 of gas, or about 28.4 GJ/m3 of liquid butane. At atmospheric pressure, commercial butane has a relatively high boiling point of about 0°C. Therefore, butane cannot be used in cold weather unless the gas temperature is maintained above 0°C or the partial pressure is decreased by dilution with a gas having a lower boiling point. Butane is usually available in bottles, tank trucks, or tank cars, but not in cylinders. Butane/air mixtures are used in place of natural gas in small communities and by natural gas companies to supplement normal supplies at peak loads. Table 4 lists heating values and densities for various fuel/air ratios. Commercial propane/butane mixtures with various ratios of propane and butane are available. Their properties generally fall between those of the unmixed fuels.
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28.6
2017 ASHRAE Handbook—Fundamentals (SI)
Table 4 Propane/Air and Butane/Air Gas Mixtures Heating Propane/Aira Butane/Airb Value, MJ/m3 % Gas % Air Density, kg/m3 % Gas % Air Density, kg/m3 18 22 26 30 34 38 42 46 50 54 58 62 66
19.16 23.41 27.67 31.93 36.18 40.44 44.70 48.95 53.21 57.47 61.72 65.98 70.24
80.84 76.59 72.33 68.07 63.82 59.56 55.30 51.05 46.79 42.53 38.28 34.02 29.76
1.41 1.44 1.46 1.49 1.52 1.54 1.57 1.60 1.63 1.65 1.68 1.71 1.73
14.81 18.11 21.40 24.69 27.98 31.27 34.57 37.86 41.15 44.44 47.74 51.03 54.32
85.19 81.89 78.60 75.31 72.02 68.73 65.43 62.14 58.85 55.56 52.26 48.97 45.68
1.48 1.52 1.56 1.60 1.64 1.68 1.72 1.76 1.80 1.84 1.88 1.92 1.96
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Adapted from Gas Engineers Handbook (1965). Air density at 0°C and 101.325 kPa is 1.292 kg/m3. aValues used for calculation: 93.97 MJ/m3; density = 1.92 kg/m3. bValues used for calculation: 121.5 MJ/m3; density = 2.53 kg/m3.
Manufactured gases are combustible gases produced from coal, coke, oil, liquefied petroleum gases, or natural gas. For more detailed information, see the Gas Engineers Handbook (1965). These fuels are used primarily for industrial in-plant operations or as specialty fuels (e.g., acetylene for welding). Renewable Gases. There are two primary processes that produce renewable gas from various feedstocks: anaerobic digestion and thermal gasification. Both processes are discussed here in general. Anaerobic Digestion (AD). In this process, complex organic matter (source material) is broken down into simpler constituents, directly through microbial action and in the absence of oxygen. The degradation process usually occurs in some form of tank, called a digester or reactor. Organic matter, perhaps first pretreated by grinding or by mechanical or chemical hydrolysis, enters the tank and is held there for a predefined length of time. For systems based on animal manure, this time ranges from a few days to a few weeks; for systems that use energy crops, residence time can be up to several tens of days. During that period, microbial activity breaks down the organic matter, and the resultant gaseous products contain a large fraction of methane and carbon dioxide along with trace amounts of other gases. Eventually, the material is expelled from the digester and replaced by new feed matter to continue the digestion/ degradation process. The new organic matter may replace the entirety of the resident matter in batch, or it may replace it semicontinuously, depending on the reactor and on the collection and processing of the source matter. The four stages of anaerobic digestion are as follows: 1. In hydrolysis, bacteria liquefy and break down organic matter comprised of complex organic polymers and cell structures. The end products are organic molecules that consist primarily of sugars, amino acids, peptides, and fatty acids. 2. In acidogenesis, acid-forming bacteria break down the products of the hydrolytic stage, forming volatile organic acids, CO2, hydrogen, and ammonia. 3. Next, in acetogenesis, bacteria convert the volatile organic acids from acidogenesis into acetic acid (CH3COOH) and acetate, CO2, and hydrogen. 4. Finally, methanogenesis uses methane-producing bacteria to change CO2 and acetic acid (products of the acidogenic and acetogenic stages) into methane (CH4). The resultant gas yield consists primarily of CH4, CO2, and other trace gases such as hydrogen sulfide (H2S).
Wastewater treatment plant (WWTP) gases are generated from waste liquids and solids from household or commercial water usage or from industrial processes. Depending on the architecture of the sewer system and local regulation, it may also contain stormwater from roofs, streets, or other runoff areas. The contents may include anything expelled (legally or not) from a household that enters the drains. If stormwater is included in the wastewater sewer flow, it may also contain components collected during runoff (e.g., soil, metals, organic compounds, animal waste, oils, solid debris such as leaves and branches). Processing influent to a large wastewater treatment plant typically has three stages: mechanical, biological, and sometimes chemical processing. The goal of such treatments is to prepare solids (treated sludge) and liquids (treated effluent) output from the WWTP that is environmentally safe and able to be landfilled (treated solids) or returned to the environment (treated effluent). One step in the processing of the wastewater sludge may be anaerobic digestion, from which methane can be produced. Landfill gas (LFG) derives from municipal solid waste (MSW) in landfills. In the United States, the primary federal law currently controlling disposal of solid and hazardous waste is the Resource Conservation and Recovery Act (RCRA), which sets criteria under which landfills can accept municipal solid waste and nonhazardous industrial solid waste. It also prohibits open dumping of waste, and ensures that hazardous waste is managed from the time of its creation to the time of its disposal. For energy production, MSW can be used in either of two ways: it can be gasified directly through thermochemical processes [see the section on Thermal Gasification (TG)], or it can be deposited into a landfill and undergo AD. Although most surveys indicate the material composition of all landfill constituents, the important components of MSW for the production of landfill gas are the organic fractions, which form the substrates that anaerobically decompose in the landfill. Typical or approximate contents of the organic fractions of MSW appear in Table 5. Landfill gas is the result of these anaerobic processes acting on the organic matter within a landfill. In a sense, the landfill itself substitutes for an anaerobic digester tank: a closed volume that contains putrescible matter and over time becomes devoid of oxygen. Methane and carbon dioxide are the principal components of the gas, though the overall composition of raw LFG can vary depending on the materials in the landfill: it can contain significant amounts of hydrogen sulfide as well as trace amounts of ammonia, mercury, chlorine, fluorine, siloxanes, and volatile metallic compounds. Variation in source MSW, its organic contents, temperature conditions, moisture conditions, compaction densities, landfill operational procedures, and other landfill attributes account for variations in LFG content. Typical compounds and their reported concentration ranges are shown in Table 6. Methane concentration is generally reported as being around 55 mol%, and carbon dioxide is often measured at 40%. Nitrogen, hydrogen, oxygen, and hydrogen sulfide are found in smaller but significant quantities. The composition of raw biogas can vary depending on the materials being digested. Methane and carbon dioxide are the principal components of landfill gas, though the overall composition of raw LFG can contain significant amounts of H2S as well as trace amounts of ammonia, mercury, chlorine, fluorine, siloxanes, and volatile metallic compounds (EPA 2017). However, the composition of biogas generated from dairy manure tends to be more consistent because the dairy industry is regulated as a producer of milk for human consumption. Typical compounds and their reported concentration ranges for digesterbased biogas are shown in Table 7. Methane concentration can be as high as 70% but is generally reported at around 60%. Landfill gas, unless the landfill is specifically designed for gas production, typically has a slightly lower methane fraction (e.g., around 55%).
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Combustion and Fuels Table 5
28.7
Components of Organic Portion of Municipal Solid Waste
Component Composition
Mass, %
Moisture Cellulose, sugar, starch Lipids Protein Other organics Inert materials
Compound
100
Table 8 Table 6
Landfill Gas Composition
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Compound Methane (CH4) Carbon dioxide (CO2) Nitrogen (N2) Hydrogen (H2) Carbon monoxide (CO) Oxygen (O2) Sulfides, disulfides, mercaptans, etc. Ammonia (NH3) Trace elements, amines, sulfur compounds, nonmethane volatile organic carbons halocarbons
Concentration
Methane (CH4) Carbon dioxide (CO2) Nitrogen (N2) Hydrogen (H2) Carbon monoxide (CO) Oxygen (O2) Hydrogen sulfide (H2S)
20.7 46.6 4.5 2.1 1.2 24.9
Total
Table 7 Typical Compounds and Concentrations in Biogas from Anaerobic Digester 54 to 70% 27 to 45% 0.5 to 3% 1 to 10% 0 to 0.1% 0 to 0.1% 600 to 7000+ ppm
Typical Compounds and Concentrations Found in Syngas from Thermal Gasification
mol% 45 to 60 40 to 60 2 to 5 0 to 0.2 0 to 0.2 0.1 to 1 0 to 1 0.1 to 1 0.01 to 0.6
Adding food wastes into a manure-based digester (codigestion) seems to improve biogas production and may increase methane concentration, but is not addressed in this chapter. CO2, the other major biogas component, is often measured around 40%. Nitrogen, hydrogen, oxygen, and H2S are found in smaller quantities. Similarly to natural gas, biogas derived from biomass feedstocks also must undergo one or more cleanup processes to remove unwanted components and to upgrade it suitably for natural gas pipelines. Quality control is required to prevent or minimize entry of raw, unconditioned biogas or less-than-pipeline-quality biomethane into the natural gas grid. Many methods and processes can be used to remove contaminants from subquality gas streams. Some are appropriate for use on farms, and others are only economical at gas flows measuring 30 000 m3 per day or more and where sulfur removal rates are measured in megagrams per day. The ability of a process to remove unwanted compounds depends on many factors, and assessment of the true practicality of a method for a given application requires careful evaluation. Thermal Gasification (TG). This process encompasses a fairly broad range of processes and reactions that convert carbonaceous feedstocks (coal, heavy oils, wood, biomass, sludge, etc.) into a mixture of gases, primarily hydrogen, carbon monoxide, steam, carbon dioxide, some methane, small amounts of ethane and higher hydrocarbons, small amounts of hydrogen sulfide, and nitrogen (if gasification is conducted with air). Depending on feedstock and operating conditions, TG of biomass typically generates tars and oils that are undesirable by-products. Thermal gasification is conducted in reducing (substoichiometric or incompletely combusting) atmospheres. Some of the process heat for the endothermic gasification reactions is typically provided by burning some of the carbon in the feedstock. Process heating can be direct or indirect. Indirect heating of the gasifier is called allthermal gasification. A typical range of syngas compositions from oxygen- or air-blown operation is presented in Table 8. This mixture of gases is known as synthesis gas or syngas, and can be further catalytically converted into methane to generate refinery gas (RG). The syngas can also be converted into liquid products by Fischer-Tropsch synthesis [see DOE (2017) for a description] for use as transportation fuel, or transformed into a host of chemical products such as methanol, dimethyl ether, fuel gas/
Air-Blown Steam-Blown Oxygen-Blown Fixed Bed Fluidized Bed Entrained Flow
Compound Typical Range Calorific value, MJ/m3 Hydrogen (H2), mol% Carbon monoxide (CO), mol% Carbon dioxide (CO2), mol% Methane (CH4), mol% Nitrogen (N2), mol%
4 to 6 11 to 16 13 to 18 12 to 16 2 to 6 45 to 60
12 to 14 35 to 45 22 to 25 20 to 23 9 to 11 1
10 to 12 23 to 28 45 to 55 10 to 15 1 5
town gas, ethylene/propylene, or acetic acid. It can also be combusted directly in a gas turbine to drive a generator. In some cases, a catalyst is included with the feedstock to accelerate reactions and allow a reduced operating temperature. TG can be carried out at temperatures in the range of 650 to 1100°C and at pressures ranging from ambient to greater than 7000 kPa. If the TG process is conducted at ambient or fairly low pressure, then the product RG must be compressed so it can be injected into the transmission or distribution line at the appropriate pressure.
4.
LIQUID FUELS
Significant liquid fuels include various fuel oils for firing combustion equipment and engine fuels for on-site energy systems. Liquid fuels, with few exceptions, are mixtures of hydrocarbons derived by refining crude petroleum. In addition to hydrocarbons, crude petroleum usually contains small quantities of sulfur, oxygen, nitrogen, vanadium, other trace metals, and impurities such as water and sediment. Refining produces a variety of fuels and other products. Nearly all lighter hydrocarbons are refined into fuels (e.g., liquefied petroleum gases, gasoline, kerosene, jet fuels, diesel fuels, light heating oils). Heavy hydrocarbons are refined into residual fuel oils and other products (e.g., lubricating oils, waxes, petroleum coke, asphalt). Crude petroleums from different oil fields vary in hydrocarbon molecular structure. Crude is paraffin-base (principally chainstructured paraffin hydrocarbons), naphthene- or asphaltic-base (containing relatively large quantities of saturated ring-structural naphthenes), aromatic-base (containing relatively large quantities of unsaturated, ring-structural aromatics, including multi-ring compounds such as asphaltenes), or mixed- or intermediate-base (between paraffin- and naphthene-base crudes). Except for heavy fuel oils, the crude type has little significant effect on resultant distillate products and combustion applications.
Types of Fuel Oils Fuel oils for heating are broadly classified as distillate fuel oils (lighter oils) or residual fuel oils (heavier oils). ASTM Standard D396 has specifications for fuel oil properties that subdivide the oils into various grades. Grades No. 1 and 2 are distillates; grades 4, 5 (Light), 5 (Heavy), and 6 are residual. Specifications for the grades are based on required characteristics of fuel oils for use in different types of burners.
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28.8 Grade No. 1 is a light distillate intended for vaporizing-type burners. High volatility is essential to continued evaporation with minimum residue. This fuel is also used in extremely cold climates for residential heating using pressure-atomizing burners. Grade No. 2 is heavier than No. 1 and is used primarily with pressure-atomizing (gun) burners that spray oil into a combustion chamber. Vapor from the atomized oil mixes with air and burns. This grade is used in most domestic burners and many medium-capacity commercial/industrial burners. A dewaxed No. 2 oil with a pour point of –50°C is supplied only to areas where regular No. 2 oil would jell. Grade No. 2—low sulfur is a relatively new category that has a sulfur content of 0.05%. Lower fuel sulfur content reduces fouling rates of boiler heat exchangers (Butcher et al. 1997). Grade No. 4 is an intermediate fuel that is considered either a heavy distillate or a light residual. Intended for burners that atomize oils of higher viscosity than domestic burners can handle, its permissible viscosity range allows it to be pumped and atomized at relatively low storage temperatures. Grade No. 5 (Light) is a residual fuel of intermediate viscosity for burners that handle fuel more viscous than No. 4 without preheating. Preheating may be necessary in some equipment for burning and, in colder climates, for handling. Grade No. 5 (Heavy) is a residual fuel more viscous than No. 5 (Light), but intended for similar purposes. Preheating is usually necessary for burning and, in colder climates, for handling. Grade No. 6, sometimes referred to as Bunker C, is a highviscosity oil used mostly in commercial and industrial heating. It requires preheating in the storage tank to allow pumping, and additional preheating at the burner to allow atomizing. Low-sulfur residual oils are marketed in many areas to allow users to meet sulfur dioxide emission regulations. These fuel oils are produced (1) by refinery processes that remove sulfur from the oil (hydrodesulfurization), (2) by blending high-sulfur residual oils with low-sulfur distillate oils, or (3) by a combination of these methods. These oils have significantly different characteristics from regular residual oils. For example, the viscosity/temperature relationship can be such that low-sulfur fuel oils have viscosities of No. 6 fuel oils when cold, and of No. 4 when heated. Therefore, normal guidelines for fuel handling and burning can be altered when using these fuels. Another liquid fuel of increasing interest is biodiesel. It is made from biological sources (e.g., vegetable oils, used cooking oils, tallow). ASTM Standard D6751 addresses biodiesel; requirements are largely similar to those for petroleum diesel (cetane number, flash point, etc.; see the section on Types and Properties of Liquid Fuels for Engines). In practice, biodiesel is almost always blended, most often with ASTM heating oils when used for stationary heating applications, because of cost and cold-flow properties of 100% biodiesel. However, the benefits of a renewable fuel that has very low net carbon dioxide emission in its life cycle, reduced particulate and sulfur emissions, and lower NOx emissions in many heating applications balance the need for mixing. Fuel oil grade selection for a particular application is usually based on availability and economic factors, including fuel cost, clean air requirements, preheating and handling costs, and equipment cost. Installations with low firing rates and low annual fuel consumption cannot justify the cost of preheating and other methods that use residual fuel oils. Large installations with high annual fuel consumption cannot justify the premium cost of distillate fuel oils.
Characteristics of Fuel Oils Characteristics that determine grade classification and suitability for given applications are (1) viscosity, (2) flash point, (3) pour point, (4) water and sediment content, (5) carbon residue, (6) ash, (7) distillation qualities or distillation temperature ranges, (8) density, (9) sulfur content, (10) heating value, (11) carbon/ hydrogen
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 2 Approximate Viscosity of Fuel Oils content, (12) aromatic content, and (13) asphaltene content. Not all of these are included in ASTM Standard D396. Viscosity is an oil’s resistance to flow. It is significant because it indicates the ease with which oil flows or can be pumped and the ease of atomization. Differences in fuel oil viscosities are caused by variations in the concentrations of fuel oil constituents and different refining methods. Approximate viscosities of fuel oils are shown in Figure 2. Flash point is the lowest temperature to which an oil must be heated for its vapors to ignite in a flame. Minimum permissible flash point is usually prescribed by state and municipal laws. Pour point is the lowest temperature at which a fuel can be stored and handled. Fuels with higher pour points can be used when heated storage and piping facilities are provided. Water and sediment content should be low to prevent fouling the facilities. Sediment accumulates on filter screens and burner parts. Water in distillate fuels can cause tanks to corrode and emulsions to form in residual oil. Carbon residue is obtained by a test in which the oil sample is destructively distilled in the absence of air. When commercial fuels are used in proper burners, this residue has almost no relationship to soot deposits, except indirectly when deposits are formed by vaporizing burners. Ash is the noncombustible material in an oil. An excessive amount indicates the presence of materials that cause high wear on burner pumps. The distillation test shows the volatility and ease of vaporization of a fuel. Relative density is the ratio of the density of a fuel oil to the density of water at a specific temperature. Relative densities cover a range in each grade, with some overlap between distillate and residual grades. Air pollution considerations are important in determining the allowable sulfur content of fuel oils. Sulfur content is frequently limited by legislation aimed at reducing sulfur oxide emissions from combustion equipment; usual maximum allowable sulfur content levels are 1.0, 0.5, or 0.3%. Table 9 lists sulfur levels of some marketed fuel oils. Research (Lee et al. 2002a, 2002b) suggests that fuel
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Combustion and Fuels
28.9
Table 9 Sulfur Content of Marketed Fuel Oils Grade of Oil Total fuel samples
No. 5 No. 5 No. 4 (Light) (Heavy) No. 6
No. 1
No. 2
31
61
13
15
16
96
0.03 0.50 0.20
0.46 1.44 0.83
0.90 3.50 1.46
0.57 2.92 1.46
0.32 4.00 1.41
17 2 0 0
13 11 3 0
15 15 9 2
16 16 11 0
96 93 60 8
Sulfur content, % mass minimum 0.001 maximum 0.120 average 0.023 No. samples with S over 0.3% over 0.5% over 1.0% over 3.0%
Types and Properties of Liquid Fuels for Engines
0 0 0 0
Data for No. 1 and No. 2 oil derived from Dickson and Sturm (1994). Data for No. 4, 5, and 6 oil derived from Shelton (1974).
Table 10
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Grade No. 1 2 4 5L 5H 6
Typical Density and Higher Heating Value of Standard Grades of Fuel Oil Density, kg/m3 833 to 800 874 to 834 933 to 886 951 to 921 968 to 945 1012 to 965
Higher Heating Value, GJ/m3 38.2 to 37.0 39.5 to 38.2 41.3 to 39.9 41.8 to 40.9 42.4 to 41.6 43.5 to 42.2
sulfur content affects the sulfate content of particulate emissions, which are reported to be associated with adverse health effects. Sulfur in fuel oils is also undesirable because sulfur compounds in flue gas are corrosive. Although low-temperature corrosion can be minimized by maintaining the stack at temperatures above the dew point of the flue gas, this limits the overall thermal efficiency of combustion equipment. The presence of sulfur oxides in the flue gas raises the dew point temperature (see the section on Combustion Calculations). For certain industrial applications (e.g., direct-fired metallurgy, where work is performed in the combustion zone), fuel sulfur content must be limited because of adverse effects on product quality. Sulfur contents of typical fuel oils are listed in Table 9. Heating value is an important property, although ASTM Standard D396 does not list it as one of the criteria for fuel oil classification. Table 10 shows the relationship between heating value and density for several oil grades. In the absence of more specific data, heating values can be calculated as derived from the North American Combustion Handbook (1978): Higher heating value, MJ/kg = 51.92 – 8.79 10–62
(2)
Distillate fuel oils (grades 1 and 2) have a carbon/hydrogen content of 84 to 86% carbon, with the remainder predominantly hydrogen. Heavier residual fuel oils (grades 4, 5, and 6) may contain up to 88% carbon and as little as 11% hydrogen. An approximate relationship for determining the hydrogen content of fuel oils is Hydrogen, % = 26 – (15 × Relative density)
(3)
ASTM Standard D396 is more a classification than a specification, distinguishing between six generally nonoverlapping grades, one of which characterizes any commercial fuel oil. Quality is not defined, as a refiner might control it; for example, the standard lists the distillation temperature 90% point for grade No. 2 as having a maximum of 338°C, whereas commercial practice rarely exceeds 315°C.
The primary stationary engine fuels are diesel and gas turbine oils, natural gases, and LPGs. Other fuels include sewage gas, manufactured gas, and other commercial gas mixtures. Gasoline and the JP series of gas turbine fuels are rarely used for stationary engines. Only properties of diesel and gas turbine fuel oils are covered here; properties of natural and liquefied petroleum gases are found in the section on Gaseous Fuels. For properties of gasolines and JP turbine fuel, consult texts on internal combustion engines and gas turbines. Properties of currently marketed gasolines can be found in ASTM Standard D4814. Properties of the three grades of diesel fuel oils (1-D, 2-D, and 4D) are listed in ASTM Standard D975. Grade No. 1-D includes the class of volatile fuel oils from kerosene to intermediate distillates. They are used in high-speed engines with frequent and relatively wide variations in loads and speeds and where abnormally low fuel temperatures are encountered. Grade No. 2-D includes the class of lower-volatility distillate gas oils. They are used in high-speed engines with relatively high loads and uniform speeds, or in engines not requiring fuels with the higher volatility or other properties specified for grade No. 1-D. Grade No. 4-D covers the more viscous distillates and blends of these distillates with residual fuel oils. They are used in low- and medium-speed engines involving sustained loads at essentially constant speed. Property specifications and test methods for grade No. 1-D, 2-D, and 4-D diesel fuel oils are essentially identical to specifications of grade No. 1, 2, and 4 fuel oils, respectively. However, diesel fuel oils have an additional specification for cetane number, which measures ignition quality and influences combustion roughness. Cetane number requirements depend on engine design, size, speed and load variations, and starting and atmospheric conditions. An increase in cetane number over values actually required does not improve engine performance. Thus, the cetane number should be as low as possible to ensure maximum fuel availability. ASTM Standard D975 provides several methods for estimating cetane number from other fuel oil properties. ASTM Standard D2880 for gas turbine fuel oils relates gas turbine fuel oil grades to fuel and diesel fuel oil grades. Test methods for determining properties of gas turbine fuel oils are essentially identical to those for fuel oils. However, gas turbine specifications limit quantities of some trace elements that may be present, to prevent excessive corrosion in gas turbine engines. For a detailed discussion of fuels for gas turbines and combustion in gas turbines, see Chapters 5 and 9, respectively, in Hazard (1971).
5.
SOLID FUELS
Solid fuels include coal, coke, wood, and waste products of industrial and agricultural operations. Of these, only coal is widely used for heating and cooling applications. Coal’s complex composition makes classification difficult. Chemically, coal consists of carbon, hydrogen, oxygen, nitrogen, sulfur, and a mineral residue, ash. Chemical analysis provides some indication of quality, but does not define its burning characteristics sufficiently. Coal users are principally interested in the available energy per unit mass of coal and the amount of ash and dust produced, but are also interested in burning characteristics and handling and storing properties. A description of coal qualities and characteristics from the U.S. Bureau of Mines as well as other information can be obtained from the U.S. Geological Survey at energy .er.usgs.gov/products/databases/USCoal/index.htm and the Energy Information Administration at www.eia.doe.gov/fuelcoal.html.
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2017 ASHRAE Handbook—Fundamentals (SI) Table 11 Classification of Coals by Ranka
Class I
Anthracite
Group 1. Metaanthracite 2. Anthracite 3. Semianthracite
II Bituminousd
1. Low-volatile bituminous coal 2. Medium-volatile bituminous coal 3. High-volatile Type A bituminous coal
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4. High-volatile Type B bituminous coal 5. High-volatile Type C bituminous coal
Limits of Fixed Carbon or Energy Content, Mineral-Matter-Free Basis
Requisite Physical Properties
Dry FC, 98% or more (Dry VM, 2% or less) Dry FC, 92% or more, and less than 98% (Dry VM, 8% or less, and more than 2%) Dry FC, 86% or more, and less than 92% (Dry VM, 14% or less, and more than 8%)
Nonagglomerating
Dry FC, 78% or more, and less than 86% (Dry VM, 22% or less, and more than 14%) Dry FC, 69% or more, and less than 78% (Dry VM, 31% or less, and more than 22%) Dry FC, less than 69% (Dry VM, more than 31%), and moist,c about 32.6 MJ/kge or more Moist,c about 30.2 MJ/kg or more, and less than 32.6 MJ/kge Moist,c about 25.6 MJ/kg or more, and less than 30.2 MJ/kge
Either agglomeratingb or nonweatheringf
III Subbituminous 1. Subbituminous Type A coal 2. Subbituminous Type B coal 3. Subbituminous Type C coal
Moist,c about 25.6 MJ/kg or more, and less than 30.2 MJ/kge Moist,c about 22.1 MJ/kg or more, and less than 25.6 MJ/kge Moist,c about 19.3 MJ/kg or more, and less than 22.1 MJ/kge
Both weathering and nonagglomeratingb
IV Lignitic
Moist,c less than 19.3 MJ/kg Moist,c less than 19.3 MJ/kg
Consolidated Unconsolidated
1. Lignite 2. Brown coal
Source: Data from ASTM Standard D388. FC = fixed carbon; VM = volatile matter; MMF = mineral-matter-free aClassification does not include a few coals of unusual physical and chemical properties that come within limits of fixed carbon or energy content of highvolatile bituminous and subbituminous ranks. All these coals either contain less than 48% dry, MMF FC, or have more than about 36.1 MJ/kg, which is moist, MMF.
b If
agglomerating, classify in group 1 of class II. refers to coal containing natural bed moisture but without visible water on coal surface. may be noncaking varieties in each group of class II. e Coals with 69% or more fixed carbon on dry, MMF basis are classified according to FC, regardless of energy content. f There are three varieties of coal in group 5: variety 1, agglomerating and nonweathering; variety 2, agglomerating and weathering; and variety 3, nonagglomerating and nonweathering.
cMoist
d There
Types of Coals Commonly accepted definitions for classifying coals are listed in Table 11. This classification is arbitrary because there are no distinct demarcation lines between coal types. Anthracite is a clean, dense, hard coal that creates little dust in handling. It is comparatively difficult to ignite, but burns freely once started. It is noncaking and burns uniformly and smokelessly with a short flame. Semianthracite has a higher volatile content than anthracite. It is not as hard and ignites more easily. Otherwise, its properties are similar to those of anthracite. Bituminous coal includes many types of coal with distinctly different compositions, properties, and burning characteristics. Coals range from high-grade bituminous, such as those found in the eastern United States, to low-rank coals, such as those found in the western United States. Caking properties range from coals that melt or become fully plastic, to those from which volatiles and tars are distilled without changing form (classed as noncaking or free-burning). Most bituminous coals are strong and nonfriable enough to allow screened sizes to be delivered free of fines. Generally, they ignite easily and burn freely. Flame length is long and varies with different coals. If improperly fired, much smoke and soot are possible, especially at low burning rates. Semibituminous coal is soft and friable, and handling creates fines and dust. It ignites slowly and burns with a medium-length flame. Its caking properties increase as volatile matter increases, but the coke formed is weak. With only half the volatile matter content of bituminous coals, burning produces less smoke; hence, it is sometimes called smokeless coal. Subbituminous coal, such as that found in the western United States, is high in moisture when mined and tends to break up as it dries or is exposed to the weather; it is likely to ignite spontaneously when piled or stored. It ignites easily and quickly, has a mediumlength flame, and is noncaking and free-burning. The lumps tend to break into small pieces if poked. Very little smoke and soot are formed.
Lignite is woody in structure, very high in moisture when mined, of low heating value, and clean to handle. It has a greater tendency than subbituminous coals to disintegrate as it dries and is also more likely to ignite spontaneously. Because of its high moisture, freshly mined lignite ignites slowly and is noncaking. The char left after moisture and volatile matter are driven off burns very easily, like charcoal. The lumps tend to break up in the fuel bed and pieces of char that fall into the ash pit continue to burn. Very little smoke or soot forms.
Characteristics of Coal The characteristics of coals that determine classification and suitability for given applications are the proportions of (1) volatile matter, (2) fixed carbon, (3) moisture, (4) sulfur, and (5) ash. Each of these is reported in the proximate analysis. Coal analyses can be reported on several bases: as-received, moisture-free (or dry), and mineral-matter-free (or ash-free). As-received is applicable for combustion calculations; moisture-free and mineral-matter-free, for classification purposes. Volatile matter is driven off as gas or vapor when the coal is heated according to a standard temperature test. It consists of a variety of organic gases, generally resulting from distillation and decomposition. Volatile products given off by heated coals differ materially in the ratios (by mass) of the gases to oils and tars. No heavy oils or tars are given off by anthracite, and very small quantities are given off by semianthracite. As volatile matter increases to as much as 40% of the coal (dry and ash-free basis), increasing amounts of oils and tars are released. However, for coals of higher volatile content, the quantity of oils and tars decreases and is relatively low in the subbituminous coals and in lignite. Fixed carbon is the combustible residue left after the volatile matter is driven off. It is not all carbon. Its form and hardness are an indication of fuel coking properties and, therefore, guide the choice of combustion equipment. Generally, fixed carbon represents that portion of fuel that must be burned in the solid state. Moisture is difficult to determine accurately because a sample can lose moisture on exposure to the atmosphere, particularly when reducing the sample size for analysis. To correct for this loss, total
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Table 12 Typical Ultimate Analyses for Coals
Rank
As Received, MJ/kg
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Anthracite 29.5 Semianthracite 31.6 Low-volatile 33.4 bituminous Medium-volatile 32.6 bituminous High-volatile bituminous Type A 32.1 B 29.1 C 25.6 Subbituminous Type B 20.9 C 19.8 Lignite 16.0
6.
Constituents, Percent by Mass O
H
C
N
5.0 5.0 5.0
2.9 3.9 4.7
80.0 80.4 81.7
0.9 1.1 1.4
0.7 10.5 1.1 8.5 1.2 6.0
S
Ash
5.0
5.0
81.4
1.4
1.5
6.0
9.3 13.8 20.6
5.3 5.5 5.8
75.9 67.8 59.6
1.5 1.4 1.1
1.5 3.0 3.5
6.5 8.5 9.4
29.5 35.7 44.0
6.2 6.5 6.9
52.5 46.4 40.1
1.0 0.8 0.7
1.0 1.0 1.0
9.8 9.6 7.3
moisture content of a sample is customarily determined by adding the moisture loss obtained when air-drying the sample to the measured moisture content of the dried sample. Moisture does not represent all of the water present in coal; water of decomposition (combined water) and of hydration are not given off under standardized test conditions. Ash is the noncombustible residue remaining after complete coal combustion. Generally, the mass of ash is slightly less than that of mineral matter before burning. Sulfur is an undesirable constituent in coal, because sulfur oxides formed when it burns contribute to air pollution and cause combustion system corrosion. Table 12 lists the sulfur content of typical coals. Legislation has limited the sulfur content of coals burned in certain locations. Heating value may be reported on an as-received, dry, dry and mineral-matter-free, or moist and mineral-matter-free basis. Higher heating values of coals are frequently reported with their proximate analysis. When more specific data are lacking, the higher heating value of higher-quality coals can be calculated by the Dulong formula: Higher heating value, MJ/kg = 33.829C + 144.28[H – (O/8)] + 9.42S
(4)
where C, H, O, and S are the mass fractions of carbon, hydrogen, oxygen, and sulfur in the coal obtained from the ultimate analysis. Other important parameters in judging coal suitability include • Ultimate analysis, which is another method of reporting coal composition. Percentages of C, H, O, N, S, and ash in the coal sample are reported. Ultimate analysis is used for detailed fuel studies and for computing a heat balance when required in heating device testing. Typical ultimate analyses of various coals are shown in Table 12. • Ash-fusion temperature, which indicates the fluidity of the ash at elevated temperatures. It is helpful in selecting coal to be burned in a particular furnace and in estimating the possibility of ash handling and slagging problems. • The grindability index, which indicates the ease with which a coal can be pulverized and is helpful in estimating ball mill capacity with various coals. There are two common methods for determining the index: Hardgrove (see Hardgrove Grindability Index at www.acarp.com.au/Media/ACARP-WP-5-Hardgrove GrindabilityIndex.pdf) and ball mill. • The free-swelling index, which denotes the extent of coal swelling on combustion on a fuel bed and indicates the coking characteristics of coal.
COMBUSTION CALCULATIONS
Calculations of the quantities of (1) air required for combustion and (2) flue gas products generated during combustion are frequently needed for sizing system components and as input to efficiency calculations. Other calculations, such as values for excess air and theoretical CO2, are useful in estimating combustion system performance. Frequently, combustion calculations can be simplified by using relative molecular mass. The relative molecular mass of a compound equals the sum of the atomic masses of the elements in the compound. Molecular mass can be expressed in any mass units. The gram molecular mass or gram mole is the molecular mass of the compound expressed in grams. The molecular mass of any substance contains the same number of molecules as the molecular mass of any other substance. Corresponding to measurement standards common to the industries, calculations involving gaseous fuels are generally based on volume, and those involving liquid and solid fuels generally use mass. Some calculations described here require data on concentrations of carbon dioxide, carbon monoxide, and oxygen in the flue gas. Gas analyses for CO2, CO, and O2 can be obtained by volumetric chemical analysis and other analytical techniques, including electromechanical cells used in portable electronic flue gas analyzers.
Air Required for Combustion Stoichiometric (or theoretical) air is the exact quantity of air required to provide oxygen for complete combustion. The three most prevalent components in hydrocarbon fuels (C, H2, and S) are completely burned as in the following fundamental reactions: C + O2 CO2 H2 + 0.5O2 H2O S + O2 SO2 In the reactions, C, H2, and S can be taken to represent 1 kg mole of carbon, hydrogen, and sulfur, respectively. Using approximate atomic masses (C = 12, H = 1, S = 32, and O = 16), 12 kg of C are oxidized by 32 kg of O2 to form 44 kg of CO2, 2 kg of H2 are oxidized by 16 kg of O2 to form 18 kg of H2O, and 32 kg of S are oxidized by 32 kg of O2 to form 64 kg of SO2. These relationships can be extended to include hydrocarbons. The mass of dry air required to supply a given quantity of oxygen is 4.32 times the mass of the oxygen. The mass of air required to oxidize the fuel constituents listed in Table 1 was calculated on this basis. Deduct oxygen contained in the fuel, except the amount in ash, from the amount of oxygen required, because this oxygen is already combined with fuel components. In addition, when calculating the mass of supply air for combustion, allow for water vapor, which is always present in atmospheric air. Combustion calculations for gaseous fuels are based on volume. Avogadro’s law states that, for any gas, one mole occupies the same volume at a given temperature and pressure. Therefore, in reactions involving gaseous compounds, the gases react in volume ratios identical to the gram mole ratios. That is, to oxidize hydrogen in the preceding reaction, one volume (or 1 kg mole) of hydrogen reacts with one-half volume (or 0.5 kg mole) of oxygen to form one volume (or 1 kg mole) of water vapor. The volume of air required to supply a given volume of oxygen is 4.78 times the volume of oxygen. The volumes of dry air required to oxidize the fuel constituents listed in Table 1 were calculated on this basis. Volume ratios are not given for fuels that do not exist in vapor form at reasonable temperatures or pressures. Again, oxygen contained in the fuel should be deducted from the quantity of oxygen required, because this oxygen is already combined with fuel
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components. Allow for water vapor, which increases the volume of dry air by 1 to 3%. From the relationships just described, the theoretical mass ma of dry air required for stoichiometric combustion of a unit mass of any hydrocarbon fuel is ma = 0.0144(8C + 24H + 3S – 3O)
(5)
where C, H, S, and O are the mass percentages of carbon, hydrogen, sulfur, and oxygen in the fuel. Analyses of gaseous fuels are generally based on hydrocarbon components rather than elemental content. If fuel analysis is based on mass, the theoretical mass ma of dry air required for stoichiometric combustion of a unit mass of gaseous fuel is ma = 2.47CO + 34.28H2 + 17.24CH4 + 16.09C2H6 + 15.68C3H8 + 15.47C4H10 + 13.27C2H2 + 14.78C2H4 + 6.08H2S – 4.32O2
(6)
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(7)
where CO, H2, and so forth are the volumetric fractions of each constituent in the fuel gas. Illuminants include a variety of compounds not separated by usual gas analysis. In addition to ethylene (C2H4) and acetylene (C2H2), the principal illuminants included in Equation (7), and the dry air required for combustion, per unit volume of each gas, are as follows: propylene (C3H6), 21.44; butylene (C4H8), 28.58; pentene (C5H10), 35.73; benzene (C6H6); 35.73, toluene (C7H8), 42.88; and xylene (C8H10), 50.02. Because toluene and xylene are normally scrubbed from the gas before distribution, they can be disregarded in computing air required for combustion of gaseous fuels. The percentage of illuminants present in gaseous fuels is small, so the values can be lumped together, and an approximate value of 30 unit volumes of dry air per unit volume of gas can be used. If ethylene and acetylene are included as illuminants, a value of 20 unit volumes of dry air per unit volume of gaseous illuminants can be used. For many combustion calculations, only approximate values of air requirements are necessary. If approximate values for theoretical air are sufficient, or if complete information on the fuel is not available, the values in Tables 13 and 14 can be used. Another value used for estimating air requirements is 0.24 m3 of air for 1 MJ of fuel. In addition to the amount theoretically required for combustion, excess air must be supplied to most practical combustion systems to ensure complete combustion: Air supplied – Theoretical air Excess air, % = ----------------------------------------------------------------------Theoretical air
Type of Fuel
Air Required kg/kg Fuel
Solid
MJ/kg 0.314 Liquid MJ/kg 0.305 Gas MJ/kg 0.288
Approx. Precision, m3/Unit Fuel* % MJ/kg 0.26 MJ/kg 0.35 MJ/m3 0.24
3
Exceptions Fuels containing more than 30% water Results low for gasoline and kerosene 11.2 MJ/m3 or less
3 5
Source: Data based on Shnidman (1954). *Unit fuel for solid and liquid fuels in kg, for gas in L.
Table 14 Theoretical Air Requirements for Stoichiometric Combustion of Various Fuels
If fuel analysis is reported on a volumetric or molecular basis, it is simplest to calculate air requirements based on volume and, if necessary, convert to mass. The theoretical volume Va of air required for stoichiometric combustion of a unit volume of gaseous fuels is Va = 2.39CO + 2.39H2 + 9.57CH4 + 16.75C2H6 + 23.95C3H8 + 31.14C4H10 + 11.96C2H2 + 14.38C2H4 + 7.18H2S – 4.78O2 + 30.47 illuminants
Table 13 Approximate Air Requirements for Stoichiometric Combustion of Fuels by Category
Type of Fuel
kg/kg fuel 9.6 11.2 10.3 6.2 11.2 Mg/m3 fuel 12.34 12.70 13.42 13.66 m3/m3 fuel 9.6 31.1 24.0
The amount of dry air supplied per unit mass of fuel burned can be obtained from the following equation, which is reasonably precise for most solid and liquid fuels: C 3.04N 2 Dry air supplied = -------------------------CO 2 + CO
(9)
where Dry air supplied = unit mass per unit mass of fuel C = unit mass of carbon burned per unit mass of fuel, corrected for carbon in ash CO2, CO, N2 = percentages by volume from flue gas analysis
These values of dry air supplied and theoretical air can be used in Equation (8) to determine excess air. Excess air can also be calculated from unit volumes of stoichiometric combustion products and air, and from volumetric analysis of the flue gas: P U – CO 2 Excess air, % = 100 --- --------------------- A CO 2
(8)
The excess air level at which a combustion process operates significantly affects its overall efficiency. Too much excess air dilutes flue gas excessively, lowering its heat transfer temperature and increasing sensible flue gas loss. Conversely, too little excess air can lead to incomplete combustion and loss of unburned combustible gases. Combustion efficiency is usually maximized when just enough excess air is supplied and properly mixed with combustible gases to ensure complete combustion. The general practice is to supply 5 to 50% excess air, depending on the type of fuel burned, combustion equipment, and other factors.
Theoretical Air Required for Combustion
Solid fuels Anthracite Semibituminous Bituminous Lignite Coke Liquid fuels No. 1 fuel oil No. 2 fuel oil No. 5 fuel oil No. 6 fuel oil Gaseous fuels Natural gas Butane Propane
(10)
where U = ultimate carbon dioxide of flue gases resulting from stoichiometric combustion, % CO 2 = carbon dioxide content of flue gases, % P = dry products from stoichiometric combustion, unit volume per unit volume of gas burned A = air required for stoichiometric combustion, unit volume per unit volume of gas burned
Because the ratio P/A is approximately 0.9 for most natural gases, a value of 90 can be substituted for 100(P/A) in Equation (10) for rough calculation.
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Table 15 Approximate Maximum Theoretical (Stoichiometric) CO2 Values, and CO2 Values of Various Fuels with Different Percentages of Excess Air
Type of Fuel Gaseous fuels Natural gas Propane gas (commercial) Butane gas (commercial) Mixed gas (natural and carbureted water gas) Carbureted water gas Coke oven gas Liquid fuels No. 1 and 2 fuel oil No. 6 fuel oil
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Solid fuels Bituminous coal Anthracite Coke
Percent CO2 at Given Excess Air Values
Theoretical or Maximum CO2, %
20%
40%
60%
12.1 13.9 14.1 11.2
9.9 11.4 11.6 12.5
8.4 9.6 9.8 10.5
7.3 8.4 8.5 9.1
17.2 11.2
14.2 9.2
12.1 7.8
10.6 6.8
15.0 16.5
12.3 13.6
10.5 11.6
9.1 10.1
18.2 20.2 21.0
15.1 16.8 17.5
12.9 14.4 15.0
11.3 12.6 13.0
Because excess air calculations are almost invariably made from flue gas analysis results and theoretical air requirements are not always known, another convenient method of expressing Equation (8) is 100 O 2 – CO 2 Excess air, % = ---------------------------------------------------------------0.264 N2 – O 2 – CO 2
Water Vapor and Dew Point of Flue Gas
11CO 2 + 8O 2 + 7 CO + N 2 Dry flue gas = ---------------------------------------------------------------------3 CO 2 + CO
(13)
where
(11)
Dry flue gas = kg/kg of fuel C = kg of carbon burned per kg of fuel, corrected for carbon in ash CO2, O2, CO, N2 = percentages by volume from flue gas analysis
(12)
The total dry gas volume of flue gases from combustion of one unit volume of gaseous fuels for various percentages of CO2 is
where O2, CO, and N2 are percentages by volume from the flue gas analysis, dry basis.
Volume of CO 2 produced 100 Dry flue gas = -------------------------------------------------------------- ---------- Unit vol. of gas burned CO 2
(14)
where
Theoretical CO2 The theoretical (or ultimate, stoichiometric, or maximum) CO2 concentration attainable in the products from the combustion of a hydrocarbon fuel with air is obtained when the fuel is completely burned with the theoretical quantity of air and zero excess air. Theoretical CO2 varies with the carbon/hydrogen ratio of the fuel. For combustion with excess air present, theoretical CO2 values can be calculated from the flue gas analysis: CO 2 Theoretical CO2, % = U = -------------------------------------1 – O 2 20.95
Fig. 3
Adapted from Gas Engineers Handbook (1965). Printed with permission of Industrial Press and American Gas Association.
(12)
where CO2 and O2 are percentages by volume from the flue gas analysis, dry basis. Table 15 gives approximate theoretical CO2 values for stoichiometric combustion of several common types of fuel, as well as CO2 values attained with different amounts of excess air. In practice, desirable CO2 values depend on the excess air, fuel, firing method, and other considerations.
Quantity of Flue Gas Produced The mass of dry flue gas produced per mass of fuel burned is required in heat loss and efficiency calculations. This mass is equal to the sum of the mass of (1) fuel (minus ash retained in the furnace), (2) air theoretically required for combustion, and (3) excess air. For solid fuels, this mass, determined from the flue gas analysis, is
Dry flue gas = unit volume per unit volume of gaseous fuel CO2 = percentage by volume from the flue gas analysis
Excess air quantity can be estimated by subtracting the quantity of dry flue gases resulting from stoichiometric combustion from the total volume of flue gas.
Water Vapor and Dew Point of Flue Gas Water vapor in flue gas is the total of the water (1) contained in the fuel, (2) contained in the stoichiometric and excess air, and (3) produced from combustion of hydrogen or hydrocarbons in the fuel. The amount of water vapor in stoichiometric combustion products may be calculated from the fuel burned by using the water data in Table 1. The dew point is the temperature at which condensation begins and can be determined using Figure 3. The volume fraction of water vapor Pwv in the flue gas can be determined as follows: Vw Pwv = ------------------------------------------ 100V c P c + V w
(15)
where Vw = total water vapor volume (from fuel; stoichiometric, excess, and dilution air; and combustion) Vc = unit volume of CO2 produced per unit volume of gaseous fuel Pc = percent CO2 in flue gas
Using Figure 4, the dew points of solid, liquid, or gaseous fuels may be estimated. For example, to find the dew point of flue gas resulting from the combustion of a solid fuel with a mass ratio
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Fig. 4
Theoretical Dew Points of Combustion Products of Industrial Fuels
Adapted from Gas Engineers Handbook (1965). Printed with permission of Industrial Press and American Gas Association.
(hydrogen to carbon-plus-sulfur) of 0.088 and sufficient excess air to produce 11.4% oxygen in the flue gas, start with the mass ratio of 0.088. Proceed vertically to the intersection of the solid fuels curve and then to the theoretical dew point of 46°C on the dewpoint scale (see dashed lines in Figure 4). Follow the curve fixed by this point (down and to the right) to 11.4% oxygen in the flue gas (on the abscissa). The actual dew point is 34°C and is found on the dew-point scale. The dew point can be estimated for flue gas from natural gas having a higher heating value (HHV) of 38 MJ/m3 with 6.3% oxygen or 31.5% air. Start with 38 MJ/m3 and proceed vertically to the intersection of the gaseous fuels curve and then to the theoretical dew point of 59°C on the dew-point scale. Follow the curve fixed by this point to 6.3% oxygen or 31.5% air in the flue gas. The actual dew point is 53°C. The presence of sulfur dioxide, and particularly sulfur trioxide, influences the vapor pressure of condensate in flue gas, and the dew point can be raised by as much as 14 to 42°C, as shown in Figure 5. For a manufactured gas with an HHV of 20.5 MJ/m3 containing 340 mg of sulfur per cubic metre being burned with 40% excess air, the proper curve in Figure 5 is determined as follows: Fig. 5 3
Mass sulfur in fuel, mg/m 340 ----------------------------------------------------------------3- = ---------- = 16.6 20.5 Fuel heating value, MJ/m
(16)
This curve lies between the 0 and 20 curves and is close to the 20 curve. The dew point for any percentage of excess air from zero to 100% can be determined on this curve. For this flue gas with 40%
Influence of Sulfur Oxides on Flue Gas Dew Point
excess air, the dew point is about 80°C, instead of 65°C for zero sulfur at 40% excess air.
Sample Combustion Calculations Example 2. Analysis of flue gases from burning a natural gas shows 10.0% CO2, 3.1% O2, and 86.9% N2 by volume. Analysis of the fuel is 90%
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Combustion and Fuels
28.15
CH4, 5% N2, and 5% C2H6 by volume. Find U (maximum theoretical percent CO2), and percentage of excess air. Solution: From Equation (12),
where mg (mass of dry flue gas per mass of fuel, kg/kg) is calculated as in Equation (13). 3. Heat loss in water vapor in products formed by combustion of hydrogen
10.0 U = -------------------------------------- = 11.74% CO2 1 – 3.1 20.95
q3 = (9H2/100)[(h)tg – (hf )ta] 4. Heat loss in water vapor in the combustion air
From Equation (10), using 100(P/A) = 90,
q4 = Mma[(h)tg – (hg )ta]
11.74 – 10.0 90 Excess air = ----------------------------------------- = 15.7% 10 Example 3. For the same analysis as in Example 2, find, per cubic metre of fuel gas, the volume of dry air required for combustion, the volume of each constituent in the flue gases, and the total volume of dry and wet flue gases. Solution: From Equation (7), the volume of dry air required for combustion is
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= 9.45
per
m3
CO q5 = 23 591C ------------------------- CO + CO 2
Oxygen, O2 In excess air Carbon dioxide, CO2 From methane From ethane
(0.9CH4)(9.57 2.0) = 6.81 (0.05C2H6)(16.75 3.5) = 0.66 = 0.05 0.791 0.157 9.45 = 1.17 Total nitrogen = 8.69 m3 0.209 0.157 9.45 = 0.31 m3
(0.9CH4)(1.0) = 0.90 (0.05C2H6)(2.0) = 0.10 Total carbon dioxide = 1.00 m3 Water vapor, H2O (does not appear in some flue gas analyses) From methane (0.9CH4)(2.0) = 1.8 From ethane (0.05C2H6)(3.0) = 0.15 Total water vapor = 1.95 m3 Total volume of dry gas per cubic metre of fuel gas 8.69 + 0.31 + 1.00 = 10.0 m3 Total volume of wet gases per cubic metre of fuel gas (neglecting water vapor in combustion air) 10.0 + 1.95 = 11.95 m3 The cubic metres of dry flue gas per cubic metre of fuel gas can also be computed from Equation (14): (1.00)(100)/10.0 = 10.0 m3
7.
(21)
7. Unaccounted-for heat losses, q7
(The volume of dry air may also be calculated using Table 14.) From Table 1, the cubic metres of flue gas constituents per cubic metre of fuel gas are as follows: Nitrogen, N2 From methane From ethane Nitrogen in fuel Nitrogen in excess air
(20)
6. Heat loss from unburned carbon in the ash or refuse q6 = 33 957[(Cu /100) – C]
of fuel gas
(19)
where ma is calculated as in Equations (5) and (6). 5. Heat loss from incomplete combustion of carbon
9.57CH4 + 16.75C2H6 = (9.57 0.90) + (16.75 0.05) m3
(18)
The following symbols are used in Equations (17) to (21): q1 = useful heat, kJ/kg of fuel q2 = heat loss in dry flue gases, kJ/kg of fuel q3 = heat loss in water vapor from combustion of hydrogen, kJ/kg of fuel q4 = heat loss in water vapor in combustion air, kJ/kg of fuel q5 = heat loss from incomplete combustion of carbon, kJ/kg of fuel q6 = heat loss from unburned carbon in ash, kJ/kg of fuel q7 = unaccounted-for heat losses, kJ/kg of fuel cpg = mean specific heat of flue gases at constant pressure [from 1.01 to 1.06 kJ/(kg·K) for flue gas temperatures from 150 to 540°C], kJ/(kg·K) (h)tg = enthalpy of superheated steam at flue gas temperature and 101.4 kPa, kJ/kg (hf )ta = enthalpy of saturated water liquid at air temperature, kJ/kg (hg)ta = enthalpy of saturated steam at combustion air temperature, kJ/kg ma = mass of combustion air per mass of fuel used, kg/kg of fuel mg = mass of dry flue gas per mass of fuel, kg/kg of fuel ta = temperature of combustion air, °C tg = temperature of flue gases at exit of heating device, °C H2 = hydrogen in fuel, % by mass (from ultimate analysis of fuel) M = humidity ratio of combustion air, mass of water vapor per mass of dry air CO, CO2 = carbon monoxide and carbon dioxide in flue gases, % by volume C = mass of carbon burned per unit of mass of fuel, corrected for carbon in ash, kg/kg of fuel
WC u – W a C a C = -------------------------------100W
(22)
EFFICIENCY CALCULATIONS
In analyzing heating appliance efficiency, an energy balance is made that accounts (as much as possible) for disposition of all thermal energy released by combustion of the fuel quantity consumed. The various components of this balance are generally expressed in terms of megajoules per kilogram of fuel burned or as a percentage of its higher heating value. The following are major components of an energy balance and their calculation methods: 1. Useful heat q1, or heat transferred to the heated medium; for convection heating equipment, this value is computed as the product of the mass rate of flow and enthalpy change. 2. Heat loss as sensible heat in the dry flue gases q2 = mg cpg (tg – ta)
(17)
where Cu = percentage of carbon in fuel by mass from ultimate analysis Wa = mass of ash and refuse Ca = percent of combustible in ash by mass (combustible in ash is usually considered to be carbon) W = mass of fuel used
Useful heat (item 1) is generally measured for a particular piece of combustion equipment. Flue gas loss is the sum of items 2 to 6. However, for cleanburning gas- and oil-fired equipment, items 5 and 6 are usually negligible and flue gas loss is the sum of items 2, 3, and 4. Flue gas losses (the sum of items 2, 3, and 4) can be determined with sufficient precision for most purposes from the curves in Figure 6, if O2 content and flue gas temperature are known. Values
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28.16
2017 ASHRAE Handbook—Fundamentals (SI) Generally, item 5 is negligible for modern combustion equipment in good operating condition. Item 6 is generally negligible for gas and oil firing, but should be determined for coal-firing applications.
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of the losses were computed from typical ultimate analyses, assuming 1% water vapor (by mass) in the combustion air. Curves for medium-volatile bituminous coal can be used for high-volatile bituminous coal with no appreciable error.
Fig. 6 Flue Gas Losses with Various Fuels (Flue gas temperature rise shown. Loss based on 18°C room temperature.)
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Combustion and Fuels
28.17
Item 7 consists primarily of radiation and convection losses from combustion equipment surfaces and losses caused by incomplete combustion not included in items 5 and 6. Heat loss from incomplete combustion is determined by subtracting the sum of items 1 to 6 from the fuel heating value. Radiation and convection losses are not usually determined by direct measurement, but if the heating appliance is located within the heated space, radiation and convection losses can be considered useful heat rather than lost heat and can be omitted from heat loss calculations or added to item 1. If CO is present in flue gases, small amounts of unburned hydrogen and hydrocarbons may also be present. The small losses caused by incomplete combustion of these gases would be included in item 7, if item 7 was determined by subtracting items 1 to 6 from the fuel heating value. The overall thermal efficiency of combustion equipment is defined as
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Useful heat Thermal efficiency, % = 100 -------------------------------------------------Heating value of fuel
(23)
Equation (24) can be used to estimate efficiency for equipment where item 7 is small or radiation and convection are useful heat: Thermal efficiency, % = Heating value of fuel – q 2 + q 3 + q 4 + q 5 + q 6 100 ------------------------------------------------------------------------------------------------------------------Heating value of fuel
(24)
Using heating values based on gas volume, a gas appliance’s thermal efficiency can be computed with sufficient precision by the following equation: 100 Q h – Q f l = ----------------------------------Qh
(25)
where = thermal efficiency, % Qh = higher heating value of fuel gas per unit volume Qfl = flue gas losses per unit volume of fuel gas
To produce heat efficiently by burning any common fuel, flue gas losses must be minimized by (1) providing adequate heat-absorbing surface in the appliance, (2) keeping heat transfer surfaces clean on both fire and water or air sides, and (3) reducing excess air to the minimum level consistent with complete combustion and discharge of combustion products.
Seasonal Efficiency The method just presented is useful for calculating the steadystate efficiency of a heating system or device. Unfortunately, the seasonal efficiency can be significantly different from the steadystate efficiency. The primary factor affecting seasonal efficiency is flue loss during the burner-off period. The warm stack that exists at the end of the firing period can cause airflow in the stack while the burner is off, which can remove heat from furnace and heat exchanger components, the structure itself, and pilot flames. Also, if combustion air is drawn from the heated space within the structure, the heated air lost must be at least partly replaced with cold infiltrated air. For further discussion of seasonal efficiency, see Chapters 10 and 33 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment and Chapter 19 of this volume.
Table 16 NOx Emission Factors for Combustion Sources NOx Emission Factor, mg/MJ of Heat Input Source Gas-fired equipment Small industrial boilers Commercial boilers Residential furnaces Distillate-oil-fired small industrial boilers, commercial boilers, and residential furnaces Residual-oil-fired small industrial boilers and commercial boilers
8.
Without Emission Controls
With Emission Controls
60 43 39
9 9 14
60
160
COMBUSTION CONSIDERATIONS
Air Pollution Combustion processes constitute the largest single source of anthropogenic (human-caused) air pollution. Pollutants can be grouped into five categories: • Products of incomplete fuel combustion - Combustible aerosols (solid and liquid), including smoke, soot, and organics, but excluding ash - Carbon monoxide CO - Gaseous hydrocarbons • Carbon dioxide CO2 • Oxides of nitrogen (collectively referred to as NOx) - Nitric oxide NO - Nitrogen dioxide NO2 • Emissions resulting from fuel contaminants - Sulfur oxides, primarily sulfur dioxide SO2 and small quantities of sulfur trioxide SO3 - Ash - Trace metals • Emissions resulting from additives - Combustion-controlling additives - Mercaptans - Other additives Emission levels of nitrogen oxides and products of incomplete combustion are directly related to the combustion process and can be controlled, to some extent, by process modification. Emissions from fuel contaminants are related to fuel selection and are slightly affected by the combustion process. Emissions from additives must be considered in the overall evaluation of the merits of using additives. Carbon dioxide as a pollutant has gained attention because of its suspected effect on global warming. Carbon dioxide is produced by HVAC&R equipment (either directly or as a result of generating the electric power to operate the HVAC&R equipment), transportation, industry, and other sources. Carbon dioxide emissions can be minimized by increasing appliance operating efficiencies and using fuels with higher hydrogen content. Nitrogen oxides are produced during combustion, either (1) by thermal fixation (reaction of nitrogen and oxygen at high combustion temperatures), or (2) from fuel nitrogen (oxidation of organic nitrogen in fuel molecules). Unfortunately, high excess air and high flame temperature techniques, which ensure complete fuel combustion, tend to promote NOx formation. NO levels in flames where the reactants are premixed tend to peak with excess air levels around 10%. Higher excess air levels generally reduce the amount of NOx and flame temperatures.
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28.18 Table 16 lists some NOx emission factors for fuel-burning equipment. Differences in emissions are caused by flame temperature and different levels of fuel nitrogen. Carbon monoxide emissions depend less on fuel type and typically range from 13 to 17 mg/MJ of heat input. For gas-fired commercial and industrial boilers, particulate emissions range from 2.2 to 2.6 mg/MJ. For distillate-oil-fired commercial and industrial boilers, particulates are typically 6.0 mg/MJ. For residential oil-fired equipment, particulate emission factors are 1.3 mg/MJ. For residual-oil-fired equipment, particulate emissions depend on the sulfur content and, to a lesser extent, the mineral content. For a sulfur content of 1%, the particulate emission rate is typically 36 mg/MJ. Emission levels of products of incomplete fuel combustion can be reduced by reducing burner cycling, ensuring adequate excess air, improving mixing of air and fuel (by increasing turbulence, improving distribution, and improving liquid fuel atomization), increasing residence time in the hot combustion zone (possibly by decreasing the firing rate), increasing combustion zone temperatures (to speed reactions), and avoiding quenching the flame before reactions are completed. Relative humidity of combustion air affects the amount of NOx produced and must be considered when specifying acceptable NOx emission rates and measuring NOx production during appliance tests. The relative contribution of each of these mechanisms to the total NOx emissions depends on the amount of organic nitrogen in the fuel. Natural gas normally contains very little nitrogen. Virtually all NOx emissions with gas firing are due to the thermal mechanism. Nitrogen content of distillate oil varies, but an average of 20 ppm of fuel NOx is produced (about 20 to 30% of the total NOx). Levels in residual oil can be significantly higher, with fuel NOx contributing heavily to the total emissions. Thermal fixation depends strongly on flame maximum temperature. For example, increasing the flame temperature from 1400 to 1500°C increases thermal NOx tenfold. Therefore, methods to control thermal NOx are based on methods to reduce the maximum flame temperature. Flue gas recirculation is perhaps the most effective method for commercial and industrial boilers. In gas-fired boilers, NOx can be reduced 70% with 15 to 20% recirculation of flue gas into the flame. The NOx reduction decreases with increasing fuel nitrogen content. With distillate-oil firing, reductions of 60 to 70% can be achieved. In residual-oil-fired boilers, flue gas recirculation can reduce NOx emissions by 15 to 30%. The maximum rate of flue gas recirculation is limited by combustion instability and CO production. Two-stage firing is the only technique that reduces NOx produced both by thermal fixation and fuel nitrogen in industrial and utility applications. The fuel-rich or air-deficient primary combustion zone retards NOx formation early in combustion (when NOx forms most readily from fuel nitrogen), and avoids peak temperatures, reducing thermal NOx. Retrofit low-NOx burners that control air distribution and fuel air mixing in the flame zone can be used to achieve staged combustion. With oil firing, NOx reductions of 20 to 50% can be obtained with low-NOx burners. Application of flue gas recirculation and other control methods to residential, oil-fired heating systems was reviewed by Butcher et al. (1994). The following are some methods of reducing NOx emissions from gas-fired appliances (Murphy and Putnam 1985): • • • • • • •
Total premix Burner adjustment Flame inserts (radiation screens or rods) Staged combustion and delayed mixing Secondary air baffling Catalytic and radiant burners Pulse
2017 ASHRAE Handbook—Fundamentals (SI) Radiation screens or rods (flame inserts) surrounding or inserted into the flame absorb radiation to reduce flame temperature and retard NOx formation. Proprietary appliance burners with no flame inserts have been produced to comply with the very strict NOx emission limitations of California’s Air Quality Management Districts. The U.S. EPA sets limits on air pollutant emissions (Source Performance Standards) from boilers larger than 3 MW of heat input. In addition, states set emission regulations that are at least as strict at the federal limits and may apply to smaller equipment. The EPA’s automobile emission standard is 0.62 g of NO2 per kilometre, which is equivalent to 750 ng/J of NOx emission. California’s maximum is 0.25 g/km, equivalent to 300 ng/J. California’s Air Quality Management Districts for the South Coast (Los Angeles) and the San Francisco Bay Area limit NOx emission to 14 ng/J of useful heat for some natural gas-fired central furnaces. For further discussion of air pollution aspects of fuel combustion, see EPA (1971a, 1971b).
Portable Combustion Analyzers (PCAs) These battery-powered electronic instruments, also known as flue gas analyzers (FGAs), have been widely used for over 30 years. Their sensors measure various gases found in flues of combustion appliances, whether fired by gas, oil, or solid fuels. Users place the analyzer’s probe in the flue of an appliance to extract combustion gases for measurement. These gases pass through a water trap to create a dry gas for measurement, and then have excess particles removed by a filter before entering the PCA with help from its internal pump. The PCA’s sensors measure or calculate real-time readings of oxygen (O2), carbon monoxide (CO), and carbon dioxide (CO2). Most PCAs also measure temperature, pressure, and draft; some measure the ratio of CO to CO2, detect gas leaks, or measure other gases such as nitric oxide (NO), nitrogen dioxide (NO2), sulfur dioxide (SO2), and hydrocarbons (CH4) and transfer test results to a printer, PC, smartphone, or tablet. If in doubt, ask the manufacturer which model is appropriate for a specific application. Pay particular attention to the user manual’s care instructions. These include switching on in outdoor air to allow sensors to tare (zero out) correctly, not leaving a PCA overnight in a cold vehicle, and ensuring the PCA is annually recertified by an authorized service provider to keep its metrological performance within specification. An annual calibration certificate is essential. As a minimum, it should include • • • •
PCA serial number and type Calibration date Next calibration due date or certificate expiration date Reference to test gases and instruments used during calibration process • Ambient conditions at time of test • Calibration results of parameters calibrated • Uncertainty measurement
All calibrated PCAs should have a label or sticker applied including the name of the company who performed the work and the PCA’s next service date. The label should reference the PCA using a serial number or certificate number, be tamperproof, and be placed where it can easily be seen. PCAs generally rely on service software, usually only available from the manufacturer or an approved service center. Using nonauthorized companies may invalidate the PCA’s warranty and may lead to incorrect replacement parts being used. An American PCA performance standard, AHRI Standard 1260P, is in development, but most PCAs are already certified to European Standard CEN Standard 50379, which has three parts:
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Combustion and Fuels • Part 1 defines general requirements of and test methods for PCAs • Part 2 defines PCA performance requirements where statutory inspections are required • Part 3 defines PCA performance requirements where nonstatutory inspections are required (e.g., in most European countries other than Germany and Austria) CEN Standard 50379 defines which gases are included (CO, NO, SO2, O2, and CO2 if fitted to the PCA) and what accuracy, response, and performance criteria PCAs must meet. The standard also tests for sensor cross sensitivity, long life, drop, and vibration to ensure reliability in normal applications under competent use. Some PCAs are used to measure ambient levels of CO and CO2 in residential or commercial environments. In Europe, these PCAs must meet the performance requirements of CEN Standard 50543; some PCAs are already certified to this standard as well as to CEN Standard 50379.
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Condensation and Corrosion Fuel-burning systems that cycle on and off to meet demand cool down during the off cycle. When the appliance starts again, condensate forms briefly on surfaces until they are heated above the dewpoint temperature. Low-temperature corrosion occurs in system components (heat exchangers, flues, vents, chimneys) when their surfaces remain below the dew-point temperature of flue gas constituents (water vapor, sulfides, chlorides, fluorides, etc.) long enough to cause condensation. Corrosion increases as condensate dwell time increases. Acids in flue gas condensate are the principal substances responsible for low-temperature corrosion in fuel-fired systems. Sulfuric, hydrochloric, and other acids are formed when acidic compounds in fuel and air combustion products combine with condensed moisture in appliance heat exchangers, flues, or vents. Corrosion can be avoided by maintaining these surfaces above the flue gas dew point. In high-efficiency, condensing-type appliances and economizers, flue gas temperatures are intentionally reduced below the flue gas dew-point temperatures to achieve efficiencies approaching 100%. In these systems, surfaces subjected to condensate must be made of corrosion-resistant materials. The most corrosive conditions exist at the leading edge of the condensing region, especially areas that experience evaporation during each cycle (Stickford et al. 1988). Draining condensate retards the concentration of acids on system surfaces; regions from which condensate partially or completely drains away before evaporation are less severely attacked than regions from which condensate does not drain before evaporation. The metals most resistant to condensate corrosion are stainlesssteel alloys with high chromium and molybdenum content, and nickel-chromium alloys with high molybdenum content (Stickford et al. 1988). Aluminum experiences general corrosion rather than pitting when exposed to flue gas condensate. If applied in sufficiently thick cross section to allow for metal loss, aluminum can be used in condensing regions. Most ceramic and high-temperature polymer materials resist the corrosive effects of flue gas condensate. These materials may have application in the condensing regions, if they can meet the structural and temperature requirements of a particular application. In coal-fired power plants, the rate of corrosion for carbon steel condensing surfaces by mixed acids (primarily sulfuric and hydrochloric) is reported to be maximum at about 50°C ± 10 K (Davis 1987). Mitigation techniques include (1) acid neutralization with a base such as NH3 or Ca(OH)2; (2) use of protective linings of glassfilled polyester or coal-tar epoxy; and (3) replacing steel with molybdenum-bearing stainless steels, nickel alloys, polymers, or other corrosion-resistant materials. Other elements in residual fuel oils and coals that contribute to high-temperature corrosion include sodium, potassium, and vanadium. Each fuel-burning system com-
28.19 ponent should be evaluated during installation, or when modified, to determine the potential for corrosion and the means to retard corrosion (Paul et al. 1988). If fuel-burning appliances accumulate condensate that does not evaporate, the condensate must be routed into a trapped drainage system. Because the condensate may be acidic, the drainage system must be suitable and environmentally acceptable. Condensate freezing must be considered in cold climates.
Abnormal Combustion Noise in Gas Appliances During development of a new boiler, furnace, or other gas-fired appliance, tonal noise can be unacceptable. Because the frequency of the tone is equal to a resonance frequency of the system, this problem is often called a combustion resonance, but this term is misleading: changing the appliance’s resonance frequency merely changes the frequency of the tone without much effect on the amplitude. The proper term is combustion-driven oscillation, which is caused by feedback instability. Pressure oscillations in the combustion chamber (which manifest themselves as objectionable noise) also interact with the flame, modulating the instantaneous rate of combustion, which, in turn, causes more pressure oscillations (Putnam 1971). This feedback involves the acoustic response of the combustion chamber and of the fuel-air supply system, as well as that of the flame. For some combinations of response properties, the feedback loop is unstable. Instabilities may result from either perturbation of the mixture flow or of the air/fuel ratio. Instability because of a fluctuating air/fuel ratio is more likely to occur at low frequencies (Herrin et al. 2012). The feedback loop suggested by Baade (1978) and modified by ASHRAE research project RP-1517 (Herrin et al. 2012) is very useful for understanding and solving oscillation problems. Baade identified three transfer functions that must be determined. Transfer functions Z and H are acoustic and can be determined experimentally or by simulation. Transfer function Gf is related to the flame and is best determined experimentally, though a few models are available. Figure 7 shows a schematic of Baade’s feedback loop stability model for predicting combustion oscillations. Perturbations to the volume velocity of the flame, which are external to the feedback loop, are indicated by q˜ ext . The driving point impedance Z of the combustion chamber is the ratio of the oscillating pressure p˜ to the volume velocity in the combustion chamber q˜ tot or Q˜ . The transfer function H relates the perturbation of the mixture flow q˜ i to the acoustic pressure in the combustion chamber p˜ . As shown in Figure 7, the acoustic impedance of the chamber relates the flame perturbations to the sound pressure in the chamber. Acoustic impedance is primarily a function of the geometry of the combustion chamber and length of the exhaust. The sound pressure in the chamber in turn modulates either the mixture supply or air/fuel ratio. The respective transfer functions relating the sound pressures to the mixture supply or air/fuel ratio fluctuations are mainly controlled by the geometry of the intake. Mixture supply or air/fuel ratio fluctuations then drive the flame perturbations completing the feedback loop. The flame transfer function, which relates flame perturbations to the mixture supply and air/fuel ratio fluctuations, is related to several factors, including the burner type. Predicting instability in a design is generally not practical for domestic or small commercial appliances because there is not enough information to predict the acoustic response of some of the components, particularly the flame. A model of the feedback loop is very useful, however, for solving existing oscillation problems, where the only concern is the particular frequency at which the oscillation occurs. Reducing the response of the flame, air/fuel mixture supply, or combustion chamber at that frequency should be the focus. This concept can be easily
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28.20
2017 ASHRAE Handbook—Fundamentals (SI) REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore.
Fig. 7 Feedback Loop Stability Model Defined by Baade (1978, 2004)
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(Herrin et al. 2012)
demonstrated with a small brazing torch in a tube of variable length (Baade 1987, 2004). In some systems, the flame can be modified to reduce its response at the oscillation frequency. Often, this involves simply changing the fuel/air ratio further away from the stochiometric ratio (Elsari and Cummings 2003; Goldschmidt et al. 1978), thus lengthening the flame, which can also be done by increasing the size of burner ports (Matsui 1981; Schimmer 1979). Other possibilities for reducing flame response are using a suitable mix of differently sized burner ports (Kagiya 2000) and modifying the heat transfer characteristics of the burner matrix (Schreel et al. 2002). The fuel supply system response can be reduced by avoiding resonance at or near the frequency of oscillation (Kilham et al. 1964) or by tuning the supply system to an antiresonance at that frequency (Neumann 1974). Higher-flow-resistance burners add acoustic damping to the fuel supply system. Designs for this can be evaluated by modeling the mixture supply system using transmission matrices (Munjal 1987) and computer programs for matrix multiplication, which are widely available (Baade and Tomarchio 2008). Be careful to avoid any unnecessary resonances in the combustion chamber (Herrin et al. 2012). Structural resonances are sometimes the primary cause of a combustion oscillation, and can be identified by using an impact hammer and measuring the components’ vibrational response using an accelerometer. Increasing damping can effectively resolve instabilities, if the system can increase damping enough: any damping less than the critical amount will have very little effect. Acoustic damping or sound absorption can be added by installing perforated panels or fiber. However, damping treatments are only likely to be effective for instabilities at high frequencies. In some systems, the oscillation frequency may be a function of the flue pipe length. In such cases, consider changing the length as well as adding damping. For large systems, active feedback may be used to eliminate combustion oscillations (Sattinger et al. 2000). However, active feedback is not likely to be cost effective for residential and small commercial systems.
Soot Soot deposits on flue surfaces of a boiler or heater act as an insulating layer over the surface, reducing heat transfer to the water or air. Soot can also clog flues, reduce draft and available air, and prevent proper combustion. Proper burner adjustment can minimize soot accumulation. Using off-specification fuel can contribute to soot generation.
AHRI. [In development.] Performance rating of flue gas combustion analyzers. Standard 1260P. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. ASME. 2015. Boiler and pressure vessel code. American Society of Mechanical Engineers, New York. ASTM. 2015. Standard classification of coals by rank. Standard D388-15. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for fuel oils. ANSI/ASTM Standard D396-15c. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for diesel fuel oils. ANSI/ASTM Standard D975-15c. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2013. Standard specification for liquefied petroleum (LP) gases. ANSI/ASTM Standard D1835-13. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard specification for gas turbine fuel oils. ANSI/ASTM Standard D2880-15. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2016. Specification for automotive spark-ignition engine fuel. Standard D4814-16b. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Specification for biodiesel fuel blend stock (B100) for middle distillate fuels. Standard D6751-15ce1. American Society for Testing and Materials, West Conshohocken, PA. Baade, P.K. 1978. Design criteria and models for preventing combustion oscillations. ASHRAE Transactions 84(1):449. Baade, P.K. 1987. Demonstration of methods for solving combustion “resonance” noise problems. NOISE-CON ’87 Proceedings, pp. 195-200. Baade, P.K. 2004. How to solve abnormal combustion noise problems. Sound and Vibration 4(7):22-27. Baade, P.K., and M.J. Tomarchio. 2008. Tricks and tools for solving abnormal combustion noise problems. Sound and Vibration (July):12-17. Butcher, T.A., L. Fisher, B. Kamath, T. Kirchstetter, and J. Batey. 1994. Nitrogen oxides (NOx) and oil burners. Proceedings of the 1994 Oil Heat Technology Conference and Workshops. BNL Report 52430. Brookhaven National Laboratory, Upton, NY. Butcher, T.A., S.W. Lee, Y. Celebi, and W. Litzke. 1997. Fouling of heattransfer surfaces in oil-fired boilers for domestic heating. Journal of the Institute of Energy 70:151-159. CEN. 2012. Specification for portable electrical apparatus designed to measure combustion flue gas parameters of heating appliances. Part 1—General requirements and test methods. Part 2—Performance requirements for apparatus used in statutory inspections and assessment. Standard 50379-1 and 2. Comité Européen de Normalisation, Brussels. CEN. 2011. Electronic portable and transportable apparatus designed to detect and measure carbon dioxide and/or carbon monoxide in indoor ambient air—Requirements and test methods. Standard 50543. Comité Européen de Normalisation, Brussels. Coward, H.F., and G.W. Jones. 1952. Limits of flammability of gases and vapors. Bulletin 503. U.S. Bureau of Mines, Washington, D.C. Davis, J.R., ed. 1987. Metals handbook, 9th ed., vol. 13, Corrosion. ASM International, Metals Park, OH. Dickson, C.L., and G.P. Sturm, Jr. 1994. Heating oils. National Institute for Petroleum and Energy Research, Bartlesville, OK. DOE. 2017. Fischer-Tropsch synthesis. National Energy Technology Laboratory, U.S. Department of Energy, Washington, D.C. www.netl.doe.gov /research/coal/energy-systems/gasification/gasifipedia/ftsynthesis. Elsari, M., and A. Cummings. 2003. Combustion oscillations in gas fired appliances: Eigen-frequencies and stability regimes. Applied Acoustics 64(6):565-580. EPA. 1971a. Standards of performance for new stationary sources, Group I. Federal Register 36, August 17. U.S. Environmental Protection Agency, Washington, D.C.
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Combustion and Fuels
28.21
EPA. 1971b. Standards of performance for new stationary sources, Group I, Part II. Federal Register 36, December 23. U.S. Environmental Protection Agency, Washington, D.C. EPA. 2017. Basic information about landfill gas. Landfill Methane Outreach Program (LMOP). U.S. Environmental Protection Agency, Washington, D.C. www.epa.gov/lmop/basic-information-about-landfill-gas. Fleck, B.A., S.C. Arnold, M.Y. Ackerman, J.D. Dale, W.E. Klaczek, and D.J. Wilson. 2007. Field testing and residential fan-assisted gas-fired furnaces: Effects of altitude and assessment of current derating standards. ASHRAE Research Project RP-1182, Final Report. Gas engineers handbook. 1965. Industrial Press, New York. Goldschmidt, V., R.G. Leonard, J.F. Riley, G. Wolfbrandt, and P.K. Baade. 1978. Transfer functions of gas flames: Methods of measurement and representative data. ASHRAE Transactions 84(1):466-476. GPA. 1997. Liquefied petroleum gas specifications and test methods. Standard 2140-97. Gas Processors Association, Tulsa, OK. Hartman, I. 1958. Dust explosions. In Mechanical engineers’ handbook, 6th ed., Section 7, pp. 41-48. McGraw-Hill, New York. Hazard, H.R. 1971. Gas turbine fuels. In Gas turbine handbook. Gas Turbine Publications, Stamford, CT. Herrin, D., L. Zhou, and T. Li. 2012. Validation of a low-order acoustic model of boilers and its application for diagnosing combustion driven oscillations. ASHRAE Research Project RP-1517, Final Report. Kagiya, S. 2000. Practical burner design for the suppression of combustion oscillations. Annual Technical Report Digest, vol. 10. Tokyo Gas Co. Kilham, J.K., E.G. Jackson, and T.J.B. Smith. 1964. Oscillatory combustion in tunnel burners. 10th Symposium (International) on Combustion, England, pp. 1231-1240. The Combustion Institute, Pittsburgh, PA. Lee, S.W., I. He, T. Herage, V. Razbin, E. Kelly, and B. Young. 2002a. Influence of fuel sulphur in particulate emissions from pilot-scale research furnaces. Natural Resources Canada. CETC 02-08 (CF). Lee, S.W., I. He, T. Herage, B. Young, and E. Kelly. 2002b. Fuel sulphur effects on particulate emissions from oil combustion systems under accelerated laboratory conditions. Natural Resources Canada. CETC 02-09 (CF). Matsui, Y. 1981. An experimental study on pyro-acoustic amplification of premixed laminar flames. Combustion and Flame 43:199-209. Munjal, M.L. 1987. Acoustics of ducts and mufflers. Wiley Interscience, Hoboken, NJ. Murphy, M.J., and A.A. Putnam. 1985. Burner technology bulletin: Control of NOx emissions from residential gas appliances. Report GRI-85/0132. Battelle Columbus Division for Gas Research Institute. Neumann, E.G. 1974. An impedance condition for avoiding acoustic oscillations generated by gas flames. Acustica 30:229-235. NFPA. 1962. Fire-hazard properties of flammable liquids, gases and volatile solids. In Fire protection handbook, 12th ed., Tables 6-126, pp. 6-131 ff. National Fire Protection Association, Quincy, MA. North American combustion handbook, 3rd ed. 1986. North American Manufacturing Co., Cleveland, OH. Paul, D.D., A.L. Rutz, S.G. Talbert, J.J. Crisafolli, G.R. Whitacre, and R.D. Fischer. 1988. User’s manual for Vent-II Ver. 3.0—A dynamic microcomputer program for analyzing gas venting systems. Report GRI-88/ 0304. Battelle Columbus Division for Gas Research Institute. Putnam, A. 1971. Combustion-driven oscillations in industry. Elsevier, New York. Sattinger, S.S., Y. Neumeier, A. Nabi, B.T. Zinn, D.J. Amos, and D.D. Darling. 2000. Sub-scale demonstration of the active feedback control of gas-turbine combustion instabilities. ASME Transactions, Journal of Engineering for Gas Turbines and Power 122(2):262-268.
Schimmer, H. 1979. Selbsterregte Schwingungen in Brennkammern—Ihre Entstehung und Massnahmen zu ihrer Vermeidung. Gas Waerme International 26:17-23. Schreel, K.R.A.M., R. Rook, and L.P.H. de Goey. 2002. The acoustic response of burner stabilized flat flames. Proceedings of the Combustion Institute, Sapporo, Japan, vol. 29, pp. 115-121. Scott, G.S., G.W. Jones, and F.E. Scott. 1948. Determination of ignition temperatures of combustible liquids and gases. Analytical Chemistry 20: 238-241. Shelton, E.M. 1974. Burner oil fuels. Petroleum Products Survey 86. U.S. Bureau of Mines, Washington, D.C. Shnidman, L. 1954. Gaseous fuels. American Gas Association, Arlington, VA. Stickford, G.H., S.G. Talbert, B. Hindin, and D.W. Locklin. 1988. Research on corrosion-resistant materials for condensing heat exchangers. Proceedings of the 39th Annual International Appliance Technical Conference. Suchovsky, C., R. Sheridan, and J. Nagorka. 2011. Recommendations based on field testing and analysis of high altitude installations of gas-fired boilers and water heaters. ASHRAE Research Project RP-1388, Final Report. Trinks, W. 1947. Simplified calculation of radiation from non-luminous furnace gases. Industrial Heating 14:40-46. U.S. Bureau of Mines. Semiannually. Mineral industry surveys, motor gasolines. Washington, D.C. Zabetakis, M.G. 1956. Research on the combustion and explosion hazards of hydrogen-water vapor-air mixtures. Division of Explosives Technology, Progress Report 1. U.S. Bureau of Mines, Washington, D.C.
BIBLIOGRAPHY ASA. 2013. Acoustical terminology. ANSI/ASA Standard S1.1-2013. American National Standards Institute, New York, and Acoustical Society of America, Melville, NY. Bonne, U., and A. Patani. 1982. Combustion system performance analysis and simulation study. Report GRI-81/0093 (PB 83-161 406). Honeywell SSPL, Bloomington, MN. EPA. 2016. Compilation of air pollutant emission factors. Report AP-42. U.S. Environmental Protection Agency, Washington, D.C. www.epa.gov /air-emissions-factors-and-quantification/ap-42-compilation-air-emission -factors. Fricker, N., and C.A. Roberts. 1979. An experimental and theoretical approach to combustion driven oscillations. Gas Waerme International 28(13). Gas Appliance Technology Center, Gas Research Institute. Manufacturer update on status of GATC research on heat-exchanger corrosion, May 1984. Battelle Columbus Laboratories and American Gas Association Laboratories. Lewis, B., and G. von Elbe. 1987. Combustion, flames, and explosion of gases, 3rd ed. Academic Press, New York. NFPA/AGA. 2015. National fuel gas code, Section 11.1.2. ANSI/NFPA Standard 54-2015. National Fire Protection Association, Quincy, MA. ANSI/AGA Standard Z223.1-2012. American Gas Association, Washington, D.C. Stickford, G.H., S.G. Talbert, and D.W. Locklin. 1987. Condensate corrosivity in residential condensing appliances. Proceedings of the International Symposium on Condensing Heat Exchangers, Paper 3, BNL Report 52068, 1 and 2. Brookhaven National Laboratory, Upton, NY.
Related Commercial Resources
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Related Commercial Resources CHAPTER 29
REFRIGERANTS Refrigerant Properties .................................................................................................................. 29.1 Refrigerant Performance .............................................................................................................. 29.6 Safety ............................................................................................................................................. 29.6 Leak Detection .............................................................................................................................. 29.9 Compatibility with Construction Materials ................................................................................ 29.10
EFRIGERANTS are the working fluids in refrigeration, airconditioning, and heat-pumping systems. They absorb heat from one area, such as an air-conditioned space, and reject it into another, such as outdoors, usually through evaporation and condensation, respectively. These phase changes occur both in absorption and mechanical vapor compression systems, but not in systems operating on a gas cycle using a fluid such as air. (See Chapter 2 for more information on refrigeration cycles.) The design of the refrigeration equipment depends strongly on the selected refrigerant’s properties. Tables 1 and 2 list standard refrigerant designations, some properties, and safety classifications from ASHRAE Standard 34. Refrigerant selection involves compromises between conflicting desirable thermophysical properties. A refrigerant must satisfy many requirements, some of which do not directly relate to its ability to transfer heat. Chemical stability under conditions of use is an essential characteristic. Safety codes may require a nonflammable refrigerant of low toxicity for some applications. Environmental consequences of refrigerant leaks must also be considered. Cost, availability, efficiency, compatibility with compressor lubricants and equipment materials, and local and national regulations are other concerns. Latent heat of vaporization is another important property. On a molar basis, fluids with similar boiling points have almost the same latent heat. Because compressor displacement is defined on a volumetric basis, refrigerants with similar boiling points produce similar refrigeration effect with a given compressor. On a mass basis, latent heat varies widely among fluids. Efficiency of a theoretical vapor compression cycle is maximized by fluids with low vapor heat capacity. This property is associated with fluids having a simple molecular structure and low molecular mass. Transport properties (e.g., thermal conductivity and viscosity) affect performance of heat exchangers and piping. High thermal conductivity and low viscosity are desirable. No single fluid satisfies all the attributes desired of a refrigerant; consequently, various refrigerants are used. This chapter describes the basic characteristics of various refrigerants, and Chapter 30 lists thermophysical properties.
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R
1.
REFRIGERANT PROPERTIES
Global Environmental Properties Chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs) can affect both stratospheric ozone and climate change, whereas hydrofluorocarbons (HFCs) can affect climate change. Minimizing all refrigerant releases from systems is important not only because of environmental impacts, but also because charge losses lead to insufficient system charge levels, which in turn result in suboptimal operation and lowered efficiency. Stratospheric Ozone Depletion. The stratospheric ozone layer filters out the UV-B portion of the sun’s ultraviolet (UV) radiation. Overexposure to this radiation increases the risk of skin cancer, The preparation of this chapter is assigned to TC 3.1, Refrigerants and Secondary Coolants.
29.1 Copyright © 2017, ASHRAE
cataracts, and impaired immune systems. It also can damage sensitive crops, reduce crop yields, and stress marine phytoplankton (and thus human food supplies from the oceans). In addition, exposure to UV radiation degrades plastics and wood. Stratospheric ozone depletion has been linked to the presence of chlorine and bromine in the stratosphere. Chemicals with long atmospheric lifetimes can migrate to the stratosphere, where the molecules break down from interaction with ultraviolet light or through chemical reaction. Chemicals such as CFCs and HCFCs release chlorine, which reacts with stratospheric ozone. Ozone-depleting substances, including CFCs and HCFCs, are to be phased out of production under the Montreal Protocol (UNEP 2009). In the United States, production and importation of CFCs were banned completely in 1996. HCFCs are being phased down, with complete phaseout set for 2030. In 2010, to meet the Montreal Protocol phasedown schedule, U.S. regulations banned production and importation of HCFC-142b and HCFC-22 for use in new equipment. Reclaimed CFC and HCFC refrigerants that meet the requirements of AHRI Standard 700 can continue to be used for servicing existing systems. A complete list of U.S. regulations for CFC and HCFC refrigerants, including phaseout schedules, may be found at www.epa.gov/ozone-layer-protection. A summary of the phaseout schedules for CFCs and HCFCs for both developed and developing countries may be found at www.unep.org/ozonaction/topics/hcfc .asp. Global Climate Change. The average global temperature is determined by the balance of energy from the sun heating the earth and its atmosphere and of energy radiated from the earth and the atmosphere to space. Greenhouse gases (GHGs), such as CO2 and water vapor, as well as small particles trap heat at and near the surface, maintaining the average temperature of the Earth’s surface about 34 K warmer than would be the case if these gases and particles were not present (the greenhouse effect). Global warming (also called global climate change) is a concern because of an increase in the greenhouse effect from increasing concentrations of GHGs attributed to human activities. The major GHG of concern is CO2 released to the atmosphere when fossil fuels (coal, oil, and natural gas) are burned for energy. Methane (CH4), nitrous oxide (N2O), CFCs, HCFCs, HFCs, hydrofluoroethers (HFEs), hydrofluoro-olefins (HFOs), perfluorocarbons (PFCs), nitrogen trifluoride (NF3), and sulfur hexafluoride (SF6) are also GHGs. In 1988, the United Nations Environment Programme (UNEP) and the World Meteorological Organization (WMO) established the Intergovernmental Panel on Climate Change (IPCC) to provide an objective source of information about the causes of climate change, its potential environmental and socioeconomic consequences, and the adaptation and mitigation options to respond to it. According to the IPCC (2014), atmospheric concentration of carbon dioxide has increased by more than 40% over the past 250 years, primarily from burning fossil fuels, with some contribution from forestry and land use. Concentration of methane has increased by over 150%, and of nitrous oxide by about 20%. IPCC (2014) deems atmospheric concentrations of fluorochemicals, including fluorocarbon gases
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29.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Refrigerant Data and Safety Classifications
Refrigerant Number
Chemical Namea,b
Licensed for single user. © 2017 ASHRAE, Inc.
Methane Series 11 Trichlorofluoromethane 12 Dichlorodifluoromethane 12B1 Bromochlorodifluoromethane 13 Chlorotrifluoromethane 13B1 Bromotrifluoromethane 14 Tetrafluoromethane (carbon tetrafluoride) 21 Dichlorofluoromethane 22 Chlorodifluoromethane 23 Trifluoromethane 30 Dichloromethane (methylene chloride) 31 Chlorofluoromethane 32 Difluoromethane (methylene fluoride) 40 Chloromethane (methyl chloride) 41 Fluoromethane (methyl fluoride) 50 Methane
Normal Boiling Point,a °C
Safety Group
Chemical Formulaa
Molecular Massa
CCl3F CCl2F2 CBrClF2 CClF3 CBrF3 CF4 CHCl2F CHClF2 CHF3 CH2Cl2 CH2ClF CH2F2 CH3Cl CH3F CH4
137.4 120.9 165.4 104.5 148.9 88.0 102.9 86.5 70.0 84.9 68.5 52.0 50.4 34.0 16.0
24 –30 –4 –81 –58 –128 9 –41 –82 40 –9 –52 –24 –78 –161
CCl2FCClF2 CClF2CClF2 CClF2CF3 CF3CF3 CHCl2CF3 CHClFCF3 CHF2CF3 CH2FCF3 CH3CCl2F CH3CClF2 CH3CF3 CH3CHF2 CH3CH3
187.4 170.9 154.5 138.0 153.0 136.5 120.0 102.0 117.0 100.5 84.0 66.0 30.0
48 4 –39 –78 27 –12 –48 –26 32 –10 –47 –24 –89
A1 A1 A1 A1 B1 A1 A1 A1
A1 A1 A1 A1 A1 B1 A1 A1 B2 A2L B2 A3
Ethane Series 113 114 115 116 123 124 125 134a 141b 142b 143a 152a 170
1,1,2-trichloro-1,2,2-trifluoroethane 1,2-dichloro-1,1,2,2-tetrafluoroethane Chloropentafluoroethane Hexafluoroethane 2,2-dichloro-1,1,1-trifluoroethane 2-chloro-1,1,1,2-tetrafluoroethane Pentafluoroethane 1,1,1,2-tetrafluoroethane 1,1-dichloro-1-fluoroethane 1-chloro-1,1-difluoroethane 1,1,1-trifluoroethane 1,1-difluoroethane Ethane
Ethers E170
Dimethyl ether
CH3OCH3
46.0
–25
A3
Propane Series 218 Octafluoropropane 227ea 1,1,1,2,3,3,3-heptafluoropropane 236fa 1,1,1,3,3,3-hexafluoropropane 245fa 1,1,1,3,3-pentafluoropropane 290 Propane
CF3CF2CF3 CF3CHFCF3 CF3CH2CF3 CF3CH2CHF2 CH3CH2CH3
188.0 170.0 152.0 134.0 44.0
–37 –16 –1 15 –42
A1 A1 A1 B1 A3
–(CF2)4–
200.0
–6
A1
CH3CH2CH2CH3 CH(CH3)2CH3 CH3(CH2)3CH3 (CH3)2CHCH2CH3
58.1 58.1 72.15 72.15
0 –12 36.1 27.8
A3 A3 A3 A3
CH3CH2OCH2CH3 HCOOCH3
74.1 60.0
35 32
B2
CH3NH2 CH3CH2(NH2)
31.1 45.1
–7 17
Cyclic Organic Compounds (see Table 2 for blends) C318 Octafluorocyclobutane Miscellaneous Organic Compounds Hydrocarbons 600 Butane 600a 2-methylpropane (isobutane) 601 Pentane 601a 2-methylbutane (isopentane) Oxygen Compounds 610 Ethyl ether 611 Methyl formate Sulfur Compounds 620 (Reserved for future assignment) Nitrogen Compounds 630 Methanamine (methyl amine) 631 Ethanamine (ethyl amine)
A2 A2L A2 A3
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Refrigerants
29.3 Table 1
Refrigerant Number
Refrigerant Data and Safety Classifications (Continued)
Chemical Namea,b
Chemical Formulaa
Inorganic Compounds 702 Hydrogen 704 Helium 717 Ammonia 718 Water 720 Neon 728 Nitrogen 732 Oxygen 740 Argon 744 Carbon dioxide 744A Nitrous oxide 764 Sulfur dioxide
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Unsaturated Organic Compounds 1150 Ethene (ethylene) 1233zd(E) Trans-1-chloro-3,3,3-trifluoro-1-propene 1234yf 2,3,3,3-tetrafluoro-1-propene 1234ze(E) Trans-1,3,3,3-tetrafluoro-1-propene 1270 Propene (propylene) 1336mzz(Z) Cis-1,1,1,4,4,4-hexafluoro-2-butene Source: ANSI/ASHRAE Standard 34-2010. aChemical name, chemical formula, molecular mass, and normal boiling point are not part of this standard.
Molecular Massa
Normal Boiling Point,a °C
Safety Group
H2 He NH3 H2O Ne N2 O2 Ar CO2 N2O SO2
2.0 4.0 17.0 18.0 20.2 28.1 32.0 39.9 44.0 44.0 64.1
–253 –269 –33 100 –246 –196 –183 –186 –78c –90 –10
A3 A1 B2L A1 A1 A1
CH2=CH2 CF3CH=CHCl CF3CF=CH2 CF3CH=CHF CH3CH=CH2 CF3CH=CHCF3
28.1 130.5 114.0 114.0 42.1 164.1
–104 18 –29.4 –19.0 –48 33
A3 A1 A2L A2L A3 A1
bPreferred
cSublimes.
A1 A1 B1
chemical name is followed by the popular name in parentheses.
Table 2 Data and Safety Classifications for Refrigerant Blends Refrig. No. Composition (Mass %) Zeotropes 400 R-12/114 (must be specified) 401A R-22/152a/124 (53.0/13.0/34.0) 401B R-22/152a/124 (61.0/11.0/28.0) 401C R-22/152a/124 (33.0/15.0/52.0) 402A R-125/290/22 (60.0/2.0/38.0) 402B R-125/290/22 (38.0/2.0/60.0) 403A R-290/22/218 (5.0/75.0/20.0) 403B R-290/22/218 (5.0/56.0/39.0) 404A R-125/143a/134a (44.0/52.0/4.0) 405A R-22/152a/142b/C318 (45.0/7.0/5.5/42.5) 406A R-22/600a/142b (55.0/4.0/41.0) 407A R-32/125/134a (20.0/40.0/40.0) 407B R-32/125/134a (10.0/70.0/20.0) 407C R-32/125/134a (23.0/25.0/52.0) 407D R-32/125/134a (15.0/15.0/70.0) 407E R-32/125/134a (25.0/15.0/60.0) 407F R-32/125/134a (30.0/30.0/40.0) 408A R-125/143a/22 (7.0/46.0/47.0) 409A R-22/124/142b (60.0/25.0/15.0) 409B R-22/124/142b (65.0/25.0/10.0) 410A R-32/125 (50.0/50.0) 410B R-32/125 (45.0/55.0) 411A R-1270/22/152a (1.5/87.5/11.0) 411B R-1270/22/152a (3.0/94.0/3.0) 412A R-22/218/142b (70.0/5.0/25.0) 413A R-218/134a/600a (9.0/88.0/3.0) 414A R-22/124/600a/142b (51.0/28.5/4.0/16.5) 414B R-22/124/600a/142b (50.0/39.0/1.5/9.5) 415A R-22/152a (82.0/18.0) 415B R-22/152a (25.0/75.0) 416A R-134a/124/600 (59.0/39.5/1.5) 417A R-125/134a/600 (46.6/50.0/3.4) 417B R-125/134a/600 (79.0/18.3/2.7) 418A R-290/22/152a (1.5/96.0/2.5) 419A R-125/134a/E170 (77.0/19.0/4.0) 420A R-134a/142b (88.0/12.0) 421A R-125/134a (58.0/42.0)
Composition Tolerances
(±2.0 /+0.5,–1.5/±1.0) (±2/+0.5,–1.5/±1.0) (±2/+0.5,–1.5/±1.0) (±2.0/+0.1,–1.0/±2.0) (±2/+0.1,–1/±2) (+0.2,–2/±2/±2) (+0.2,–2/±2/±2) (±2/±1/±2) (±2/±1/±1 /±2) sum of R-152a and R-142b = (+0.0, –2.0) (±2/±1/±1) (±2/±2/±2) (±2/±2/±2) (±2/±2/±2) (±2/±2/±2) (±2,±2,±2) (±2,±2,±2) (±2/±1/±2) (±2/±2/±1) (±2/±2/±1) (+0.5,–1.5/+1.5,–0.5) (±1/±1) (+0,–1/+2,–0/+0,–1) (+0,–1/+2,–0/+0,–1) (±2/±2/±1) (±1/±2/±0,–1) (±2/±2/±0.5/+0.5,–1) (±2/±2/±0.5/+0.5,–1) (±1/±1) (±1/±1) (+0.5,–1/+1,–0.5/+1,–0.2) (±1.1/±1/+0.1,–0.4) (±1/±1/+0.1,–0.5) (±0.5/±1/±0.5 (±1/±1/±1) (±1,–0/+0,–1) (±1/±1)
Molec- Normal Normal ular Bubble Dew Safety Massa Point, °C Point, °C Group
94.4 92.8 101 101.6 94.7 92 103.3 97.6 111.9 89.9 90.1 102.9 86.2 91 83.8 82.1 87 97.4 96.7 72.6 75.6 82.4 83.1 92.2 104 96.9 101.6 81.9 70.2 111.9 106.7 113.1 84.6 109.3 101.8 111.8
–34.4 –35.7 –30.5 –49.2 –47.2 –44.0 –43.8 –46.6 –32.9 –32.7 –45.2 –46.8 –43.8 –39.4 –42.8 –46.1 –45.5 –35.4 –36.5 –51.6 –51.5 –39.7 –41.6 –36.4 –29.3 –34.0 –34.4 –37.5 –27.7 –23.4 –38.0 –44.9 –41.2 –42.6 –25.0 –40.8
–28.8 –30.8 –23.8 –47.0 –44.9 –42.3 –42.3 –45.8 –24.5 –23.5 –38.7 –42.4 –36.7 –32.7 –35.6 –39.7 –45.0 –27.5 –29.7 –51.5 –51.4 –37.2 –41.3 –28.8 –27.6 –25.8 –26.1 –34.7 –26.2 –21.8 –32.9 –41.5 –40.1 –36.0 –24.2 –35.5
A1 A1 A1 A1 A1 A1 A1 A1 A1 A2 A1 A1 A1 A1 A1 A1 A1 A1 A1 A1 A1 A2 A2 A2 A2 A1 A1 A2 A2 A1 A1 A1 A2 A2 A1 A1
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29.4
2017 ASHRAE Handbook—Fundamentals (SI) Table 2
Data and Safety Classifications for Refrigerant Blends (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Refrig. No. Composition (Mass %) 421B 422A 422B 422C 422D 423A 424A 425A 426Aª 427Aª 428Aª 429A 430A 431A 432A 433A 433B 433C 434A 435A 436A 436B 437A 438A 439A 440A 441A 442A 443A 444A 444B 445A 446A 447A 448A 449A 449B 450A 451A 451B 452A 453A 454A 454B
R-125/134a (85.0/15.0) R-125/134a/600a (85.1/11.5/3.4) R-125/134a/600a (55.0/42.0/3.0) R-125/134a/600a (82.0/15.0/3.0) R-125/134a/600a (65.1/31.5/3.4) R-134a/227ea (52.5/47.5) R-125/134a/600a/600/601a (50.5/47.0/0.9/1.0/0.6) R-32/134a/227ea (18.5/69.5/12.0) R-125/134a/600a/601a (5.1/93.0/1.3/0.6) R-32/125/143a/134a (15.0/25.0/10.0/50.0) R-125/143a/290/600a (77.5/20.0/0.6/1.9) R-E170/152a/600a (60.0/10.0/30.0) R-152a/600a (76.0/24.0) R-290/152a (71.0/29.0) R-1270/E170 (80.0/20.0) R-1270/290 (30.0/70.0) R-1270/290 (5.0/95.0) R-1270/290 (25.0/75.0) R-125/143a/134a/600a (63.2/18.0/16.0/2.8) R-E170/152a (80.0/20.0) R-290/600a (56.0/44.0) R-290/600a (52.0/48.0) R-125/134a/600/601 (19.5/78.5/1.4/0.6) R-32/125/134a/600/601a (8.5/45.0/44.2/1.7/0.6) R-32/125/600a (50.0/47.0/3.0) R-290/134a/152a (0.6/1.6/97.8) R-170/290/600a/600 (3.1/54.8/6.0/36.1) R-32/125/134a/152a/227ea (31.0/31.0/30.0/3.0/5.0) R-1270/290/600a (55.0/40.0/5.0) R-32/152a/1234ze(E) (12.0/5.0/83.0) R-32/152a/1234ze(E) (41.5/10.0/48.5) R-744/134a/1234ze(E) (6.0/9.0/85.0) R-32/1234ze(E)/600 (68.0/2.09/3.0) R-32/125/1234ze(E) (68.0/3.5/28.5) R-32/125/1234yf/134a/1234ze(E) (26.0/26.0/20.0/ 21.0/7.0) R-32/125/1234yf/134a (24.3/24.7/25.3/25.7) R-32/125/1234yf/134a (25.2/24.3/23.2/27.3) R-134a/1234ze(E) (42.0/58.0) R-1234yf/134a (89.8/10.2) R-1234yf/134a (88.8/11.2) R-32/125/1234yf (11.0/59.0/30.0) R32/125/134a/227ea/600/601a (20.0/20.0/53.8/5.0/ 0.6/0.6) R-32/1234yf (35/65) R-32/1234yf (68.9/31.1)
Azeotropesb Refrig. No. Composition (Mass %) 500 501 502 503 504 505 506 507Ad 508Ad 508B 509Ad 510Ad 511Ad 512A 513A
R-12/152a (73.8/26.2) R-22/12 (75.0/25.0)c R-22/115 (48.8/51.2) R-23/13 (40.1/59.9) R-32/115 (48.2/51.8) R-12/31 (78.0/22.0)c R-31/114 (55.1/44.9) R-125/143a (50.0/50.0) R-23/116 (39.0/61.0) R-23/116 (46.0/54.0) R-22/218 (44.0/56.0) R-E170/600a (88.0/12.0) R-290/E170 (95.0/5.0) R-134a/152a (5.0/95.0) R-1234yf/134a (56.0/44.0)
Molec- Normal Normal ular Bubble Dew Safety Massa Point, °C Point, °C Group
Composition Tolerances (±1/±1) (±1/±1/+0.1,–0.4) (±1/±1/+0.1,–0.5) (±1/±1/+0.1,–0.5) (+0.9,–1.1/±1/+0.1,–0.4) (±1/±1) (±1/±1/+0.1,–0.2/+0.1,–0.2/+0.1,–0.2) (±0.5/±0.5/±0.5) (±1/±1/+0.1,–0.2/+0.1,–0.2) (±2/±2/±2/±2) (±1/±1/+0.1,–0.2/+0.1,–0.2) (±1/±1/±1) (±1/±1) (±1/±1) (±1/±1) (±1/±1) (±1/±1) (±1/±1) (±1/±1/±1/+0.1,–0.2) (±1/±1) (±1/±1) (±1/±1) (+0.5,–1.8/+1.5,–0.7/+0.1,–0.2/+0.1/–0.2) (+0.5,–1.5/±1.5/±1.5/+0.1,–0.2/+0.1/–0.2) (±1/±1) (±0.1/±0.6/±0.5) (±0.3/±2/±0.6/±2) (±1.0/±1.0/±1.0/+0.5/±1.0) (±2.0/±2.0/±1.2) (±1.0/±1.0/±2.0) (±.1.0/±1.0/±1.0) (±1.0/±1.0/±2.0) (+0.5,–1.0/+2.0,–0.6/+1.0,–1.0) (+1.5,–0.5/+1.5,–0.5/+1.0,–1.0) (+0.5 –2.0/+2.0 –0.5/+0.5– 2.0/+2.0–1.0/+0.5–2.0) (+2.0 –1.0/+1.0–0.2/+0.2– 1.0/+1.0–0.2) (+0.3 –1.5 /+1.5 –0.3 /+0.3 –1.5 /+1.5 –0.3) (±2.0/±2.0) (±0.2 /±0.2) (±0.2 /±0.2) (±1.7/±1.8/+0.1/–1.0) (±1.0 /±1.0 /±1.0 /±0.5 /+0.1 –0.2 /+0.1 –0.2) (±2.0/±2.0) (±1.0/±1.0) Composition Tolerances
–45.7 –46.5 –40.5 –45.3 –43.2 –24.2 –39.1 –38.1 –28.5 –43.0 –48.3 –26.0 –27.6 –43.1 –46.6 –44.6 –42.7 –44.3 –45.0 –26.1 –34.3 –33.4 –32.9 –43.0 –52.0 –25.5 –41.9 –46.5 –44.8 –34.3 –44.6 –50.3 –49.4 –49.3 –45.9
–42.6 –44.1 –35.6 –42.3 –38.4 –23.5 –33.3 –31.3 –26.7 –36.3 –47.5 –25.6 –27.4 –43.1 –45.6 –44.2 –42.5 –43.9 –42.3 –25.9 –26.2 –25.0 –29.2 –36.4 –51.8 –24.3 –20.4 –39.9 –41.2 –24.3 –34.9 –23.5 –44.0 –44.2 –39.8
A1 A1 A1 A1 A1 A1 A1 A1 A1 A1 A1 A3 A3 A3 A3 A3 A3 A3 A1 A3 A3 A3 A1 A1 A2 A2 A3 A1 A3 A2L A2L A2L A2L A2L A1
87.21 86.37 108.67 72.76 112.56 112.56 88.78
–46.0 –46.1 –23.4 –30.8 –31.0 –31.0 –42.2
–39.9 –40.2 –22.8 –30.5 –30.6 –30.6 –35.0
A1 A1 A1 A2L A2L A2L A1
80.47 62.61
–48.4 –50.9
–41.6 –50.0
A2L A2L
Azeotropic Temperatures, °C
Molecular Massa
0 –41 19 –88 17 115 18 –40 –86 –45.6 0 –25.2 –20 to 40 –20 to 40 27
99.3 93.1 112.0 87.5 79.2 103.5 93.7 98.9 100.1 95.4 124.0 47.24 44.19 67.24 108.4
(±0.5/±0.5) (±1/±1) (±1/±1) (±1.0/±1.0)
Source: ANSI/ASHRAE Standard 34-2010. aMolecular mass and normal boiling point are not part of this standard. bAzeotropic refrigerants exhibit some segregation of components at conditions of temperature and pressure other than those at which they were formulated. Extent of segregation depends on the particular azeotrope and hardware system configuration.
116.9 113.6 108.5 116.3 109.9 126 108.4 90.3 101.6 90.4 107.5 50.8 64 48.8 42.8 43.5 44 43.6 105.7 49.04 49.33 49.87 103.7 99.1 71.2 66.2 48.2 81.77 43.47 96.7 92.78 103.1 62 63.04 86.28
cExact
Normal Boiling Safety Point, °C Group –33 –41 –45 –88 –57 –30 –12 –46.7 –86 –88.3 –47 –25.2 –42.1 –24.0 –29.1
A1 A1 A1
A1 A1 A1 A1 A3 A3 A2 A1
composition of this azeotrope is in question, and additional experimental studies are needed. R-508, and R-509 are allowed designations for R-507A, R-508A, and R-509A because of a change in designations after assignment of R-500 through R-509. Corresponding changes were not made for R-500 through R-506.
dR-507,
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerants
29.5
(CFCs, HCFCs, and HFCs) and sulfur hexafluoride, to be about 2% of the 2010 GHG (CO2 equivalent) emissions. Global Environmental Characteristics of Refrigerants. Atmospheric release of CFC and HCFC refrigerants (see Table 3) contributes to depletion of the ozone layer. The measure of a material’s ability to deplete stratospheric ozone is its ozone depletion potential (ODP), a value relative to R-11’s value of 1.0. It is the nonzero ODP of these refrigerants that led to the phaseout of their production and use under the Montreal Protocol. The global warming potential (GWP) of a GHG is an index describing its relative ability to trap radiant energy compared to CO2 (R-744), which has a very long atmospheric lifetime. Measurements of climate impact of refrigerant emissions are hence often reported in CO2 equivalents. GWP may be calculated for any particular integration time horizon (ITH). Typically, a 100 year ITH is used for calculation of GWPs for regulatory purposes, and may be designated as GWP100. Halocarbons (CFCs, HCFCs, and HFCs) and many nonhalocarbons (e.g., hydrocarbons, carbon dioxide) are GHGs. HFOs, or unsaturated HFCs, and blends using them are being developed and promoted as low-GWP alternatives to the existing halocarbon refrigerants. HFOs are also GHGs but their GWPs are drastically lower than those of HFCs. The energy refrigeration appliances consume is often produced from fossil fuels, which results in emission of CO2, a contributor to global warming. This indirect effect associated with energy consumption is frequently much larger than the direct effect of refrigerant emissions. The total equivalent warming impact (TEWI) of an HVAC&R system is the sum of direct refrigerant emissions expressed in terms of CO2 equivalents, and indirect emissions of CO2 from the system’s energy use over its service life (Fischer et. al. 1991). Another measure is life-cycle climate performance (LCCP), which includes TEWI and adds direct and indirect emissions effects associated with manufacturing the refrigerant and endof-life disposal (ARAP 1999). Ammonia (R-717), hydrocarbons, HCFCs, most HFCs, and HFOs have shorter atmospheric lifetimes than CFCs because they are largely destroyed in the lower atmosphere by reactions with OH Table 3A Refrigerant Environmental Properties Refrigerant CFC-11 CFC-12 CFC-13 CFC-113 CFC-114 CFC-115 HCFC-22 HCFC-123 HCFC-124 HCFC-142b HFC-23 HFC-32 HFC-125 HFC-134a HFC-143a HFC-152a HFC-227ea HFC-236fa HFC-245fa PFC-116 PFC-218 C318 R-744
Atmospheric Lifetime, yearsa 45 100 640 85 300 1700 12 1.3 5.8 17.9 270 4.9 29 14 52 1.4 34.2 240 7.6 10 000 2600 3200
ODPb
GWP100a
1 1 1 0.8 1 0.6 0.055 0.02 0.022 0.065 0 0 0 0 0 0 0 0 0 0 0 0.00 0.00
4750 10 900 14 400 6130 10 000 7370 1810 77 609 2310 14 800 675 3500 1430 4470 124 3220 9810 1030 12 200 8830 10 300 1
Sources: IPCC (2007), UNEP (2009). aAtmospheric lifetimes and GWP 100s from Table 2.14 of IPCC (2007). bODP values stipulated for reporting under the Montreal Protocol from UNEP (2009), Section 1.1, Annexes A, B, and C, pp. 25-27.
radicals. A shorter atmospheric lifetime generally results in lower ODP and GWP100 values. Table 3A gives values for atmospheric lifetime, ODP, and GWP100 of refrigerants being phased out under the Montreal Protocol and of their replacements, alone or as components of blends from the IPCC (2007). Table 3B shows the latest scientific assessment values for these refrigerants and for new refrigerants developed since IPCC (2007) was published. Because HFCs do not contain chlorine or bromine, their ODP values are negligible (Ravishankara et al. 1994) and thus are shown as 0 in Tables 3A and 3B. Nonhalocarbon refrigerants listed have zero ODP and very low GWP100. As shown in Tables 3A and 3B, there are differences between the values stipulated for reporting under the Montreal and Kyoto protocols and the latest scientific values. These differences are not large enough to significantly alter design decisions based on the numbers in the table. All these values have rather wide error bands and may change with each assessment of the science. Changes in GWP assessments are largely dominated by changes in understanding of CO2, which is the reference chemical. Table 4 provides ODP values for blends based on Montreal Protocol reporting values of the component fluids (Table 3A). It also gives IPCC (2007, 2013) GWP100 values based on component fluid values in Tables 3A and 3B, respectively. Table 3B Refrigerant Environmental Properties Refrigerant CFC-11 CFC-12 CFC-13 CFC-113 CFC-114 CFC-115 HCFC-22 HCFC-123 HCFC-124 HCFC-142b HCFO-1233zd(E) HE-E170 HFC-23 HFC-32 HFC-125 HFC-134a HFC-143a HFC-152a HFC-227ea HFC-236fa HFC-245fa HFO-1234yf HFO-1234ze(E) HFO-1336mzz(Z) PFC-116 PFC-218 C318 HC-290 HC-600 HC-600a HC-601a HC-1270 R-717 R-744
Atmospheric Lifetime, yearsa 45 100 640 85 190 1020 11.9 1.3 5.9 17.2 0.071 0.015b 222 5.2 28.2 13.4 47.1 1.5 38.9 242 7.7 0.029 0.045 0.07 10 000 2600 3200 0.034b 0.016b 0.009b 0.001b
ODPb
GWP100a
1 0.73 1 0.81 0.50 0.26 0.034 0.01 0.02 0.057 0.00034 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00
4660 10 800 13 900 5820 8590 7670 1760 79 527 1980 1 1b 12 400 (11 700)c 677 (650)c 3170 (2800)c 1300 (1300)c 4800 (3800)c 138 (140)c 3350 (2900)c 8060 (6300)c 858 <1 <1 2 11 100 (9200)c 8900 (7000)c 9540 (8700)c 5b 4b ~20b ~20b 1.8b 1 (1)c
Sources: IPCC (2013). aAtmospheric lifetimes and GWP 100s from IPCC (2013) except where indicated. bFrom Table 2-7 of Calm et al. (2015) cGWP 100 values stipulated for reporting under Kyoto Protocol from Table 2-5 of Calm et al. (2015).
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
29.6
2017 ASHRAE Handbook—Fundamentals (SI)
Table 4 Environmental Properties of Refrigerant Blends; based on Montreal Protocol Reporting ODP and IPCC AR4 and AR5 GWP100 of Components
Licensed for single user. © 2017 ASHRAE, Inc.
GWP100* Refrigerant ODP* AR4 AR5 401A 401B 401C 402A 402B 403A 403B 404A 405A 406A 407A 407B 407C 407D 407E 407F 408A 409A 409B 410A 410B 411A 411B 412A 413A 414A 414B 415A 415B 416A 417A 417B 418A 419A 420A 421A 421B 422A 422B 422C 422D 423A 424A 425A 426A 427A
0.02 0.03 0.02 0.01 0.02 0.03 0.02 0.00 0.02 0.04 0.00 0.00 0.00 0.00 0.00 0.00 0.02 0.03 0.03 0.00 0.00 0.03 0.03 0.04 0.00 0.03 0.03 0.03 0.009 0.008 0.00 0.00 0.03 0.00 0.007 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00
1180 1290 933 2790 2420 3120 4460 3920 5330 1940 2110 2800 1770 1630 1550 1820 3150 1580 1560 2090 2230 1600 1710 2290 2050 1480 1360 1510 546 1080 2350 3030 1740 2970 1540 2630 3190 3140 2530 3090 2730 2280 2440 1510 1510 2140
1130 1240 876 2570 2260 3100 4460 3940 4970 1780 1920 2550 1620 1490 1420 1670 3260 1480 1470 1920 2050 1560 1660 2170 1950 1380 1270 1470 544 975 2130 2740 1690 2690 1380 2380 2890 2850 2290 2800 2470 2270 2210 1430 1370 2020
Refrigerant ODP* 428A 429A 430A 431A 432A 433A 433B 433C 434A 435A 436A 436B 437A 438A 439A 440A 441A 442A 443A 444A 444B 445A 446A 447A 448A 449A 449B 450A 451A 451B 452A 453A 454A 454B 500 501 502 503 504 507A 508A 508B 509A 510A 511A 512A 513A
0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.50 0.29 0.20 0.60 0.10 0.00 0.00 0.00 0.01 0.00 0.00 0.00 0.00
GWP100* AR4
AR5
3610 3120 19 20 99 110 38 44 2 2 3 4 3 5 3 4 3250 3080 26 28 11 12 11 12 1810 1640 2260 2060 1980 1830 144 156 5 5 1890 1750 2.5 4 93 89 296 295 135 118 461 461 584 572 1390 1360 1400 1280 1410 1300 605 547 150 133 164 146 2140 1950 1770 1640 239 238 466 467 8080 8010 4080 4020 4660 4790 14 600 13 300 4140 4300 3990 3990 13 200 11 600 13 400 11 700 5740 5760 3 3 3 5 189 196 631 573
*Derived based on weighted averages of blend component values. AR4 = IPCC (2007) AR5 = IPCC (2013)
Physical Properties Table 5 lists some physical properties of commonly used refrigerants, a few very-low-boiling-point cryogenic fluids, some newer refrigerants, and some older refrigerants of historical interest, arranged in increasing order of atmospheric boiling point. Table 5 also includes the freezing point, critical properties, and refractive index. Of these properties, normal boiling point is most important because it is a direct indicator of the temperature at which a refrigerant can be used. The freezing point must be lower than any contemplated usage. The critical properties describe a material at the point where the distinction between liquid and gas is lost. At higher
temperatures, no separate liquid phase is possible for pure fluids. In refrigeration cycles involving condensation, a refrigerant must be chosen that allows this change of state to occur at a temperature somewhat below the critical. Cycles that reject heat at supercritical temperatures (e.g., cycles using carbon dioxide) are also possible. Lithium Bromide/Water and Ammonia/Water Solutions. These are the most commonly used working fluids in absorption refrigeration systems. (See Chapter 30 for property data.)
Electrical Properties Tables 6 and 7 list electrical characteristics of refrigerants that are especially important in hermetic systems.
Sound Velocity The practical velocity of a gas in piping or through openings is limited by the velocity of sound in the gas. Chapter 30 has sound velocity data for many refrigerants. The velocity increases when temperature is increased and decreases when pressure is increased. REFPROP software (NIST 2010) can be used to calculate the sound velocity at superheated conditions. The velocity of sound can be calculated from the equation Va =
dp d S =
dp d T
(1)
where Va p S T
= = = = = =
sound velocity, m/s pressure, Pa density, kg/m3 cp /cv = ratio of specific heats entropy, kJ/(kg·K) temperature, K
Sound velocity can be estimated from tables of thermodynamic properties. Change in pressure with a change in density (dp/d) can be estimated at either constant entropy or constant temperature. It is simpler to estimate at constant temperature, but the ratio of specific heats must also be known.
2.
REFRIGERANT PERFORMANCE
Chapter 2 describes several methods of calculating refrigerant performance, and Chapter 30 includes tables of thermodynamic properties of refrigerants. Table 8 shows the theoretical calculated performance of a number of refrigerants for a standard cycle of various evaporation temperatures and 303 K condensation. For blend refrigerants, the average temperature in the evaporator and condenser is used. In most cases, suction vapor is assumed to be saturated, and compression is assumed adiabatic or at constant entropy. For R-113 and R-600a, for example, these assumptions cause some liquid in the discharge vapor. In these cases, it is assumed that discharge vapor is saturated and that suction vapor is slightly superheated. Note that actual operating conditions and performance may differ significantly from numbers in the table because of additional factors such as compressor efficiency and transport properties.
3.
SAFETY
Tables 1 and 2 summarize toxicity and flammability characteristics of many refrigerants. In ASHRAE Standard 34, refrigerants are classified according to the hazard involved in their use. The toxicity and flammability classifications yield six safety groups (A1, A2, A3, B1, B2, and B3) for refrigerants. Group A1 refrigerants are the least hazardous, group B3 the most hazardous. The letter designates toxicity class based on allowable exposure: • Class A: Refrigerants that have an occupational exposure limit (OEL) of 400 ppm or greater. • Class B: Refrigerants that have an OEL of less than 400 ppm.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Refrigerants
29.7 Table 5 Physical Properties of Selected Refrigerantsa
Refrigerant Chemical Name or Composition (% by Mass)
Licensed for single user. © 2017 ASHRAE, Inc.
No.
Chemical Formula
Boiling Pt.f (NBP) at Molecular Mass 101.325 kPa, °C
Freezing Point, °C
Critical Critical Critical Refractive Temperature, Pressure, Density, Index of °C kPa kg/m3 Liquidb,c
728
Nitrogen
N2
28.013
–195.8
–210.0
–146.96
3395.8
313.3
729 740
Air Argon
— Ar
28.959 39.948
–194.25 –185.85
— –189.34
–140.59 –122.46
3789.6 4863.0
335.94 535.6
732
Oxygen
O2
31.999
–182.96
–218.79
–118.57
5043.0
436.14
50 14 170 508A 508B 23 13 744 504 32 410A 125 1270 143a 507A 404A 502 407C 290 22 115 500 717 12 1234yf 134a 152a 1234ze (E) 124 600a 142b C318 600 1336mzz (Z) 114 1233zd(E)
Methane CH4 Tetrafluoromethane CF4 Ethane C2H6 R-23/116 (39/61) — R-23/116 (46/54) — Trifluoromethane CHF3 Chlorotrifluoromethane CClF3 Carbon dioxide CO2 R-32/115 (48.2/51.8) — Difluoromethane CH2F2 R-32/125 (50/50) — Pentafluoroethane C2HF5 Propylene C3H6 Trifluoroethane CH3CF3 R-125/143a (50/50) — R-125/143a/134a (44/52/4) — R-22/115 (48.8/51.2) — R-32/125/134a (23/25/52) — Propane C3H8 Chlorodifluoromethane CHClF2 Chloropentafluoroethane CClF2CF3 R-12/152a (73.8/26.2) — Ammonia NH3 Dichlorodifluoromethane CCl2F2 2,3,3,3-tetrafluoroprop-1-ene CF3CF=CH2 Tetrafluoroethane CF3CH2F Difluoroethane CHF2CH3 Trans-1,3,3,3CF3CH=CHF tetrafluoropropene Chlorotetrafluoroethane CHClFCF3 Isobutane C4H10 Chlorodifluoroethane CClF2CH3 Octafluorocyclobutane C4F8 Butane C4H10 Cis-1,1,1,4,4,4-hexafluoro-2- CF3CH=CHCF3 butene Dichlorotetrafluoroethane CClF2CClF2 Trans-1-chloro-3,3,3-trifluoro- CF3CH=CHCl 1-propene Trichlorofluoromethane CCl3F Dichlorotrifluoroethane CHCl2CF3 Dichlorofluoroethane CCl2FCH3 Trichlorotrifluoroethane CCl2FCClF2 Water H2O
16.043 88.005 30.07 100.1 95.394 70.014 104.46 44.01 79.249 52.024 72.585 120.02 42.08 84.041 98.859 97.604 111.63 86.204 44.096 86.468 154.47 99.303 17.03 120.91 114.04 102.03 66.051 114.04
–161.48 –128.05 –88.581 –87.60 –87.6 –82.018 –81.48 –78.4d –57.906 –51.651 –51.446 –48.09 –47.62 –47.241 –46.741 –46.222 –45.174 –43.627 –42.11 –40.81 –39.25 –33.603 –33.327 –29.752 –29.45 –26.074 –24.023 –18.95
–182.46 –183.61 –182.8 — — –155.13 –181.15 –56.558e — –136.81 — –100.63 –185.2 –111.81 — — — — –187.62 –157.42 –99.39 — –77.655 –157.05
136.48 58.122 100.5 200.03 58.122 164.1
–11.963 –11.75 –9.15 –5.975 –0.49 92.1
—
170.92 130.5
3.586 18.1
—
137.37 152.93 116.95 187.38 18.015
23.708 27.823 32.05 47.585 99.974
11 123 141b 113 7183
Notes: cFor the sodium D line. aData from NIST (2010) REFPROP v. 9.0. bTemperature of measurement (°C, unless kelvin is noted) shown in dSublimes. parentheses. Data from CRC (1987), unless otherwise noted.
The numeral denotes flammability: • Class 1: No flame propagation in air at 60°C and 101.3 kPa • Class 2: Exhibits flame propagation in air at 60°C and 101.3 kPa, lower flammability limit (LFL) greater than 0.10 kg/m3 at 23°C and 101.3 kPa, and heat of combustion less than 19 000 kJ/kg • Optional class 2L: class 2 refrigerants may be classified as 2L if they exhibit a maximum burning velocity of no more than 100 mm/s at 23.0°C and 101.3 kPa
–103.3 –118.59
–82.586 –45.64 32.72 10.192 11.205 26.143 28.85 30.978 62.138 78.105 71.358 66.023 91.061 72.707 70.617 72.046 80.507 86.034 96.74 96.145 79.95 102.09 132.25 111.97 94.7 101.06 113.26 109.37
4599.2 3750.0 4872.2 3650.8 3771.6 4832 3879 7377.3 4428.8 5782.0 4902.6 3617.7 4554.8 3761.0 3705 3728.9 4016.8 4629.8 4251.2 4990.0 3129.0 4168.6 11 333.0 4136.1 3382.2 4059.3 4516.8 3636.3
1.205 (83 K) 589.3 nm — 1.233 (84 K) 589.3 nm 1.221 (92 K) 589.3 nm — — — — — — 1.146 (25)2 1.195 (15) — — — — 1.3640 (–50)1 — — — — — 1.3397 (–42) 1.234 (25)2 1.221 (25)2 — 1.325 (16.5) 1.288 (25)2
162.66 625.66 206.18 567.58 568.45 526.5 582.88 467.6 504.68 424 459.53 573.58 230.03 431.0 490.77 486.53 568.70 484.23 220.4 523.84 614.8 495.1 225.0d 565.0 475.55 511.9 — 368 — 489.24
–199.15 –159.42 –130.43 –39.8 –138.27
122.28 134.66 137.11 115.23 151.98 171.3
3624.3 3629.0 4055.0 2777.5 3796.0 2903
560.0 225.5 466.0 619.97 227.94 504.67
–92.5
145.68 165.6
3257.0 3580
579.97 1.294 (25) 480.769 —
–110.47 –107.15 –103.5 –36.22 0.01
197.96 183.68 204.4 214.06 373.95
4407.6 3661.8 4212.0 3392.2 22 064.0
eAt
527 kPa. fBubble point used for blends
554.0 550.0 458.6 560.0 322.0
— 1.3514 (–25)1 — 1.3562 (–15)1 —
1.362 (25)2 — — 1.357 (25)2 —
References: 1Kirk and Othmer (1956). 2Bulletin B-32A (DuPont). 3Handbook of Chemistry (1967).
• Class 3: Exhibits flame propagation in air at 60°C and 101.3 kPa and LFL less than or equal to 0.10 kg/m3at 23.0°C and 101.3 kPa or heat of combustion greater than or equal to 19 000 kJ/kg Refrigerant blends are assigned the flammability safety classification of the worst case of fractionation of the blend (i.e., the composition during fractionation that results in the highest concentration of flammable components). For class 2 or 3 refrigerants or refrigerant blends that show no flame propagation when tested at
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29.8
2017 ASHRAE Handbook—Fundamentals (SI)
Table 6 Electrical Properties of Liquid Refrigerants
Table 7 Electrical Properties of Refrigerant Vapors
Refrigerant
No. 11
12
13
Licensed for single user. © 2017 ASHRAE, Inc.
22
Trichlorofluoromethane
Dichlorodifluoromethane
Chlorotrifluoromethane Chlorodifluoromethane
23
Trifluoromethane
32
Difluoromethane
113
Trichlorotrifluoroethane
114
123 124
Dichlorotetrafluoroethane
2,2-dichloro-1,1,1trifluoroethane 2-chloro-1,1,1,2tetrafluoroethane Chlorotetrafluoroethane Pentafluoroethane
28.9 a 25 25 28.9 a 25 25 25 –30 20 23.9 a 25 25 –30 20 a 25 30 a 25 31.1 a 25 a
2.28 1.92 2.5 2.32 2.13 1.74 2.1 2.100 2.14 2.3 1.64 6.11 6.12 6.6 6.42 6.3 5.51 14.27 14.67 2.44 1.68 2.6 2.17 1.83 2.2 4.50
25
4.89
25 20 25 134a 1,1,1,2-tetrafluoroethane a 25 143a 1,1,1-trifluoroethane 25 236fa 1,1,1,3,3,3-hexafluoropropane 25 245fa 1,1,1,3,3-pentafluoropropane 25 290 Propane a 404A R-125/143a/134a (44/52/4) a 25 407C R-32/125/134a (23/25/52) a 25 410A R-32/125 (50/50) a 25 500 R-12/152a (73.8/26.2) a 507A R-125/143a (50/50) a 25 508A R-23/116 (39/61) –30 0 508B R-23/116 (46/54) –30 0 717 Ammonia 20.6 744 Carbon dioxide 0 1234yf 2,3,3,3-tetrafluoro-1-propene 25 21
4.0 4.94 5.10 9.51 9.87 9.78 7.89 6.82 1.27 7.58 8.06 8.74 10.21 7.78 5.37 1.80 6.97 7.94 6.60 5.02 7.24 5.48 15.5 1.59 7.6 7.7
124a 125
Refrigerant
Volume Temp., Dielectric Resistivity, Chemical Name or °C Constant M·m Ref. Composition (% By Mass)
a = ambient temperature References: 1 Data from E.I. DuPont de Nemours & Co., Inc. 2 Beacham and Divers (1955) 3 Eiseman (1955) 4 Fellows et al. (1991)
63 680 90
53 900 >120
120
0.83 75
45 490 >120 66 470 >70 14 700
1 2 3 9 1 2 3 4 9 4 1 2 3 9 3 4 6 9 1 2 3 1 2 3 4 9
50
17 700
73 840 8450 7420 3920 55 750 5570
5 CRC (1987) 6 Bararo et al. (1997) 7 Pereira et al. (1999) 8 Meurer et al. (2001) 9 Gbur and Byrne (2001) 10 Muller et al. (2011) 11 Data from Honeywell
3 7 9 4 9 9 9 9 2 8 9 8 9 8 9 2 8 9 1 1 1 1 5 5 10 11
No. 11
Dielec- Relative Chemical Name Prestric Dielectric or Composition sure, Temp., Con- Strength, kPa °C stant Nitrogen = 1 (% by mass)
Trichlorofluoromethane
50.7 a 101.3 12 Dichlorodifluoro50.7 methane a 101.3 101.3 13 Chlorotrifluoro50.7 methane 101.3 14 Tetrafluoromethane 50.7 101.3 22 Chlorodifluoro50.7 methane a 101.3 101.3 32 Difluoromethane 101.3 113 Trichlorotria fluoroethane 40.5 114 Dichlorotetra50.7 fluoroethane a 101.3 116 Hexafluoroethane 95.2 124 2-chloro-1,1,1,2101.3 tetrafluoroethane 125 Pentafluoroethane 101.3 134a 1,1,1,2101.3 tetrafluoroethane 142b Chlorodifluoroethane 94.2 143a Trifluoroethane 86.1 101.3 170 Ethane 101.3 236fa 1,1,1,3,3,3101.3 hexafluoropropane 245fa 1,1,1,3,3101.3 pentafluoropropane 290 Propane a 404A R-125/143a/134a 101.3 (44/52/4) 407C R-32/125/134a (23/ 101.3 25/52) 410A R-32/125 (50/50) 101.3 500 R-12/152a (73.8/ a 26.2) 507A R-125/143a (50/50) 101.3 508A R-23/116 (39/61) a a 101.3 508B R-23/116 (46/54) a a 101.3 717 Ammonia 101.3 a 729 Air 101.3 744 Carbon dioxide 101.3 101.3 1150 Ethylene 101.3 101.3
26.1 b 22.8 28.9 b 22.8 25 28.9 22.8 24.4 22.8
1.002 1.0060
3 2 4 3 2 4 6 3 4 3 4 3 2 4 6 6 2 4 3 2 4 3 6
1.0072 1.0125
6 6
27.2 1.013 25 1.013 25 1.0170 0 1.0015 25 1.0121
3 3 6 1 6
b 22.8 25 25 b 22.8 26.7 b 22.8 22.8 25 25 25
1.0019 1.009
Volume Resistivity, G·m Ref.
3.1 1.0016 1.012
452c 2.4
72.77
1.0064 1.0013 1.4 1.0006 1.0 1.0035 1.004 1.0068 1.0102 1.010 1.0021 1.002
25
1.0066
b 25
1.009 1.0121
25
1.0113
25 b
1.0078 1.024
25 –30 0 25 –30 0 25 0 0 0 0 b 0 22.8
1.0119 1.12 1.31 1.0042 1.13 1.34 1.0042 1.0072
Notes: a = saturation vapor pressure b = ambient temperature c = measured breakdown voltage, volts/mil References: 1 CRC (1987)
74.35
460c 1.3
2113
440c 2.6
94.18
295c 2.8
148.3
6 440c
105.3
2 6 6
470c
0.82 1.00059 1.00099 0.88 1.00144 1.21
76.45
6 2 6 5 5 6 5 5 6 1 4 1 1 4 1 4
2 Beacham and Divers (1955) 3 Fuoss (1938) 4 Charlton and Cooper (1937) 5 Data from E.I. DuPont de Nemours & Co., Inc. 6 Gbur (2005)
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Refrigerants
29.9 Table 8 Comparative Refrigerant Performance per Kilowatt of Refrigeration
Licensed for single user. © 2017 ASHRAE, Inc.
No.
ConNet Refrigdenser Refrig- erant PresCom- erating Circusure, pression Effect, lated, MPa Ratio kJ/kg g/s
Specific Volume Com- Power pressor ConLiquid of Circu- Suction Displace- sumpment, tion, lated, Gas, L/s kW L/s m3/kg
ComCoeffi- pressor cient Disof charge Perfor- Temp., mance °C
1.349 1.012 0.199 0.199 0.190 0.183 0.152 0.110
7.213 4.655 1.305 1.460 1.421 1.304 1.192 1.167
5.35 4.6 6.57 7.34 7.46 7.14 7.81 10.61
132.1 153.6 269.1 101.1 104.9 97.8 155.3 1079.1
7.57 6.51 3.72 9.89 9.54 10.22 6.44 0.93
0.0128 0.0236 0.0075 0.0097 0.0093 0.0086 0.0055 0.0016
0.0285 0.0548 0.2266 0.0949 0.1005 0.0924 0.1448 1.0425
0.2160 0.3567 0.8422 0.9360 0.9565 0.9470 0.9326 0.9643
0.5892 0.5947 0.3471 0.3887 0.3853 0.3651 0.3369 0.3327
1.698 1.681 2.88 2.573 2.595 2.739 2.967 3.007
91.3 57.9 49.1 38.1 38.9 41.3 65.4 140.9
2.909 2.024 0.653 0.643 0.503 0.486 0.476 0.457 0.399 0.396 0.385 0.332 0.250 0.228 0.168 0.123
7.213 4.655 1.928 1.886 1.460 1.421 1.305 1.304 1.192 1.267 1.079 1.167 0.783 0.770 0.578 0.405
2.48 2.3 2.95 2.94 2.9 2.92 2.74 2.86 2.99 3.19 2.8 3.51 3.13 3.37 3.44 3.29
129.5 163.1 258.6 170.9 114.9 118.8 294.4 109.5 165.9 167.1 288.6 1113.0 120.5 153.0 139.6 278.0
7.72 6.13 3.87 5.85 8.70 8.42 3.40 9.13 6.03 5.98 3.47 0.90 8.30 6.54 7.16 3.60
0.0130 0.0222 0.0041 0.0057 0.0085 0.0083 0.0068 0.0077 0.0051 0.0053 0.0072 0.0015 0.0077 0.0055 0.0063 0.0066
0.0127 0.0263 0.0563 0.0406 0.0385 0.0405 0.0986 0.0386 0.0584 0.0588 0.1180 0.3689 0.0718 0.0880 0.1086 0.2984
0.0977 0.1612 0.2178 0.2381 0.3349 0.3410 0.3359 0.3527 0.3520 0.3518 0.4093 0.3313 0.5954 0.5745 0.7798 1.0723
0.2845 0.2786 0.1690 0.1728 0.1798 0.1785 0.1675 0.1724 0.1637 0.1686 0.1669 0.1599 0.1715 0.1650 0.1658 0.1620
3.514 3.588 5.924 5.78 5.564 5.598 5.975 5.799 6.105 5.93 5.987 6.254 5.835 6.063 6.03 6.171
61.3 46.6 59.7 46.6 34.2 34.6 39.3 35.4 47.8 43.9 34.9 82.1 30.0 34.8 30.0 30.0
1.018 1.000 0.703 0.640 0.626 0.588 0.558 0.458 0.401 0.388 0.377 0.280 0.201 0.134 0.045 0.021
1.928 1.886 1.304 1.267 1.192 1.079 1.167 0.880 0.783 0.744 0.770 0.578 0.405 0.283 0.110 0.054
1.89 1.89 1.85 1.98 1.9 1.84 2.09 1.92 1.96 1.92 2.04 2.06 2.01 2.11 2.44 2.57
261.1 175.0 115.3 173.7 171.0 303.9 1127.8 150.4 129.0 126.9 161.0 149.1 296.3 326.9 155.5 137.6
3.83 5.71 8.67 5.76 5.85 3.29 0.89 6.65 7.75 7.88 6.21 6.71 3.37 3.06 6.43 7.27
0.0040 0.0055 0.0073 0.0051 0.0050 0.0068 0.0015 0.0059 0.0072 0.0061 0.0052 0.0059 0.0062 0.0054 0.0044 0.0047
0.0360 0.0260 0.0252 0.0367 0.0377 0.0787 0.2254 0.0453 0.0453 0.0449 0.0542 0.0668 0.1879 0.2853 0.3309 0.5874
0.1381 0.1484 0.2187 0.2112 0.2205 0.2580 0.1998 0.3010 0.3514 0.3536 0.3364 0.4483 0.6332 0.8725 2.1269 4.2686
0.0944 0.0965 0.0956 0.0939 0.0918 0.0931 0.0893 0.0916 0.0941 0.0910 0.0918 0.0918 0.0901 0.0891 0.0878 0.0876
10.602 10.379 10.474 10.655 10.885 10.743 11.186 10.925 10.623 11.004 10.903 10.899 11.084 11.226 11.397 11.409
46.9 39.8 33.2 39.3 40.3 32.6 58.6 34.6 30.0 33.1 32.6 30.0 30.0 30.0 30.0 30.0
EvapoRefrigerant rator PresChemical Name or Composition sure, MPa (% by mass)
Evaporator –31.7°C/Condenser 30°C 744 Carbon dioxide 170 Ethane 1270 Propylene 507A R-125/143a (50/50) 404A R-125/143a/134a (44/52/4) 502 R-22/115 (48.8/51.2) 22 Chlorodifluoromethane 717 Ammonia Evaporator –6.7°C/Condenser 30°C 744 Carbon dioxide 170 Ethane 32 Difluoromethane 410A R-32/125 (50/50) 507A R-125/143a (50/50) 404A R-125/143a/134a (44/52/4) 1270 Propylene 502 R-22/115 (48.8/51.2) 22 Chlorodifluoromethane 407C R-32/125/134a (23/25/52) 290 Propane 717 Ammonia 1234yf 2,3,3,3-Tetrafluoropropene* 134a Tetrafluoroethane 1234ze(E) trans-1,3,3,3-Tetrafluoropropene* 600a Isobutane* Evaporator 7.2°C/Condenser 30°C 32 Difluoromethane 410A R-32/125 (50/50) 502 R-22/115 (48.8/51.2) 407C R-32/125/134a (23/25/52) 22 Chlorodifluoromethane 290 Propane 717 Ammonia 500 R-12/152a (73.8/26.2) 1234yf 2,3,3,3-Tetrafluoropropene* 12 Dichlorodifluoromethane 134a Tetrafluoroethane 1234ze(E) trans-1,3,3,3-Tetrafluoropropene* 600a Isobutane* 600 Butane* 123 Dichlorotrifluoroethane 113 Trichlorotrifluoroethane*
*Superheat required Data from NIST CYCLE_D 4.0, zero subcool, zero superheat unless noted, no line losses, 100% efficiencies, average temperatures.
23.0°C and 101.3 kPa (i.e., no LFL), an elevated temperature flame limit at 60°C (ETFL60) is used in lieu of the LFL for determining flammability classifications.
4.
LEAK DETECTION
Leak detection in refrigeration equipment is of major importance for manufacturers and service engineers.
Electronic Detection Electronic detectors are widely used in manufacture and assembly of refrigeration equipment. Techniques include infrared, solid electrolyte semiconductor, heated electrode/diode, and corona discharge sensors. Instrument operation depends on the variation in signal caused by the presence of refrigerant. These instruments can be refrigerant specific or may detect a variety of refrigerants. Other vapors in the local environment may interfere with the test.
The electronic detector is the most sensitive of the methods discussed here, readily capable of sensing a leak of 2.8 g of refrigerant per year. A portable model is available for field testing. Other models are available with automatic balancing systems that correct for background refrigerant vapors that might be present in the atmosphere around the test area.
Bubble Method The object to be tested is pressurized with air or nitrogen. A pressure corresponding to operating conditions is generally used. If possible, the object is fully immersed in water, and leaks are detected by observing bubbles in the liquid. Adding a detergent to the water decreases surface tension, prevents escaping gas from clinging to the side of the object, and promotes formation of a regular stream of small bubbles. In addition to dwell time, test sensitivity is influenced by clarity of the liquid, lighting, proximity of the leak site to
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29.10
2017 ASHRAE Handbook—Fundamentals (SI)
the operator, and human factors. When immersion is not practical, a solution of soap can be brushed, sprayed, or poured onto joints or other spots where leakage is suspected. Leaking gas forms soap bubbles that can be readily detected. When properly performed under favorable conditions, bubble testing methods can detect leaks as small as 2.8 g of refrigerant per year.
Pressure Change Methods The presence of leaks can be determined by pressurizing or evacuating the internals of the part or system and observing the change in pressure or vacuum over a period of time. The vacuum decay test can give an indication of proper dehydration but, like pressure decay, it does not locate the point of leakage. Test methods that are based on pressure change typically are not sensitive enough to meet the needs of refrigerant components and systems used in HVAC applications. The pressure change test methods are useful to verify that a component or system is free from gross leaks. Typical sensitivity is in the range of dozens of kilograms of refrigerant per year. Leaks can also be determined by pressurizing or evacuating and observing the change in pressure or vacuum over a period of time. This is effective in checking system tightness but does not locate the point of leakage. Licensed for single user. © 2017 ASHRAE, Inc.
UV Dye Method A stable UV-fluorescent dye is introduced into the system to be tested. Operating the system mixes the UV dye uniformly in the oil/ refrigerant system. The dye, which usually prefers oil, shows up at the leak’s location, and can be detected using an appropriate UV lamp. Ensure that the dye is compatible with system components and that no one is exposed to UV radiation from the lamp. This method only finds defects that are large enough to pass liquid, and will only work effectively in regions of the system where enough oil is available to carry the dye. Thus, this method’s sensitivity is typically significantly lower than that of the electronic detection and bubble test methods.
Ammonia Leaks Ammonia can be detected by any of the previously described methods, or by bringing a solution of hydrochloric acid near the object. If ammonia vapor is present, a white cloud or smoke of ammonium chloride forms. Ammonia can also be detected with indicator paper that changes color in the presence of a base. Ensure that adequate ventilation is provided and no one is exposed to ammonia.
5.
COMPATIBILITY WITH CONSTRUCTION MATERIALS
Metals Halogenated refrigerants can be used satisfactorily under normal conditions with most common metals, such as steel, cast iron, brass, copper, tin, lead, and aluminum [an important exception is methyl chloride (R-40) in contact with aluminum]. Under more severe conditions, various metals affect properties such as hydrolysis and thermal decomposition in varying degrees. The tendency of metals to promote thermal decomposition of halogenated compounds is in the following order: (least decomposition) Inconel 18-8 stainless steel nickel copper 1040 steel aluminum bronze brass zinc silver (most decomposition) This order is only approximate, and there may be exceptions for individual compounds or for special use conditions (Downing 1988). Magnesium alloys and aluminum containing more than 2% magnesium are not recommended for use with halogenated compounds where even trace amounts of water may be present. Zinc is not recommended for use with CFC-113. Experience with zinc and other
fluorinated compounds has been limited, but no unusual reactivity has been observed under normal conditions of use in dry systems. However, OxyChem (2009) takes a more conservative position: “Aluminum, zinc, or magnesium equipment should never be allowed to come in contact with methyl chloride.” In 2011, several suspected substitutions of R-40-containing refrigerant mixtures for R-134a in aluminum-containing refrigeration and air-conditioning systems resulted in equipment failures, explosions, and even fatalities (Powell 2012; WorldCargo News 2011). Reactions of methyl chloride with aluminum are known (Dow 2007; Linde 2010; OxyChem 2009), and several publications advise not using aluminum containers for methyl chloride (Dow 2010; European Industrial Gases Association 2010). Studies are under way to quantify the reactivity of methyl chloride concentrations in R-134a/aluminum systems, and identify methods of safe handling, neutralization, and disposal of R-40-contaminated refrigerants. R-40 contamination appears to be part of a broader global issue of fake and counterfeit refrigerants. Press reports [e.g., ACR News (2011a, 2011b, 2012)] describe discovery of R-40 and other contaminants, including hydrocarbons and illegal ozone-depleting substances, found in service market refrigerant containers with fake labels, and in refrigeration and air-conditioning systems on several continents, including North America. Using best practices in the service industry is essential, particularly using refrigerant identification methods to ensure refrigerant systems contain and are refilled with genuine refrigerant. Refrigerant manufacturers are developing additional means to deter counterfeit products. Ammonia should never be used with copper, brass, or other alloys containing copper. Metals compatibility data for ammonia, carbon dioxide, and hydrocarbons are provided by Pruett (1995). Further information on compatibility of refrigerants and lubricants with construction materials metals, elastomers, and plastics is in Chapter 6 of the 2014 ASHRAE Handbook—Refrigeration and in publications of the Air-Conditioning, Heating, and Refrigeration Technology Institute (AHRTI), the research branch of the AirConditioning, Heating, and Refrigeration Institute (AHRI). For example, Rohatgi et al. (2012) investigated thermal and chemical stability of five refrigerants [HFO-1234yf, HFO-1234ze, R-32/ HFO-1234yf (equal mass percentages), R-410A, and R-134a] and three lubricants (two types of POEs and a PVE). The report can be downloaded from AHRI’s web site.
Elastomers Linear swelling of some elastomers in the liquid phase of HCFC and HFC refrigerants is shown in Table 9 (Hamed et al. 1994). Swelling data can be used to a limited extent in comparing the effect of refrigerants on elastomers. However, other factors, such as the amount of extraction, tensile strength, and degree of hardness of the exposed elastomer, must be considered. When other fluids (e.g., lubricants) are present in addition to the refrigerant, the combined effect on elastomers should be determined. Extensive test Table 9
Swelling of Elastomers in Liquid Refrigerants at Room Temperature, % Linear Swell
PolyisoRefrig- prene erant (Sulfur Number Cure) 22 123 124 142b 32 125 134a 143a 152a
10.2 48.0 5.8 10.2 2.7 4.2 1.2 1.9 4.2
Polychloroprene
Butyl Rubber
Styrene Butadiene Rubber
6.1 15.3 2.8 6.5 1.0 2.7 1.2 1.2 3.0
3.9 16.3 3.2 6.2 1.0 2.6 0.6 1.3 1.7
9.8 40.8 4.1 7.3 2.0 3.6 1.0 1.5 2.8
Nitrile FluoroRubber elastomer 51.4 83.7 45.9 8.7 8.3 3.9 5.1 2.0 8.8
33.2 31.6 29.0 31.8 23.2 11.7 25.6 13.6 39.1
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Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerants data for compatibility of elastomers and gasketing materials with refrigerants and lubricants are reported by Hamed et al. (1994). More recent elastomer compatibility data for R-134a and R-1234yf were reported by Minor and Spatz (2008). Six elastomers were contacted with the refrigerants and a PAG lubricant at 100°C for two weeks. Linear swell percentages for the elastomers were very similar in the tests: in the range of –1.4 to +2.1% with R-134a and –1.6 to +1.6% with R-1234yf. Mass gain and hardness changes for the elastomers were also similar in the tests, except for silicone elastomer in R-1234yf having a larger decrease in hardness. Permeation of fluids through elastomers is another consideration, such as with elastomeric hoses in mobile air-conditioning systems. Refrigerant can be lost by outward permeation of refrigerant through the hoses, and water can enter the system by inward diffusion. Data for water and refrigerant permeation through many types of elastomers are presented by Downing (1988). Multilayer hose construction is used to significantly reduce water ingression and refrigerant permeation loss. A typical hose construction might be an outer cover of chlorobutyl elastomer to reduce water ingression, a layer of polyamide to reduce refrigerant permeation, and the inner tube of chloroprene. Hose manufacturers offer variations of such constructions based on their proprietary technology. Refrigerant permeation test data for R-134a and R-1234yf through these types of hose constructions were reported by Minor and Spatz (2008) and Hill and Grimm (2008). In all cases, the permeation rates of R1234yf were lower than those of R-134a. Majurin et al. (2014) also investigated materials compatibility of HFO-1234yf, HFO-1234ze, and a blend of 1234yf/1234ze/R-32 (equal mass percentages) with various elastomers and other motor materials, and found a wide range of compatible materials; the report is available on AHRI’s website. See Pruett (1994) for compatibility data for elastomers with ammonia, carbon dioxide, and hydrocarbons.
Plastics The effect of a refrigerant on a plastic material should be thoroughly examined under conditions of intended use, including the presence of lubricants. Plastics are often mixtures of two or more basic types, and it is difficult to predict the refrigerant’s effect. Mass and visual changes can be used as a general guide of effect, but changes in the plastic’s properties should also be examined. Extensive test data for compatibility of plastics with refrigerants and lubricants are reported by Cavestri (1993), including 23 plastics, 10 refrigerants, 7 lubricants, and 17 refrigerant/lubricant combinations. Refrigerants and lubricants had little effect on most of the plastics, though acrylonitrile-butadiene-styrene, polyphenylene oxide, and polycarbonate were affected enough to be considered incompatible. In a separate study by DuPont Fluoroproducts (2003), acrylonitrile butadiene styrene and polystyrene were determined to have questionable compatibility with HCFC and HFC refrigerants. R-134a and R-1234yf were evaluated for compatibility with typical plastics used in automotive air-conditioning systems (Minor and Spatz 2008). Five plastics (polyester, nylon, epoxy, polyethylene terephthalate, and polyimide) were contacted in sealed tubes containing the refrigerants and PAG lubricant, and held at 100°C for two weeks. The plastics were evaluated for changes in mass and appearance 24 h after test completion, finding essentially the same positive ratings of the two refrigerants with the plastics. For data an compatibility of plastics with ammonia, carbon dioxide, and hydrocarbons, see Pruett (2000).
Additional Compatibility Reports The Air-Conditioning, Heating, and Refrigeration Institute (AHRI) has supported research programs for refrigerants stability and materials compatibility through their Materials Compatibility and Lubricants Research (MCLR) Program, beginning in the early
29.11 1990s for replacements for CFCs. Resulting research reports are available from AHRI, with the following refrigerants stability/materials compatibility research topics: plastics (Cavestri 1993), lubricant additives (Cavestri 1997), system contaminants (Cavestri 2000), motor materials (Doerr and Kujak 1993; Doerr and Waite 1996), desiccants (Field 1995), elastomers (Hamed et al. 1994), and metals (Huttenlocher 1992). Cavestri et al. (2010) studied five refrigerants (R-417A, R-422D, R-424A, R-434A, and R-438A) and lubricants in contact with aluminum, copper, and steel coupons, and with nonmetallic materials of construction (elastomers, sealants, and plastics). They found, “A general, overall statement can be made that material changes for the R-22 alternative refrigerants investigated in this study do not have statistically significant differences compared to R-22 exposure in both time and temperature.” Two recent AHRI projects investigated materials compatibility of HFO refrigerants. Phase I covered thermal stability of 1234yf, 1234ze, and a 1234yf/R-32 blend with lubricants in the presence of metals (Rohatgi et al. 2012). Phase II focused on compatibility with plastics, elastomers, and motor materials (Majurin et al. 2014). Both reports show that HFOs have compatibility with a wide range of materials used in HVAC&R systems.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. ACR News. 2011a. Methyl chloride to blame for reefer explosions? (Nov. 6). ACR News. 2011b. Fake refrigerants becoming a “serious problem.” (Nov. 15). ACR News. 2012. Traces of dangerous refrigerant found in returned cylinders. (July 16). AHRI. 2011. Specification for fluorocarbon refrigerants. Standard 700-2011. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. ARAP. 1999. Global comparative analysis of HFC and alternative technologies for refrigeration, air conditioning, foam, solvent, aerosol propellant, and fire protection applications. Final Report to the Alliance for Responsible Atmospheric Policy, Arlington, VA. Dieckmann, J., Arthur D. Little, Inc., and Hillel Magid Consultant. unfccc.int/methods/other _methodological_issues/interactions_with_ozone_layer/items/378.php. ASHRAE. 2016. Designation and safety classification of refrigerants. ANSI/ ASHRAE Standard 34-2016. Bararo, M.T., U.V. Mardolcar, and C.A. Nieto de Castro. 1997. Molecular properties of alternative refrigerants derived from dielectric-constant measurements. Journal of Thermophysics 18(2):419-438. Beacham, E.A., and R.T. Divers. 1955. Some aspects of the dielectric properties of refrigerants. Refrigerating Engineering 7:33. Calm, J. C., G. C. Hourahan, A. Vonsild, D. Clodic, and D. Colbourne. 2015. 2014 Report of the refrigeration, air conditioning, and heat pumps technical options committee, Ch. 2: Refrigerants. United Nations Environment Programme (UNEP) Ozone Secretariat, Nairobi. ozone.unep.org /en/assessment-panels/technology-and-economic-assessment-panel. Cavestri, R.C. 1993. Compatibility of refrigerants and lubricants with engineering plastics. Report DOE/CE/23810-15. Air Conditioning and Refrigeration Technology Institute (ARTI), Arlington, VA. Cavestri, R.C. 1997. Compatibility of lubricant additives with HFC refrigerants and synthetic lubricants. Report DOE/CE/23810-76. Air Conditioning and Refrigeration Technology Institute (ARTI), Arlington, VA. Cavestri, R.C. 2000. Effect of selected contaminants in air conditioning and refrigeration equipment. Report DOE/CE/23810-111. Air Conditioning and Refrigeration Technology Institute (ARTI), Arlington, VA. Cavestri, R.C., M. El-Shazly, and D. Seeger-Clevenger. 2010. Thermal stability and chemical compatibility of R-22 replacement refrigerants. AHRI Project 8003. Air Conditioning, Heating, and Refrigeration Institute, Arlington, VA. Charlton, E.E., and F.S. Cooper. 1937. Dielectric strengths of insulating fluids. General Electric Review 865(9):438. CRC. 1987. CRC handbook of chemistry and physics, 68th ed. CRC Press, Boca Raton, FL. Doerr, R.G., and S.A. Kujak. 1993. Compatibility of refrigerants and lubricants with motor materials. Report DOE/CE/23810-13. Air Conditioning and Refrigeration Technology Institute, Arlington, VA.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
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2017 ASHRAE Handbook—Fundamentals (SI)
Doerr, R.G., and T.D. Waite. 1996. Compatibility of refrigerants and lubricants with motor materials under retrofit conditions. Report DOE/CE/ 23810-63. Air Conditioning and Refrigeration Technology Institute, Arlington, VA. Dow. 2007. Product safety assessment: Methyl chloride. Dow Chemical Company, Midland MI. Dow. 2010. Material safety data sheet: Methyl chloride. Dow Chemical Company, Midland, MI. Downing, R.C. 1988. Fluorocarbon refrigerants handbook. Prentice Hall, Englewood Cliffs, NJ. DuPont. Bulletin B-32A. Freon Products Division. E.I. DuPont de Nemours & Co., Wilmington, DE. DuPont Fluoroproducts. 2003. Technical Information Bulletins for HFC-134a, R-407C and R-410A. E.I. DuPont de Nemours & Co., Wilmington, DE. Eiseman, B.J., Jr. 1955. How electrical properties of Freon compounds affect hermetic system’s insulation. Refrigerating Engineering 4:61. European Industrial Gases Association. 2010. Gas compatibility with aluminum alloy cylinders. IGC Document 161/10/E. Brussels, Belgium. Fellows, B.R., R.G. Richard, and I.R. Shankland. 1991. Electrical characterization of alternate refrigerants. Actes Congrès International du Froid 18(2). International Institute of Refrigeration, Paris. Field, J.E. 1995. Sealed tube comparisons of the compatibility of desiccants with refrigerants and lubricants. Report DOE/CE/23810-54. Air Conditioning and Refrigeration Technology Institute, Arlington, VA. Fischer, S.K., P.J. Hughes, P.D. Fairchild, C.L. Kusik, J.T. Dieckmann, E.M. McMahon, and N. Hobay. 1991. Energy and global warming impacts of CFC alternative technologies. Sponsored by the Alternative Fluorocarbons Environmental Acceptability Study (AFEAS) and the U.S. Department of Energy (DOE). www.ciesin.org/docs/011-459/011-459.html. Fuoss, R.M. 1938. Dielectric constants of some fluorine compounds. Journal of the American Chemical Society 64:1633. Gbur, A.M., and J.J. Byrne. 2001. Determination of dieletric properties of refrigerants. ASHRAE Research Project RP-1074, Final Report. Hamed, G.R., R.H. Seiple, and O. Taikum. 1994. Compatibility of refrigerants and lubricants with elastomers. Report DOE/CE/23810-14. Air Conditioning and Refrigeration Technology Institute, Arlington, VA. Handbook of chemistry, 10th ed. 1967. McGraw-Hill, New York. Hill, W., and U. Grimm. 2008. Overview of SAE Cooperative Research Program CRP1234-2 for alternative refrigerants. SAE International AARS. Phoenix. AZ. Huttenlocher, D.F. 1992. Chemical and thermal stability of refrigerant-lubricant mixtures with metals. Report DOE/CE/23810-5. Air Conditioning and Refrigeration Technology Institute, Arlington, VA. IPCC. 2007. Climate change 2007: The physical science basis. Contribution of working group I to the fourth assessment report of the Intergovernmental Panel on Climate Change. S. Solomon, D. Qin, M. Manning, Z. Chen, M. Marquis, K.B. Averyt, M. Tignor, and H.L. Miller, eds. Cambridge University Press, Cambridge, U.K. ipcc.ch/publications_and _data/ar4/wg1/en/contents.html. IPCC. 2013. Climate change 2013: The physical science basis. Contribution of working group I to the fifth assessment report of the Intergovernmental Panel on Climate Change. T.F. Stoecker, D. Qin, G.-K. Plattner, M. Tignor, S.K. Allen, J. Boschung, A. Nauels, Y. Xia, V. Bex, and P.M. Midgely, eds. Cambridge University Press. IPCC. 2014. Climate change 2014: Synthesis report. Contribution of working groups I, II, and III to the fifth assessment report of the Intergovernmental Panel on Climate Change. Core writing team, R.K. Pachauri, and L.A. Meyer, eds. International Panel on Climate Change, Geneva. www .ipcc.ch/report/ar5/syr/. Kirk and Othmer. 1956. The encyclopedia of chemical technology. Interscience Encyclopedia, New York. Linde. 2010. Material safety data sheet: Methyl chloride. Linde Gas North America LLC, Murray Hill, NJ. Majurin, J. A., E. Sorenson, S. J. Staats, W. Gilles and S. J. Kujak. 2014. Material compatibility and lubricants research for low GWP refrigerant—Phase II: Chemical and material compatibility of low GWP refrig-
erants with HVACR material of construction. AHRI Project 08007. www.ahrinet.org/App_Content/ahri/files/RESEARCH/Technical%20 Results/AHRI_Project-8007_Final_Report.pdf. Meurer, C., G. Pietsch, and M. Haacke. 2001. Electrical properties of CFC- and HCFC-substitutes. International Journal of Refrigeration 24(2):171-175. Minor, B., and M. Spatz. 2008. HFO-1234yf low GWP refrigerant update. Presentation charts. International Refrigeration and Air Conditioning Conference, Purdue University, West Lafayette, IN. www2.dupont.com /Refrigerants/en_US/assets/downloads/SmartAutoAC/MAC_Purdue _HFO_1234yf.pdf. Muller, Y., S. Feja, and U. Grimm. 2011. Electrical properties of the liquid phase of refrigerant oil mixtures. SAE International Automotive Refrigerant and System Efficiency Symposium, Scottsdale, AZ. NIST. 2010. NIST reference fluid thermodynamic and transport properties database (REFPROP), v. 9.0. National Institute of Standards and Technology, Gaithersburg, MD. OxyChem. 2009. Reaction of methyl chloride with aluminum. OxyChem Technical Data Sheet 510-101. Wichita, KS. Pereira, L.F., F.E. Brito, A.N. Gurova, U.V. Mardolcar, and C.A. Nieta de Castro. 1999. Dipole moment, expansivity and compressibility coefficients of HFC 125 derived from dielectric constant measurements. 1st International Workshop on Thermochemical, Thermodynamic and Transport Properties of Halogenated Hydrocarbons and Mixtures, Pisa, Italy. Powell, P. 2012. Rogue refrigerant blend linked to fatalities. Air Conditioning, Heating, and Refrigeration News (January 9). Troy, MI. www .achrnews.com/articles/rogue-refrigerant-blend-linked-to-fatalities. Pruett, K.M. 1994. Chemical resistance guide for elastomers II. Compass Publications, La Mesa, CA. Pruett, K.M. 1995. Chemical resistance guide for metals and alloys. Compass Publications, La Mesa, CA. Pruett, K.M. 2000. Chemical resistance guide for plastics. Compass Publications, La Mesa, CA. Ravishankara, A.R., A.A. Turnipseed, N.R. Jensen, and R.F. Warren. 1994. Do hydrofluorocarbons destroy stratospheric ozone? Science 248:1217-1219. Rohatgi, N.D., R.W. Clark, D.R. Hurst. 2012. Material compatibility & lubricants research for low GWP refrigerants—Phase I: Thermal and chemical stability of low GWP refrigerants with lubricants. Project 09004-01. www.ahrinet.org/site/511/Resources/Research/Public-Sector -Research/Technical-Results. UNEP. 2009. Handbook for the international treaties for the protection of the ozone layer, 8th ed. United Nations Environment Programme (UNEP) Ozone Secretariat, Nairobi, Kenya. ozone.unep.org/. WorldCargo News. 2011. Alarm sounded over exploding reefers. (October 26). WCN Publishing. Leatherhead, Surrey, UK. www.worldcargonews .com/htm/w20111026.937700.htm.
BIBLIOGRAPHY Calm, J.M., and G.C. Hourahan. 2011. 2010 Report of the refrigeration, air conditioning and heat pumps technical options committee, Chapter 2: Refrigerants. United Nations Environment Programme (UNEP) Ozone Secretariat, Nairobi, Kenya. ozone.unep.org/teap/Reports/RTOC/RTOC -Assessment-report-2010.pdf. Calm, J.M., and G.C. Hourahan. 2011. Physical, safety, and environmental data for current and alternative refrigerants. Refrigeration for Sustainable Development: Proceedings of the 23rd International Congress of Refrigeration, Paper 915. International Institute of Refrigeration, Paris. IPCC. 2007. Climate change 2007: Synthesis report. Contribution of working groups I, II, and III to the fourth assessment report of the Intergovernmental Panel on Climate Change. Core Writing Team, R.K. Pachauri, and A. Reisinger, eds. International Panel on Climate Change, Geneva. www.ipcc.ch/publications_and_data/ar4/syr/en/contents.html. Lemmon, E.W., M.O. McLinden, and M.L. Huber. 2002. NIST standard reference database 23, v. 9.0. National Institute of Standards and Technology, Gaithersburg, MD. Matheson gas data book. 2001. Matheson Company, East Rutherford, NJ.
Related Commercial Resources
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Related Commercial Resources CHAPTER 30
THERMOPHYSICAL PROPERTIES OF REFRIGERANTS HIS chapter presents data for thermodynamic and transport properties of refrigerants, arranged for the occasional user. The refrigerants have a thermodynamic property chart on pressureenthalpy coordinates with an abbreviated set of tabular data for saturated liquid and vapor on the facing page. In addition, tabular data in the superheated vapor region are given for R-134a to assist students working on compression cycle examples. For each cryogenic fluid, a second table of properties is provided for vapor at a pressure of one standard atmosphere; these data are needed when such gases are used in heat transfer or purge gas applications. For zeotropic blends, including R-729 (air), tables are incremented in pressure, with properties given for liquid on the bubble line and vapor on the dew line. This arrangement is used because pressure is more commonly measured in the field while servicing equipment; it also highlights the difference between bubble and dew-point temperatures (the “temperature glide” experienced with blends). Most CFC refrigerants have been deleted. Tables for R-11, R-13, R-113, R-114, R-141b, R-142b, R-500, R-502, R-503, and R-720
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T
Refrigerant
(neon) may be found in the 1997 ASHRAE Handbook—Fundamentals. R-12 has been retained to assist in making comparisons. Hydrogen and parahydrogen (R-702 and R-702p) may be found in the 2009 ASHRAE Handbook–Fundamentals. A new table and diagram for R-1233zd(E) have been added, and the data for R-245fa and R-1234ze(E) have been revised. Viscosity data for R-1234yf have been revised. The formulations conform to international standards, where applicable: thermodynamic properties of R-12, R-22, R-32, R-123, R-125, R-134a, R-143a, R-152a, R-717 (ammonia), and R-744 (carbon dioxide) and refrigerant blends R-404A, R-407C, R-410A, and R-507 conform to ISO Standard 17584, Refrigerant Properties. Reference states used for most refrigerants correspond to the international convention of 200 kJ/kg for enthalpy and 1 kJ/(kg·K) for entropy, both for saturated liquid at 0°C. Exceptions are water and fluids with very low critical temperatures (e.g., ethylene, cryogens). These data are intended to help engineers make preliminary comparisons among unfamiliar fluids. For greater detail and a wider range of data, see the sources in the References.
Page
Halocarbon Refrigerants
Refrigerant
Page
Inorganic Refrigerants R-717 (ammonia) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.40
Methane Series R-12 (dichlorodifluoromethane). . . . . . . . . . . . . . . . . . .
30.2
R-718 (water/steam) . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.42
R-22 (chlorodifluoromethane) . . . . . . . . . . . . . . . . . . . .
30.4
R-744 (carbon dioxide) . . . . . . . . . . . . . . . . . . . . . . . . . . 30.44
R-23 (trifluoromethane) . . . . . . . . . . . . . . . . . . . . . . . . .
30.6
R-32 (difluoromethane) . . . . . . . . . . . . . . . . . . . . . . . . .
30.8
Ethane Series R-123 (2,2-dichloro-1,1,1-trifluoroethane) . . . . . . . . . . 30.10 R-124 (2-chloro-1,1,1,2-tetrafluoroethane) . . . . . . . . . . 30.12 R-125 (pentafluoroethane) . . . . . . . . . . . . . . . . . . . . . . . 30.14 R-134a (1,1,1,2-tetrafluoroethane) . . . . . . . . . . . . . . . . . 30.16 R-143a (1,1,1-trifluoroethane) . . . . . . . . . . . . . . . . . . . . 30.20 R-152a (1,1-difluoroethane) . . . . . . . . . . . . . . . . . . . . . . 30.22
Hydrocarbon Refrigerants R-50 (methane) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.46 R-170 (ethane). . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.48 R-290 (propane) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.50 R-600 (n-butane) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.52 R-600a (isobutane) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.54 R-1150 (ethylene) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.56 R-1270 (propylene) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.58
Propane Series R-245fa (1,1,1,3,3-pentafluoropropane). . . . . . . . . . . . . 30.24 Cryogenic Fluids
Propylene Series R-1233zd(E) (trans-1-chloro-3,3,3-trifluoroprop-1-ene) 30.26 R-1234yf (2,3,3,3-tetrafluoroprop-1-ene) . . . . . . . . . . . 30.28 R-1234ze(E) (trans-1,3,3,3-tetrafluoropropene). . . . . . . 30.30 Zeotropic Blends (% by mass) R-404A [R-125/143a/134a (44/52/4)] . . . . . . . . . . . . . . 30.32
R-704 (helium) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.60 R-728 (nitrogen) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.62 R-729 (air). . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.64 R-732 (oxygen) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.66 R-740 (argon) . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.68
R-407C [R-32/125/134a (23/25/52)] . . . . . . . . . . . . . . . 30.34 R-410A [R-32/125 (50/50)] . . . . . . . . . . . . . . . . . . . . . . 30.36
Absorption Solutions Ammonia/Water . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30.70
Azeotropic Blends R-507A [R-125/143a (50/50)] . . . . . . . . . . . . . . . . . . . . 30.38
Water/Lithium Bromide . . . . . . . . . . . . . . . . . . . . . . . . . 30.72
The preparation of this chapter is assigned to TC 3.1, Refrigerants and Secondary Coolants.
30.1 Copyright © 2017, ASHRAE
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2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 1
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 12
30.2
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.3
Refrigerant 12 (Dichlorodifluoromethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor
–100 0.00119 1679.1 10.0040 –90 0.00286 1652.8 4.3948 –80 0.00619 1626.3 2.1355 –70 0.01228 1599.5 1.1286 –60 0.02261 1572.3 0.63992 –50 0.03911 1544.7 0.38494 –40 0.06409 1516.5 0.24342 –30 0.10026 1487.7 0.16057 –29.75b 0.10133 1487.0 0.15900 –28 0.10910 1481.9 0.14841 –26 0.11854 1476.0 0.13736 –24 0.12860 1470.1 0.12731 –22 0.13931 1464.1 0.11815 –20 0.15070 1458.1 0.10978 –18 0.16279 1452.1 0.10213 –16 0.17562 1446.1 0.09512 –14 0.18920 1440.0 0.08870 –12 0.20358 1433.8 0.08280 –10 0.21878 1427.6 0.07737 –8 0.23483 1421.4 0.07237 –6 0.25176 1415.1 0.06777 –4 0.26960 1408.8 0.06352 –2 0.28839 1402.5 0.05959 0 0.30815 1396.1 0.05595 2 0.32891 1389.6 0.05258 4 0.35071 1383.1 0.04946 6 0.37358 1376.5 0.04656 8 0.39756 1369.9 0.04386 10 0.42267 1363.2 0.04135 12 0.44895 1356.5 0.03901 14 0.47643 1349.7 0.03683 16 0.50514 1342.8 0.03480 18 0.53513 1335.9 0.03290 20 0.56642 1328.9 0.03112 22 0.59905 1321.8 0.02946 24 0.63305 1314.6 0.02790 26 0.66846 1307.4 0.02643 28 0.70531 1300.1 0.02506 30 0.74365 1292.7 0.02377 32 0.78350 1285.2 0.02256 34 0.82491 1277.6 0.02142 36 0.86791 1269.9 0.02034 38 0.91253 1262.2 0.01933 40 0.95882 1254.3 0.01838 42 1.00680 1246.3 0.01748 44 1.05660 1238.1 0.01662 46 1.10810 1229.9 0.01582 48 1.16140 1221.5 0.01505 50 1.21660 1213.0 0.01433 52 1.27370 1204.4 0.01365 54 1.33270 1195.6 0.01300 56 1.39380 1186.6 0.01238 58 1.45680 1177.5 0.01180 60 1.52190 1168.1 0.01124 62 1.58920 1158.6 0.01071 64 1.65860 1148.9 0.01021 66 1.73020 1139.0 0.00973 68 1.80410 1128.8 0.00927 70 1.88020 1118.3 0.00883 75 2.08110 1090.9 0.00782 80 2.29750 1061.4 0.00691 85 2.53040 1029.1 0.00608 90 2.78080 993.2 0.00533 95 3.05010 952.2 0.00463 100 3.33990 903.8 0.00396 105 3.65250 842.2 0.00330 110 3.99240 742.7 0.00252 111.97c 4.13610 565.0 0.00177 *Temperatures on ITS-90 scale
113.32 121.53 129.81 138.17 146.62 155.18 163.86 172.67 172.89 174.44 176.23 178.02 179.81 181.62 183.42 185.24 187.06 188.89 190.72 192.56 194.41 196.27 198.13 200.00 201.88 203.76 205.65 207.56 209.46 211.38 213.31 215.24 217.18 219.14 221.10 223.07 225.05 227.04 229.04 231.06 233.08 235.12 237.16 239.22 241.29 243.38 245.47 247.59 249.71 251.85 254.01 256.18 258.38 260.58 262.81 265.06 267.33 269.62 271.94 277.84 283.94 290.27 296.91 303.95 311.58 320.24 331.82 347.76
306.09 310.59 315.19 319.87 324.61 329.39 334.18 338.94 339.06 339.89 340.83 341.78 342.72 343.65 344.59 345.52 346.44 347.37 348.29 349.20 350.11 351.01 351.91 352.81 353.69 354.57 355.45 356.32 357.18 358.03 358.88 359.71 360.54 361.36 362.17 362.97 363.76 364.54 365.31 366.07 366.81 367.54 368.26 368.96 369.65 370.33 370.98 371.62 372.24 372.85 373.43 373.99 374.53 375.05 375.54 376.00 376.44 376.84 377.22 377.99 378.48 378.64 378.35 377.45 375.60 372.08 363.95 347.76
Velocity of Sound, m/s cp /cv Vapor Liquid Vapor
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
0.6077 0.6538 0.6978 0.7400 0.7806 0.8197 0.8577 0.8946 0.8955 0.9019 0.9091 0.9163 0.9234 0.9305 0.9376 0.9447 0.9517 0.9587 0.9656 0.9726 0.9795 0.9863 0.9932 1.0000 1.0068 1.0136 1.0203 1.0270 1.0337 1.0404 1.0471 1.0537 1.0603 1.0669 1.0735 1.0801 1.0866 1.0932 1.0997 1.1062 1.1127 1.1192 1.1257 1.1322 1.1387 1.1451 1.1516 1.1580 1.1645 1.1710 1.1774 1.1839 1.1904 1.1969 1.2033 1.2099 1.2164 1.2229 1.2295 1.2461 1.2629 1.2801 1.2978 1.3163 1.3360 1.3581 1.3874 1.4283
0.819 0.449 1.182 1035 0.824 0.465 1.176 990 0.831 0.481 1.172 945 0.840 0.497 1.168 902 0.850 0.513 1.166 859 0.861 0.530 1.165 816 0.873 0.548 1.166 775 0.886 0.566 1.169 733 0.887 0.567 1.169 732 0.889 0.570 1.170 725 0.892 0.574 1.171 717 0.895 0.578 1.171 709 0.898 0.582 1.172 701 0.901 0.586 1.174 693 0.904 0.590 1.175 684 0.907 0.594 1.176 676 0.910 0.598 1.178 668 0.913 0.602 1.179 660 0.917 0.607 1.181 652 0.920 0.611 1.183 644 0.923 0.616 1.184 636 0.927 0.620 1.186 628 0.930 0.625 1.189 620 0.934 0.630 1.191 612 0.938 0.635 1.193 604 0.942 0.640 1.196 596 0.946 0.645 1.199 588 0.950 0.650 1.202 580 0.954 0.656 1.205 572 0.958 0.661 1.208 564 0.962 0.667 1.211 556 0.967 0.672 1.215 548 0.971 0.678 1.219 540 0.976 0.685 1.223 532 0.981 0.691 1.228 524 0.986 0.697 1.232 516 0.991 0.704 1.237 508 0.997 0.711 1.242 499 1.002 0.718 1.248 491 1.008 0.726 1.254 483 1.014 0.734 1.260 475 1.020 0.742 1.267 467 1.026 0.750 1.274 459 1.033 0.759 1.282 450 1.040 0.768 1.290 442 1.048 0.778 1.299 434 1.055 0.788 1.308 426 1.063 0.798 1.318 417 1.072 0.810 1.329 409 1.081 0.821 1.340 400 1.090 0.834 1.353 392 1.100 0.847 1.366 383 1.111 0.861 1.381 375 1.122 0.876 1.397 366 1.135 0.892 1.414 357 1.148 0.910 1.433 348 1.162 0.929 1.453 339 1.177 0.949 1.476 330 1.193 0.971 1.501 321 1.241 1.037 1.576 298 1.302 1.122 1.677 274 1.384 1.239 1.817 249 1.501 1.410 2.026 224 1.679 1.683 2.362 197 1.996 2.192 2.990 169 2.754 3.458 4.544 139 7.81 11.440 14.140 105 0 b Normal boiling point
1.7210 1.6861 1.6576 1.6344 1.6156 1.6004 1.5882 1.5784 1.5782 1.5767 1.5751 1.5735 1.5720 1.5706 1.5693 1.5680 1.5667 1.5655 1.5644 1.5633 1.5623 1.5613 1.5603 1.5594 1.5586 1.5577 1.5569 1.5561 1.5554 1.5547 1.5540 1.5533 1.5527 1.5521 1.5515 1.5509 1.5503 1.5498 1.5492 1.5487 1.5481 1.5476 1.5470 1.5465 1.5459 1.5454 1.5448 1.5443 1.5437 1.5431 1.5425 1.5418 1.5411 1.5404 1.5397 1.5389 1.5381 1.5372 1.5363 1.5337 1.5306 1.5268 1.5220 1.5159 1.5076 1.4952 1.4712 1.4283
118.5 121.4 124.1 126.7 129.1 131.2 133.0 134.5 134.5 134.7 135.0 135.2 135.4 135.6 135.8 136.0 136.1 136.3 136.4 136.5 136.6 136.6 136.7 136.7 136.7 136.7 136.7 136.6 136.5 136.5 136.3 136.2 136.1 135.9 135.7 135.5 135.2 134.9 134.7 134.3 134.0 133.6 133.2 132.8 132.4 131.9 131.4 130.9 130.3 129.7 129.1 128.4 127.7 127.0 126.3 125.5 124.6 123.8 122.9 120.4 117.7 114.7 111.4 107.8 103.7 99.3 94.0 0.0
Liquid Vapor
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C
1005.0 819.0 684.9 584.0 505.1 441.8 389.8 346.2 345.2 338.3 330.6 323.2 316.0 309.0 302.2 295.6 289.2 283.0 276.9 271.0 265.2 259.6 254.1 248.7 243.5 238.4 233.4 228.6 223.8 219.1 214.6 210.1 205.7 201.4 197.2 193.1 189.0 185.0 181.1 177.3 173.5 169.8 166.1 162.5 159.0 155.5 152.0 148.6 145.3 141.9 138.7 135.4 132.2 129.1 125.9 122.8 119.7 116.7 113.6 106.1 98.6 91.1 83.4 75.6 67.3 58.1 46.3 —
116.7 112.0 107.4 103.0 98.8 94.7 90.7 86.9 86.8 86.1 85.3 84.6 83.8 83.1 82.4 81.6 80.9 80.2 79.4 78.7 78.0 77.3 76.6 75.9 75.1 74.4 73.7 73.0 72.3 71.6 70.9 70.2 69.6 68.9 68.2 67.5 66.8 66.1 65.4 64.8 64.1 63.4 62.7 62.1 61.4 60.7 60.0 59.4 58.7 58.0 57.3 56.7 56.0 55.3 54.7 54.0 53.3 52.6 52.0 50.3 48.7 47.2 45.9 45.1 45.7 51.7 113.7
Viscosity, Pa·s 6.78 7.18 7.58 7.97 8.37 8.76 9.16 9.55 9.56 9.63 9.71 9.79 9.87 9.95 10.03 10.10 10.18 10.26 10.34 10.42 10.50 10.58 10.66 10.74 10.82 10.90 10.98 11.07 11.15 11.23 11.31 11.40 11.48 11.57 11.65 11.74 11.83 11.92 12.01 12.10 12.19 12.28 12.38 12.48 12.57 12.67 12.78 12.88 12.99 13.10 13.21 13.33 13.45 13.57 13.70 13.83 13.96 14.11 14.26 14.66 15.11 15.65 16.29 17.11 18.20 19.87 23.46 —
4.27 4.67 5.08 5.50 5.93 6.38 6.84 7.32 7.33 7.41 7.51 7.61 7.71 7.80 7.90 8.01 8.11 8.21 8.31 8.41 8.52 8.62 8.73 8.84 8.95 9.06 9.17 9.28 9.39 9.51 9.62 9.74 9.86 9.98 10.10 10.23 10.36 10.49 10.62 10.75 10.89 11.03 11.18 11.33 11.48 11.63 11.79 11.96 12.13 12.31 12.49 12.68 12.87 13.08 13.29 13.51 13.75 13.99 14.25 14.96 15.80 16.82 18.11 19.81 22.27 26.34 39.46
26.48 –100 24.90 –90 23.35 –80 21.81 –70 20.30 –60 18.81 –50 17.35 –40 15.91 –30 15.88 –29.75 15.63 –28 15.35 –26 15.06 –24 14.78 –22 14.50 –20 14.23 –18 13.95 –16 13.67 –14 13.40 –12 13.12 –10 12.85 –8 12.58 –6 12.31 –4 12.04 –2 11.77 0 11.51 2 11.24 4 10.98 6 10.71 8 10.45 10 10.19 12 9.94 14 9.68 16 9.42 18 9.17 20 8.92 22 8.67 24 8.42 26 8.17 28 7.92 30 7.68 32 7.43 34 7.19 36 6.95 38 6.72 40 6.48 42 6.25 44 6.01 46 5.78 48 5.55 50 5.33 52 5.10 54 4.88 56 4.66 58 4.44 60 4.23 62 4.01 64 3.80 66 3.59 68 3.39 70 2.88 75 2.40 80 1.93 85 1.49 90 1.07 95 0.69 100 0.35 105 0.07 110 0.00 111.97 c Critical point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 2
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 22
30.4
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.5
Refrigerant 22 (Chlorodifluoromethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –100 –90 –80 –70 –60 –50 –48 –46 –44 –42 –40.81b –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 65 70 75 80 85 90 95 96.15c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00201 0.00481 0.01037 0.02047 0.03750 0.06453 0.07145 0.07894 0.08705 0.09580 0.10132 0.10523 0.11538 0.12628 0.13797 0.15050 0.16389 0.17819 0.19344 0.20968 0.22696 0.24531 0.26479 0.28543 0.30728 0.33038 0.35479 0.38054 0.40769 0.43628 0.46636 0.49799 0.53120 0.56605 0.60259 0.64088 0.68095 0.72286 0.76668 0.81244 0.86020 0.91002 0.96195 1.01600 1.07240 1.13090 1.19190 1.25520 1.32100 1.38920 1.46010 1.53360 1.60980 1.68870 1.77040 1.85510 1.94270 2.03330 2.12700 2.22390 2.32400 2.42750 2.70120 2.99740 3.31770 3.66380 4.03780 4.44230 4.88240 4.99000
1571.3 1544.9 1518.2 1491.2 1463.7 1435.6 1429.9 1424.2 1418.4 1412.6 1409.2 1406.8 1401.0 1395.1 1389.1 1383.2 1377.2 1371.1 1365.0 1358.9 1352.7 1346.5 1340.3 1334.0 1327.6 1321.2 1314.7 1308.2 1301.6 1295.0 1288.3 1281.5 1274.7 1267.8 1260.8 1253.8 1246.7 1239.5 1232.2 1224.9 1217.4 1209.9 1202.3 1194.6 1186.7 1178.8 1170.7 1162.6 1154.3 1145.8 1137.3 1128.5 1119.6 1110.6 1101.4 1091.9 1082.3 1072.4 1062.3 1052.0 1041.3 1030.4 1001.4 969.7 934.4 893.7 844.8 780.1 662.9 523.8
8.26600 3.64480 1.77820 0.94342 0.53680 0.32385 0.29453 0.26837 0.24498 0.22402 0.21260 0.20521 0.18829 0.17304 0.15927 0.14682 0.13553 0.12528 0.11597 0.10749 0.09975 0.09268 0.08621 0.08029 0.07485 0.06986 0.06527 0.06103 0.05713 0.05352 0.05019 0.04710 0.04424 0.04159 0.03913 0.03683 0.03470 0.03271 0.03086 0.02912 0.02750 0.02599 0.02457 0.02324 0.02199 0.02082 0.01972 0.01869 0.01771 0.01679 0.01593 0.01511 0.01433 0.01360 0.01291 0.01226 0.01163 0.01104 0.01048 0.00995 0.00944 0.00896 0.00785 0.00685 0.00595 0.00512 0.00434 0.00356 0.00262 0.00191
*Temperatures on ITS-90 scale
90.71 101.32 111.94 122.58 133.27 144.03 146.19 148.36 150.53 152.70 154.00 154.89 157.07 159.27 161.47 163.67 165.88 168.10 170.33 172.56 174.80 177.04 179.30 181.56 183.83 186.11 188.40 190.70 193.01 195.33 197.66 200.00 202.35 204.71 207.09 209.47 211.87 214.28 216.70 219.14 221.59 224.06 226.54 229.04 231.55 234.08 236.62 239.19 241.77 244.38 247.00 249.65 252.32 255.01 257.73 260.47 263.25 266.05 268.89 271.76 274.66 277.61 285.18 293.10 301.46 310.44 320.38 332.09 349.56 366.90
358.97 363.85 368.77 373.70 378.59 383.42 384.37 385.32 386.26 387.20 387.75 388.13 389.06 389.97 390.89 391.79 392.69 393.58 394.47 395.34 396.21 397.06 397.91 398.75 399.57 400.39 401.20 401.99 402.77 403.55 404.30 405.05 405.78 406.50 407.20 407.89 408.56 409.21 409.85 410.47 411.07 411.66 412.22 412.77 413.29 413.79 414.26 414.71 415.14 415.54 415.91 416.25 416.55 416.83 417.07 417.27 417.44 417.56 417.63 417.66 417.63 417.55 417.06 416.09 414.49 412.01 408.19 401.87 387.28 366.90
Entropy, kJ/(kg·K) Liquid
Vapor
0.5050 0.5646 0.6210 0.6747 0.7260 0.7752 0.7849 0.7944 0.8039 0.8134 0.8189 0.8227 0.8320 0.8413 0.8505 0.8596 0.8687 0.8778 0.8868 0.8957 0.9046 0.9135 0.9223 0.9311 0.9398 0.9485 0.9572 0.9658 0.9744 0.9830 0.9915 1.0000 1.0085 1.0169 1.0254 1.0338 1.0422 1.0505 1.0589 1.0672 1.0755 1.0838 1.0921 1.1004 1.1086 1.1169 1.1252 1.1334 1.1417 1.1499 1.1582 1.1665 1.1747 1.1830 1.1913 1.1997 1.2080 1.2164 1.2248 1.2333 1.2418 1.2504 1.2722 1.2945 1.3177 1.3423 1.3690 1.4001 1.4462 1.4927
2.0543 1.9980 1.9508 1.9108 1.8770 1.8480 1.8428 1.8376 1.8327 1.8278 1.8250 1.8231 1.8186 1.8141 1.8098 1.8056 1.8015 1.7975 1.7937 1.7899 1.7862 1.7826 1.7791 1.7757 1.7723 1.7690 1.7658 1.7627 1.7596 1.7566 1.7536 1.7507 1.7478 1.7450 1.7422 1.7395 1.7368 1.7341 1.7315 1.7289 1.7263 1.7238 1.7212 1.7187 1.7162 1.7136 1.7111 1.7086 1.7061 1.7036 1.7010 1.6985 1.6959 1.6933 1.6906 1.6879 1.6852 1.6824 1.6795 1.6766 1.6736 1.6705 1.6622 1.6529 1.6424 1.6299 1.6142 1.5922 1.5486 1.4927
Velocity of Viscosity, Pa·s Sound, m/s cp /cv Liquid Vapor Vapor Liquid Vapor Liquid Vapor Specific Heat cp , kJ/(kg·K)
1.061 1.061 1.062 1.065 1.071 1.079 1.081 1.083 1.086 1.088 1.090 1.091 1.093 1.096 1.099 1.102 1.105 1.108 1.112 1.115 1.119 1.123 1.127 1.131 1.135 1.139 1.144 1.149 1.154 1.159 1.164 1.169 1.175 1.181 1.187 1.193 1.199 1.206 1.213 1.220 1.228 1.236 1.244 1.252 1.261 1.271 1.281 1.291 1.302 1.314 1.326 1.339 1.353 1.368 1.384 1.401 1.419 1.439 1.461 1.485 1.511 1.539 1.626 1.743 1.913 2.181 2.682 3.981 17.31
0.497 0.512 0.528 0.545 0.564 0.585 0.589 0.594 0.599 0.603 0.606 0.608 0.613 0.619 0.624 0.629 0.635 0.641 0.646 0.653 0.659 0.665 0.672 0.678 0.685 0.692 0.699 0.707 0.715 0.722 0.731 0.739 0.748 0.757 0.766 0.775 0.785 0.795 0.806 0.817 0.828 0.840 0.853 0.866 0.879 0.893 0.908 0.924 0.940 0.957 0.976 0.995 1.015 1.037 1.061 1.086 1.113 1.142 1.173 1.208 1.246 1.287 1.413 1.584 1.832 2.231 2.984 4.975 25.29
b Normal
1.243 1.237 1.233 1.231 1.230 1.232 1.233 1.234 1.235 1.236 1.236 1.237 1.238 1.239 1.241 1.242 1.244 1.246 1.248 1.250 1.253 1.255 1.258 1.261 1.264 1.267 1.270 1.274 1.278 1.282 1.287 1.291 1.296 1.301 1.307 1.313 1.319 1.326 1.333 1.340 1.348 1.357 1.366 1.375 1.385 1.396 1.408 1.420 1.434 1.448 1.463 1.480 1.498 1.517 1.538 1.561 1.586 1.614 1.644 1.677 1.714 1.755 1.881 2.056 2.315 2.735 3.532 5.626 26.43
boiling point
1127 1080 1033 986 940 893 884 875 865 856 851 847 838 828 819 810 800 791 782 772 763 754 744 735 726 716 707 697 688 679 669 660 650 641 632 622 613 603 594 584 575 565 555 546 536 527 517 507 497 487 478 468 458 448 437 427 417 407 396 386 375 364 337 309 280 249 215 177 128 0
143.6 147.0 150.3 153.3 156.0 158.3 158.7 159.1 159.5 159.9 160.1 160.3 160.6 160.9 161.2 161.5 161.8 162.0 162.3 162.5 162.7 162.8 163.0 163.1 163.2 163.3 163.3 163.4 163.4 163.4 163.4 163.3 163.2 163.1 163.0 162.8 162.6 162.4 162.2 161.9 161.6 161.3 161.0 160.6 160.2 159.7 159.2 158.7 158.2 157.6 157.0 156.4 155.7 155.0 154.2 153.4 152.6 151.7 150.8 149.8 148.8 147.7 144.9 141.7 138.1 134.2 129.7 124.6 118.0 0.0
845.8 699.4 591.0 507.6 441.4 387.5 377.8 368.6 359.6 351.0 346.0 342.6 334.5 326.7 319.1 311.7 304.6 297.7 291.0 284.4 278.1 271.9 265.9 260.1 254.4 248.8 243.4 238.1 233.0 227.9 223.0 218.2 213.5 208.9 204.4 200.0 195.7 191.5 187.3 183.2 179.2 175.3 171.5 167.7 163.9 160.3 156.7 153.1 149.6 146.1 142.7 139.4 136.1 132.8 129.5 126.3 123.1 120.0 116.9 113.8 110.7 107.6 100.0 92.4 84.6 76.6 68.1 58.3 44.4 —
7.25 7.67 8.09 8.52 8.94 9.36 9.45 9.53 9.62 9.70 9.75 9.79 9.87 9.96 10.04 10.12 10.21 10.29 10.38 10.46 10.55 10.63 10.72 10.80 10.89 10.98 11.06 11.15 11.24 11.32 11.41 11.50 11.59 11.68 11.77 11.86 11.96 12.05 12.14 12.24 12.33 12.43 12.53 12.63 12.74 12.84 12.95 13.06 13.17 13.28 13.40 13.52 13.64 13.77 13.90 14.04 14.18 14.32 14.47 14.63 14.80 14.98 15.46 16.02 16.70 17.55 18.71 20.48 24.76 —
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 143.1 137.8 132.6 127.6 122.6 117.8 116.9 115.9 115.0 114.0 113.5 113.1 112.2 111.2 110.3 109.4 108.5 107.5 106.6 105.7 104.8 103.9 103.0 102.1 101.1 100.2 99.3 98.4 97.5 96.6 95.7 94.8 93.9 93.1 92.2 91.3 90.4 89.5 88.6 87.7 86.8 85.9 85.0 84.1 83.2 82.3 81.4 80.5 79.6 78.7 77.8 76.9 76.0 75.1 74.1 73.2 72.3 71.4 70.4 69.5 68.6 67.6 65.3 62.9 60.6 58.6 57.4 59.3 83.5
4.46 4.84 5.25 5.68 6.12 6.59 6.69 6.79 6.89 6.99 7.05 7.09 7.19 7.29 7.40 7.51 7.61 7.72 7.83 7.94 8.06 8.17 8.29 8.40 8.52 8.65 8.77 8.89 9.02 9.15 9.28 9.42 9.56 9.70 9.84 9.99 10.14 10.29 10.45 10.61 10.77 10.95 11.12 11.30 11.49 11.69 11.89 12.10 12.31 12.54 12.77 13.02 13.28 13.55 13.83 14.13 14.45 14.78 15.14 15.52 15.92 16.36 17.61 19.16 21.16 23.87 27.82 34.55 59.15
28.12 –100 26.36 –90 24.63 –80 22.92 –70 21.24 –60 19.58 –50 19.25 –48 18.92 –46 18.59 –44 18.27 –42 18.08 –40.81 17.94 –40 17.62 –38 17.30 –36 16.98 –34 16.66 –32 16.34 –30 16.02 –28 15.70 –26 15.39 –24 15.07 –22 14.76 –20 14.45 –18 14.14 –16 13.83 –14 13.52 –12 13.21 –10 12.91 –8 12.60 –6 12.30 –4 12.00 –2 11.70 0 11.40 2 11.10 4 10.81 6 10.51 8 10.22 10 9.93 12 9.64 14 9.35 16 9.06 18 8.78 20 8.50 22 8.22 24 7.94 26 7.66 28 7.38 30 7.11 32 6.84 34 6.57 36 6.30 38 6.04 40 5.77 42 5.51 44 5.25 46 5.00 48 4.74 50 4.49 52 4.24 54 4.00 56 3.75 58 3.51 60 2.92 65 2.36 70 1.82 75 1.30 80 0.83 85 0.40 90 0.05 95 0.00 96.15 c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 3
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 23
30.6
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.7
Refrigerant 23 (Trifluoromethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –155.13a –150 –140 –130 –120 –110 –100 –90 –82.02b –80 –78 –76 –74 –72 –70 –68 –66 –64 –62 –60 –58 –56 –54 –52 –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 25 26.14c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00006 0.00014 0.00060 0.00206 0.00588 0.01447 0.03157 0.06239 0.10132 0.11370 0.12712 0.14175 0.15768 0.17496 0.19370 0.21395 0.23582 0.25937 0.28469 0.31188 0.34102 0.37220 0.40552 0.44106 0.47893 0.51921 0.56201 0.60743 0.65557 0.70653 0.76042 0.81734 0.87740 0.94071 1.00740 1.07750 1.15130 1.22880 1.31000 1.39530 1.48460 1.57810 1.67600 1.77840 1.88530 1.99710 2.11370 2.23540 2.36240 2.49470 2.63260 2.77620 2.92590 3.08170 3.24380 3.41270 3.58850 3.77150 3.96220 4.16100 4.69860 4.83200
1701.9 1685.3 1652.0 1617.9 1583.3 1548.2 1512.3 1475.7 1445.6 1437.9 1430.2 1422.5 1414.7 1406.8 1398.9 1390.9 1382.8 1374.7 1366.4 1358.1 1349.7 1341.3 1332.7 1324.0 1315.3 1306.4 1297.4 1288.3 1279.0 1269.7 1260.2 1250.5 1240.7 1230.7 1220.5 1210.1 1199.6 1188.8 1177.8 1166.6 1155.0 1143.2 1131.1 1118.7 1105.9 1092.7 1079.1 1064.9 1050.3 1035.1 1019.2 1002.5 985.0 966.5 946.7 925.6 902.6 877.4 849.1 816.4 680.1 526.5
241.360 105.950 26.330 8.21940 3.06950 1.32190 0.63807 0.33766 0.21446 0.19248 0.17333 0.15644 0.14151 0.12828 0.11652 0.10605 0.09669 0.08832 0.08081 0.07406 0.06798 0.06249 0.05753 0.05303 0.04895 0.04523 0.04185 0.03877 0.03595 0.03336 0.03100 0.02882 0.02682 0.02498 0.02328 0.02171 0.02026 0.01892 0.01767 0.01651 0.01544 0.01443 0.01350 0.01263 0.01181 0.01105 0.01033 0.00966 0.00902 0.00843 0.00786 0.00733 0.00683 0.00635 0.00589 0.00545 0.00502 0.00461 0.00421 0.00381 0.00263 0.00190
*Temperatures on ITS-90 scale
–3.67 2.54 14.53 26.48 38.45 50.47 62.56 74.76 84.60 87.10 89.59 92.09 94.60 97.12 99.64 102.18 104.72 107.28 109.85 112.43 115.02 117.62 120.24 122.88 125.53 128.19 130.87 133.57 136.29 139.02 141.78 144.56 147.36 150.18 153.03 155.90 158.80 161.73 164.69 167.68 170.71 173.77 176.88 180.02 183.21 186.45 189.75 193.10 196.51 200.00 203.56 207.22 210.97 214.84 218.84 223.00 227.37 231.98 236.93 242.36 261.94 280.97
289.21 291.73 296.65 301.56 306.45 311.27 315.98 320.51 323.96 324.81 325.63 326.44 327.24 328.03 328.79 329.55 330.29 331.01 331.71 332.40 333.06 333.71 334.33 334.94 335.52 336.07 336.61 337.11 337.59 338.04 338.46 338.85 339.21 339.53 339.82 340.06 340.27 340.43 340.55 340.62 340.63 340.59 340.49 340.33 340.09 339.78 339.39 338.91 338.33 337.64 336.83 335.89 334.78 333.50 332.01 330.27 328.22 325.79 322.84 319.17 301.55 280.97
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid
Vapor
Liquid Vapor
cp /cv Vapor
–0.0705 –0.0190 0.0746 0.1612 0.2420 0.3179 0.3898 0.4582 0.5106 0.5236 0.5364 0.5491 0.5617 0.5742 0.5866 0.5990 0.6112 0.6234 0.6356 0.6476 0.6596 0.6716 0.6835 0.6953 0.7071 0.7189 0.7306 0.7423 0.7539 0.7655 0.7771 0.7887 0.8002 0.8118 0.8233 0.8348 0.8464 0.8579 0.8695 0.8810 0.8927 0.9043 0.9160 0.9277 0.9395 0.9514 0.9634 0.9755 0.9877 1.0000 1.0125 1.0252 1.0382 1.0514 1.0650 1.0790 1.0936 1.1089 1.1252 1.1430 1.2067 1.2697
2.4111 2.3293 2.1934 2.0828 1.9919 1.9165 1.8534 1.8000 1.7630 1.7543 1.7459 1.7378 1.7299 1.7221 1.7146 1.7073 1.7001 1.6931 1.6863 1.6796 1.6731 1.6667 1.6604 1.6542 1.6482 1.6422 1.6363 1.6305 1.6248 1.6191 1.6135 1.6080 1.6025 1.5970 1.5915 1.5861 1.5806 1.5752 1.5697 1.5642 1.5586 1.5530 1.5473 1.5416 1.5357 1.5297 1.5235 1.5172 1.5107 1.5039 1.4969 1.4895 1.4817 1.4735 1.4647 1.4552 1.4448 1.4333 1.4203 1.4050 1.3396 1.2697
1.221 1.205 1.195 1.195 1.199 1.205 1.213 1.225 1.237 1.241 1.245 1.249 1.253 1.258 1.262 1.267 1.272 1.278 1.283 1.290 1.296 1.303 1.310 1.317 1.325 1.333 1.342 1.351 1.361 1.371 1.382 1.393 1.405 1.418 1.432 1.447 1.462 1.479 1.497 1.517 1.538 1.561 1.586 1.613 1.643 1.676 1.712 1.754 1.800 1.853 1.915 1.986 2.071 2.173 2.300 2.461 2.673 2.966 3.404 4.130 18.870
1.315 1.311 1.304 1.297 1.292 1.289 1.290 1.294 1.301 1.304 1.306 1.309 1.312 1.315 1.319 1.323 1.327 1.332 1.337 1.342 1.347 1.354 1.360 1.367 1.375 1.383 1.392 1.401 1.411 1.422 1.433 1.446 1.459 1.474 1.490 1.507 1.526 1.546 1.569 1.593 1.620 1.650 1.682 1.719 1.759 1.805 1.857 1.916 1.983 2.061 2.152 2.259 2.388 2.545 2.742 2.993 3.325 3.787 4.474 5.602 26.080
a Triple
point
0.500 0.506 0.522 0.541 0.565 0.593 0.625 0.661 0.694 0.702 0.711 0.721 0.730 0.740 0.750 0.760 0.771 0.782 0.793 0.805 0.817 0.830 0.843 0.856 0.870 0.885 0.900 0.915 0.932 0.949 0.967 0.985 1.005 1.026 1.048 1.072 1.097 1.123 1.152 1.182 1.215 1.251 1.290 1.333 1.38 1.432 1.491 1.556 1.630 1.715 1.814 1.929 2.067 2.234 2.441 2.704 3.053 3.536 4.252 5.429 26.84
Velocity of Sound, m/s Liquid Vapor 1211 1200 1151 1092 1033 976 921 866 823 812 801 790 779 769 758 747 736 725 714 703 692 681 670 659 648 637 625 614 603 591 580 569 557 545 534 522 510 498 486 474 462 449 437 424 411 399 386 372 359 345 332 318 303 289 274 259 243 227 210 193 141 0
135.7 138.4 143.3 147.9 152.2 156.0 159.5 162.4 164.3 164.7 165.1 165.4 165.8 166.1 166.3 166.6 166.8 167.0 167.1 167.2 167.3 167.3 167.4 167.3 167.3 167.2 167.0 166.8 166.6 166.4 166.1 165.7 165.3 164.9 164.4 163.9 163.3 162.7 162 161.3 160.5 159.7 158.8 157.9 156.9 155.8 154.7 153.5 152.3 151.0 149.6 148.1 146.5 144.9 143.2 141.3 139.4 137.3 135.0 132.5 124.3 0.0 b Normal
Viscosity, Pa·s
Thermal Cond., mW/(m·K)
Surface Tension, Temp.,* Liquid Vapor Liquid Vapor mN/m °C 2055.0 1662.0 1153.0 845.9 648.3 514.4 419.2 349.0 305.1 295.3 286.1 277.2 268.8 260.7 253.0 245.6 238.5 231.6 225.1 218.7 212.6 206.7 201.1 195.6 190.3 185.1 180.1 175.3 170.6 166.0 161.6 157.3 153.1 149.0 144.9 141.0 137.2 133.4 129.7 126.1 122.6 119.1 115.6 112.2 108.9 105.6 102.3 99.0 95.8 92.5 89.3 86.1 82.8 79.6 76.2 72.9 69.4 65.8 62.0 57.9 44.3 —
5.35 5.64 6.22 6.80 7.37 7.93 8.48 9.03 9.47 9.58 9.69 9.80 9.91 10.01 10.12 10.23 10.34 10.45 10.56 10.67 10.78 10.89 11.00 11.11 11.22 11.34 11.45 11.57 11.68 11.80 11.92 12.04 12.16 12.29 12.42 12.55 12.68 12.81 12.95 13.10 13.25 13.40 13.56 13.73 13.90 14.09 14.28 14.49 14.71 14.94 15.19 15.47 15.77 16.10 16.47 16.89 17.38 17.95 18.64 19.51 24.79 —
boiling point
268.6 250.2 220.9 198.2 180.1 165.3 152.9 142.4 135.0 133.2 131.5 129.8 128.2 126.6 125.1 123.5 122.0 120.5 119.1 117.7 116.2 114.9 113.5 112.1 110.8 109.5 108.2 106.9 105.6 104.3 103.1 101.8 100.5 99.3 98.0 96.8 95.5 94.3 93.1 91.8 90.5 89.3 88.0 86.7 85.4 84.1 82.7 81.4 80.0 78.5 77.1 75.6 74.0 72.4 70.7 68.9 67.0 65.0 62.7 60.2 53.4
3.80 4.07 4.61 5.15 5.70 6.25 6.82 7.42 7.92 8.05 8.18 8.31 8.45 8.59 8.73 8.87 9.02 9.16 9.32 9.47 9.63 9.79 9.96 10.13 10.30 10.48 10.66 10.85 11.04 11.24 11.45 11.66 11.88 12.10 12.33 12.57 12.82 13.08 13.35 13.63 13.92 14.23 14.55 14.88 15.23 15.60 15.99 16.40 16.83 17.29 17.79 18.32 18.89 19.51 20.20 20.95 21.80 22.76 23.88 25.25 35.21
34.37 33.12 30.71 28.35 26.03 23.75 21.53 19.35 17.65 17.23 16.81 16.40 15.98 15.57 15.16 14.76 14.36 13.96 13.56 13.16 12.77 12.38 11.99 11.61 11.22 10.85 10.47 10.10 9.73 9.36 9.00 8.64 8.28 7.93 7.58 7.23 6.89 6.55 6.21 5.88 5.56 5.23 4.92 4.60 4.29 3.99 3.69 3.40 3.11 2.83 2.55 2.28 2.02 1.76 1.52 1.28 1.05 0.83 0.63 0.44 0.05 0.00
–155.13 –150 –140 –130 –120 –110 –100 –90 –82.02 –80 –78 –76 –74 –72 –70 –68 –66 –64 –62 –60 –58 –56 –54 –52 –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 25 26.14
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 4
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 32
30.8
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.9
Refrigerant 32 (Difluoromethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Enthalpy, Pres- Density, Volume, kJ/kg Temp.,* sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor °C –136.81a –130 –120 –110 –100 –90 –80 –70 –60 –51.65b –50 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 75 78.11c
0.00005 0.00013 0.00048 0.00145 0.00381 0.00887 0.01865 0.03607 0.06496 0.10133 0.11014 0.17741 0.19409 0.21197 0.23111 0.25159 0.27344 0.29675 0.32157 0.34796 0.37600 0.40575 0.43728 0.47067 0.50597 0.54327 0.58263 0.62414 0.66786 0.71388 0.76226 0.81310 0.86647 0.92245 0.98113 1.04260 1.10690 1.17420 1.24450 1.31790 1.39460 1.47460 1.55790 1.64480 1.73530 1.82950 1.92750 2.02940 2.13530 2.24540 2.35970 2.47830 2.60140 2.72920 2.86160 2.99890 3.14120 3.28870 3.44150 3.59970 3.76350 3.93320 4.10890 4.29090 4.47930 4.67450 4.87680 5.41680 5.78200
1429.3 1412.7 1388.4 1363.8 1339.0 1313.9 1288.4 1262.4 1235.7 1212.9 1208.4 1180.2 1174.4 1168.6 1162.8 1156.9 1151.0 1145.0 1138.9 1132.9 1126.7 1120.6 1114.3 1108.0 1101.7 1095.2 1088.8 1082.2 1075.6 1068.9 1062.1 1055.3 1048.3 1041.3 1034.2 1027.0 1019.7 1012.2 1004.7 997.1 989.3 981.4 973.3 965.2 956.8 948.3 939.6 930.7 921.7 912.4 902.8 893.0 883.0 872.6 861.9 850.8 839.3 827.3 814.8 801.7 787.9 773.3 757.8 741.1 723.0 703.2 680.9 605.9 424.0
453.850 174.360 51.1840 17.9070 7.22200 3.27210 1.63160 0.88072 0.50786 0.33468 0.30944 0.19743 0.18134 0.16680 0.15365 0.14173 0.13091 0.12107 0.11211 0.10393 0.09646 0.08963 0.08337 0.07762 0.07234 0.06749 0.06301 0.05889 0.05508 0.05155 0.04829 0.04527 0.04246 0.03986 0.03743 0.03518 0.03308 0.03112 0.02929 0.02758 0.02598 0.02448 0.02307 0.02175 0.02051 0.01935 0.01826 0.01722 0.01625 0.01533 0.01447 0.01365 0.01287 0.01214 0.01144 0.01078 0.01015 0.00955 0.00897 0.00843 0.00790 0.00740 0.00691 0.00644 0.00598 0.00553 0.00508 0.00391 0.00236
*Temperatures on ITS-90 scale
–19.07 –8.26 7.52 23.20 38.83 54.42 70.02 85.66 101.38 114.59 117.22 133.23 136.45 139.69 142.93 146.18 149.45 152.72 156.01 159.31 162.62 165.94 169.28 172.63 175.99 179.37 182.76 186.18 189.60 193.05 196.52 200.00 203.50 207.03 210.58 214.15 217.74 221.36 225.01 228.68 232.39 236.12 239.89 243.69 247.53 251.40 255.32 259.28 263.28 267.34 271.45 275.61 279.84 284.13 288.50 292.95 297.49 302.12 306.87 311.74 316.75 321.93 327.30 332.90 338.78 345.02 351.73 372.39 414.15
Specific Heat cp , kJ/(kg·K)
Entropy, kJ/(kg·K)
Velocity of Viscosity, Pa·s Sound, m/s cp /cv Vapor Liquid Vapor Liquid Vapor
Liquid
Vapor
Liquid Vapor
444.31 –0.1050 448.77 –0.0276 455.33 0.0790 461.86 0.1782 468.31 0.2711 474.61 0.3586 480.72 0.4415 486.57 0.5204 492.11 0.5958 496.45 0.6565 497.27 0.6683 502.02 0.7382 502.91 0.7519 503.78 0.7655 504.63 0.7791 505.47 0.7926 506.27 0.8060 507.06 0.8193 507.83 0.8326 508.57 0.8458 509.28 0.8589 509.97 0.8720 510.64 0.8850 511.28 0.8979 511.89 0.9109 512.47 0.9237 513.02 0.9365 513.54 0.9493 514.03 0.9620 514.49 0.9747 514.91 0.9874 515.30 1.0000 515.65 1.0126 515.96 1.0252 516.24 1.0377 516.47 1.0503 516.66 1.0628 516.80 1.0753 516.90 1.0878 516.95 1.1003 516.95 1.1128 516.90 1.1253 516.79 1.1378 516.62 1.1503 516.39 1.1629 516.09 1.1755 515.72 1.1881 515.29 1.2007 514.77 1.2134 514.17 1.2262 513.49 1.2391 512.71 1.2520 511.82 1.2650 510.83 1.2781 509.72 1.2914 508.48 1.3048 507.10 1.3183 505.57 1.3321 503.86 1.3461 501.95 1.3603 499.82 1.3749 497.44 1.3898 494.76 1.4052 491.73 1.4211 488.26 1.4377 484.25 1.4553 479.52 1.4740 461.72 1.5314 414.15 1.6486
3.2937 3.1651 3.0030 2.8668 2.7515 2.6529 2.5679 2.4939 2.4289 2.3805 2.3714 2.3200 2.3103 2.3008 2.2916 2.2824 2.2735 2.2647 2.2561 2.2476 2.2392 2.2310 2.2229 2.2149 2.2070 2.1992 2.1915 2.1839 2.1764 2.1690 2.1616 2.1543 2.1471 2.1399 2.1327 2.1256 2.1185 2.1114 2.1043 2.0972 2.0902 2.0831 2.0760 2.0688 2.0616 2.0544 2.0471 2.0397 2.0322 2.0246 2.0169 2.0091 2.0011 1.9929 1.9845 1.9759 1.9670 1.9578 1.9482 1.9382 1.9277 1.9166 1.9048 1.8922 1.8785 1.8634 1.8464 1.7880 1.6486
1.592 0.660 1.321 1.583 0.665 1.318 1.573 0.674 1.315 1.565 0.686 1.312 1.560 0.703 1.310 1.559 0.725 1.310 1.561 0.754 1.311 1.566 0.790 1.314 1.576 0.833 1.320 1.587 0.875 1.328 1.589 0.883 1.329 1.608 0.940 1.343 1.612 0.952 1.347 1.616 0.965 1.350 1.621 0.977 1.354 1.626 0.990 1.358 1.631 1.004 1.363 1.637 1.017 1.367 1.642 1.031 1.372 1.648 1.045 1.377 1.654 1.060 1.383 1.661 1.075 1.389 1.668 1.090 1.395 1.675 1.106 1.401 1.682 1.122 1.408 1.690 1.139 1.416 1.698 1.156 1.423 1.706 1.174 1.432 1.715 1.192 1.440 1.725 1.211 1.450 1.735 1.231 1.460 1.745 1.251 1.470 1.756 1.272 1.481 1.767 1.294 1.493 1.779 1.317 1.506 1.792 1.341 1.519 1.806 1.367 1.534 1.820 1.393 1.549 1.835 1.421 1.565 1.851 1.450 1.583 1.868 1.481 1.602 1.886 1.514 1.622 1.905 1.548 1.644 1.926 1.585 1.668 1.948 1.624 1.693 1.972 1.667 1.721 1.997 1.712 1.750 2.025 1.760 1.783 2.055 1.813 1.819 2.088 1.870 1.858 2.124 1.933 1.901 2.163 2.001 1.948 2.206 2.077 2.001 2.255 2.160 2.060 2.309 2.254 2.126 2.369 2.358 2.201 2.439 2.477 2.287 2.518 2.613 2.385 2.609 2.771 2.499 2.717 2.956 2.633 2.845 3.175 2.793 3.001 3.441 2.987 3.193 3.771 3.228 3.438 4.190 3.535 3.761 4.743 3.938 4.207 5.508 4.495 4.865 6.639 5.316 10.130 15.600 11.720
a Triple
point
1414 1378 1326 1273 1221 1169 1118 1066 1014 971 962 910 900 889 879 868 858 847 837 826 816 805 794 784 773 762 751 741 730 719 708 697 686 675 664 652 641 630 618 607 595 584 572 560 548 536 524 512 499 487 474 461 448 435 421 408 394 379 365 350 335 320 304 288 271 254 236 186 0
169.6 1226.0 173.6 1023.0 179.2 811.8 184.5 664.6 189.5 556.1 194.1 472.6 198.3 406.4 202.0 352.7 205.1 308.2 207.4 276.7 207.7 271.0 209.7 239.4 210.1 233.6 210.4 228.1 210.6 222.6 210.9 217.4 211.1 212.3 211.3 207.4 211.4 202.6 211.5 197.9 211.6 193.3 211.7 188.9 211.7 184.6 211.7 180.5 211.7 176.4 211.6 172.4 211.5 168.5 211.4 164.8 211.2 161.1 211.0 157.5 210.8 154.0 210.5 150.6 210.2 147.3 209.8 144.0 209.4 140.8 209.0 137.7 208.5 134.6 208.0 131.6 207.5 128.7 206.9 125.8 206.3 123.0 205.6 120.3 204.9 117.5 204.1 114.9 203.3 112.2 202.4 109.7 201.5 107.1 200.6 104.6 199.6 102.1 198.5 99.7 197.4 97.3 196.2 94.9 194.9 92.5 193.6 90.2 192.3 87.8 190.8 85.5 189.3 83.2 187.7 80.8 186.0 78.5 184.3 76.1 182.4 73.8 180.4 71.4 178.3 68.9 176.1 66.4 173.7 63.8 171.2 61.1 168.4 58.2 159.6 49.5 0.0 — b Normal
5.70 5.97 6.39 6.80 7.22 7.64 8.06 8.48 8.91 9.26 9.33 9.75 9.84 9.92 10.01 10.09 10.18 10.26 10.35 10.43 10.52 10.61 10.70 10.78 10.87 10.96 11.05 11.14 11.23 11.32 11.42 11.51 11.61 11.70 11.80 11.90 12.00 12.10 12.48 12.60 12.73 12.86 13.00 13.14 13.28 13.43 13.58 13.74 13.90 14.07 14.25 14.44 14.64 14.84 15.06 15.29 15.54 15.80 16.09 16.39 16.73 17.09 17.49 17.95 18.46 19.06 19.76 22.56 —
boiling point
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 242.9 240.6 236.1 230.6 224.3 217.4 210.0 202.2 194.2 187.4 186.0 177.8 176.1 174.5 172.8 171.2 169.5 167.9 166.3 164.6 163.0 161.3 159.7 158.1 156.5 154.9 153.2 151.6 150.0 148.4 146.8 145.3 143.7 142.1 140.5 139.0 137.4 135.9 134.3 132.8 131.2 129.7 128.2 126.6 125.1 123.6 122.1 120.6 119.1 117.6 116.1 114.6 113.1 111.6 110.1 108.6 107.0 105.5 104.0 102.5 100.9 99.4 97.8 96.3 94.8 93.3 92.0 91.5
6.95 6.97 7.02 7.12 7.27 7.45 7.68 7.96 8.28 8.60 8.66 9.10 9.19 9.29 9.39 9.50 9.60 9.71 9.83 9.95 10.07 10.19 10.32 10.46 10.60 10.74 10.89 11.04 11.21 11.38 11.55 11.73 11.93 12.13 12.34 12.56 12.79 13.04 13.29 13.56 13.85 14.16 14.48 14.83 15.19 15.59 16.01 16.46 16.95 17.47 18.04 18.65 19.32 20.05 20.85 21.73 22.69 23.77 24.97 26.31 27.83 29.55 31.54 33.85 36.59 39.90 44.04 62.91
39.01 37.47 35.23 33.02 30.83 28.68 26.56 24.48 22.42 20.74 20.41 18.44 18.05 17.66 17.27 16.89 16.50 16.12 15.74 15.36 14.99 14.61 14.24 13.87 13.50 13.14 12.77 12.41 12.05 11.69 11.34 10.99 10.63 10.29 9.94 9.60 9.25 8.91 8.58 8.24 7.91 7.59 7.26 6.94 6.62 6.30 5.99 5.68 5.37 5.07 4.77 4.47 4.18 3.89 3.61 3.33 3.06 2.79 2.52 2.26 2.01 1.76 1.52 1.29 1.06 0.85 0.64 0.19 0.00
–136.81 –130 –120 –110 –100 –90 –80 –70 –60 –51.65 –50 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 75 78.11
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 5 Pressure-Enthalpy Diagram for Refrigerant 123
Licensed for single user. © 2017 ASHRAE, Inc.
30.10
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.11
Refrigerant 123 (2,2-Dichloro-1,1,1-Trifluoroethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –80 –70 –60 –50 –40 –30 –20 –10 0 2 4 6 8 10 12 14 16 18 20 22 24 26 27.82b 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 82 84 86 88 90 92 94 96 98 100 110 120 130 140 150 160 170 180 183.68c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00013 0.00034 0.00081 0.00177 0.00358 0.00675 0.01200 0.02025 0.03265 0.03574 0.03907 0.04264 0.04647 0.05057 0.05495 0.05963 0.06463 0.06995 0.07561 0.08163 0.08802 0.09480 0.10133 0.10198 0.10958 0.11762 0.12611 0.13507 0.14452 0.15447 0.16495 0.17597 0.18755 0.19971 0.21246 0.22584 0.23985 0.25451 0.26985 0.28589 0.30264 0.32013 0.33838 0.35740 0.37722 0.39787 0.41936 0.44171 0.46494 0.48909 0.51416 0.54019 0.56720 0.59520 0.62423 0.65430 0.68544 0.71768 0.75103 0.78553 0.97603 1.19900 1.45780 1.75630 2.09870 2.49010 2.93720 3.45060 3.66180
1709.6 1687.4 1665.1 1642.6 1620.0 1597.0 1573.8 1550.1 1526.1 1521.3 1516.4 1511.5 1506.6 1501.6 1496.7 1491.7 1486.7 1481.7 1476.6 1471.5 1466.4 1461.3 1456.6 1456.2 1451.0 1445.8 1440.6 1435.4 1430.1 1424.8 1419.4 1414.1 1408.7 1403.3 1397.8 1392.3 1386.8 1381.2 1375.6 1370.0 1364.3 1358.6 1352.8 1347.0 1341.2 1335.3 1329.3 1323.4 1317.3 1311.2 1305.1 1298.9 1292.6 1286.3 1279.9 1273.5 1266.9 1260.3 1253.7 1246.9 1211.9 1174.4 1133.6 1088.3 1036.8 975.7 896.9 765.9 550.0
83.6670 32.8420 14.3330 6.84600 3.53190 1.94700 1.13640 0.69690 0.44609 0.40991 0.37720 0.34759 0.32075 0.29637 0.27420 0.25401 0.23559 0.21877 0.20338 0.18929 0.17637 0.16451 0.15453 0.15360 0.14356 0.13431 0.12577 0.11789 0.11060 0.10385 0.09759 0.09179 0.08641 0.08140 0.07674 0.07240 0.06836 0.06458 0.06106 0.05777 0.05469 0.05180 0.04910 0.04656 0.04418 0.04195 0.03985 0.03787 0.03601 0.03426 0.03261 0.03105 0.02958 0.02819 0.02687 0.02563 0.02445 0.02334 0.02228 0.02128 0.01697 0.01361 0.01094 0.00879 0.00703 0.00555 0.00425 0.00292 0.00182
*Temperatures on ITS-90 scale
123.92 133.17 142.46 151.81 161.25 170.78 180.41 190.15 200.00 201.98 203.97 205.97 207.96 209.97 211.97 213.99 216.00 218.02 220.05 222.08 224.12 226.16 228.03 228.21 230.26 232.31 234.38 236.44 238.51 240.59 242.67 244.76 246.86 248.95 251.06 253.17 255.28 257.41 259.53 261.67 263.81 265.95 268.10 270.26 272.42 274.60 276.77 278.96 281.15 283.35 285.55 287.77 289.99 292.22 294.45 296.70 298.95 301.21 303.49 305.77 317.32 329.15 341.32 353.92 367.10 381.13 396.61 416.22 437.39
335.98 341.25 346.66 352.21 357.88 363.65 369.52 375.45 381.44 382.64 383.84 385.05 386.25 387.46 388.66 389.87 391.08 392.29 393.49 394.70 395.91 397.12 398.22 398.32 399.53 400.73 401.93 403.14 404.34 405.54 406.73 407.93 409.12 410.31 411.50 412.69 413.87 415.05 416.23 417.40 418.57 419.73 420.89 422.05 423.20 424.35 425.50 426.63 427.77 428.89 430.01 431.13 432.23 433.33 434.43 435.51 436.59 437.66 438.72 439.77 444.88 449.67 454.07 457.94 461.05 463.01 462.89 456.82 437.39
Entropy, kJ/(kg·K) Liquid
Vapor
0.6712 0.7179 0.7625 0.8054 0.8468 0.8868 0.9256 0.9633 1.0000 1.0072 1.0144 1.0216 1.0287 1.0358 1.0428 1.0499 1.0569 1.0638 1.0707 1.0776 1.0845 1.0913 1.0975 1.0981 1.1049 1.1116 1.1183 1.1250 1.1317 1.1383 1.1449 1.1515 1.1581 1.1646 1.1711 1.1776 1.1840 1.1905 1.1969 1.2033 1.2096 1.2160 1.2223 1.2286 1.2349 1.2411 1.2474 1.2536 1.2598 1.2660 1.2722 1.2783 1.2845 1.2906 1.2967 1.3028 1.3089 1.3150 1.3211 1.3271 1.3572 1.3872 1.4173 1.4475 1.4782 1.5101 1.5443 1.5867 1.6325
1.7691 1.7422 1.7206 1.7034 1.6901 1.6800 1.6726 1.6675 1.6642 1.6638 1.6634 1.6631 1.6628 1.6626 1.6625 1.6624 1.6623 1.6623 1.6624 1.6625 1.6626 1.6628 1.6630 1.6630 1.6633 1.6635 1.6639 1.6642 1.6646 1.6651 1.6655 1.6660 1.6665 1.6670 1.6676 1.6682 1.6688 1.6694 1.6701 1.6707 1.6714 1.6721 1.6728 1.6735 1.6743 1.6750 1.6758 1.6766 1.6774 1.6781 1.6789 1.6797 1.6806 1.6814 1.6822 1.6830 1.6838 1.6846 1.6854 1.6862 1.6902 1.6938 1.6969 1.6992 1.7003 1.6991 1.6939 1.6763 1.6325
Velocity of Sound, m/s cp /cv Liquid Vapor Vapor Liquid Vapor Specific Heat cp , kJ/(kg·K) 0.924 0.927 0.932 0.939 0.948 0.958 0.968 0.979 0.990 0.993 0.995 0.997 0.999 1.002 1.004 1.006 1.009 1.011 1.014 1.016 1.018 1.021 1.023 1.023 1.026 1.028 1.031 1.033 1.036 1.038 1.041 1.044 1.046 1.049 1.052 1.055 1.058 1.060 1.063 1.066 1.069 1.072 1.076 1.079 1.082 1.085 1.089 1.092 1.096 1.100 1.103 1.107 1.111 1.115 1.120 1.124 1.129 1.133 1.138 1.143 1.172 1.207 1.254 1.318 1.415 1.584 1.979 4.549
0.520 0.537 0.553 0.569 0.585 0.601 0.617 0.634 0.651 0.654 0.658 0.661 0.665 0.668 0.672 0.675 0.679 0.682 0.686 0.690 0.693 0.697 0.701 0.701 0.705 0.709 0.712 0.716 0.720 0.724 0.728 0.732 0.736 0.741 0.745 0.749 0.753 0.758 0.762 0.767 0.771 0.776 0.781 0.785 0.790 0.795 0.800 0.806 0.811 0.816 0.822 0.827 0.833 0.839 0.845 0.851 0.858 0.864 0.871 0.878 0.917 0.964 1.026 1.111 1.240 1.473 2.033 5.661
b Normal
1.117 1.113 1.110 1.107 1.105 1.103 1.102 1.102 1.102 1.103 1.103 1.103 1.103 1.104 1.104 1.104 1.105 1.105 1.106 1.106 1.107 1.107 1.108 1.108 1.109 1.109 1.110 1.111 1.112 1.113 1.114 1.115 1.116 1.117 1.119 1.120 1.121 1.123 1.124 1.126 1.127 1.129 1.131 1.133 1.135 1.137 1.139 1.142 1.144 1.147 1.150 1.152 1.156 1.159 1.162 1.166 1.169 1.173 1.177 1.182 1.208 1.243 1.294 1.369 1.493 1.726 2.309 6.158
boiling point
1133 1091 1049 1006 964 923 881 841 801 793 785 777 769 761 754 746 738 730 723 715 707 700 693 692 684 677 669 662 654 647 639 632 624 617 610 602 595 588 580 573 566 558 551 544 536 529 522 515 507 500 493 486 478 471 464 457 449 442 435 427 391 354 317 279 239 198 154 102 0
108.3 110.8 113.3 115.6 117.9 120.0 122.0 123.8 125.4 125.7 126.0 126.3 126.6 126.8 127.1 127.3 127.6 127.8 128.0 128.2 128.4 128.6 128.7 128.7 128.9 129.0 129.1 129.3 129.4 129.5 129.5 129.6 129.7 129.7 129.7 129.7 129.7 129.7 129.7 129.6 129.6 129.5 129.4 129.3 129.2 129.0 128.9 128.7 128.5 128.3 128.1 127.8 127.6 127.3 127.0 126.6 126.3 125.9 125.5 125.1 122.8 119.8 116.0 111.5 106.0 99.3 91.1 80.6 0.0
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor Liquid Vapor mN/m °C Viscosity, Pa·s
2093.0 1680.0 1383.0 1160.0 986.4 848.0 735.4 642.4 564.6 550.6 537.0 523.8 511.1 498.8 486.8 475.3 464.0 453.2 442.6 432.4 422.4 412.8 404.2 403.4 394.3 385.4 376.8 368.4 360.3 352.4 344.7 337.2 329.9 322.8 315.9 309.1 302.6 296.2 289.9 283.9 277.9 272.1 266.5 261.0 255.6 250.4 245.2 240.2 235.3 230.5 225.9 221.3 216.8 212.5 208.2 204.0 199.9 195.9 191.9 188.1 169.9 153.4 138.1 123.8 110.2 96.8 82.7 64.3 —
6.68 7.09 7.50 7.91 8.31 8.70 9.09 9.47 9.84 9.91 9.99 10.06 10.13 10.20 10.28 10.35 10.42 10.49 10.56 10.63 10.70 10.77 10.84 10.84 10.91 10.98 11.05 11.12 11.19 11.26 11.33 11.40 11.46 11.53 11.60 11.67 11.74 11.80 11.87 11.94 12.01 12.07 12.14 12.21 12.28 12.35 12.42 12.49 12.55 12.63 12.70 12.77 12.84 12.91 12.98 13.06 13.14 13.21 13.29 13.37 13.80 14.29 14.89 15.65 16.68 18.19 20.71 26.59 —
107.4 104.8 102.0 99.1 96.1 93.0 89.8 86.7 83.7 83.1 82.5 81.9 81.3 80.7 80.1 79.5 79.0 78.4 77.8 77.3 76.7 76.1 75.6 75.6 75.0 74.5 74.0 73.4 72.9 72.4 71.8 71.3 70.8 70.3 69.8 69.3 68.8 68.3 67.8 67.3 66.8 66.3 65.9 65.4 64.9 64.5 64.0 63.5 63.1 62.6 62.2 61.7 61.3 60.8 60.4 59.9 59.5 59.1 58.6 58.2 56.0 53.9 51.7 49.5 47.2 44.8 42.3 39.7
3.22 3.79 4.35 4.92 5.49 6.05 6.61 7.18 7.74 7.86 7.97 8.08 8.20 8.31 8.43 8.54 8.66 8.77 8.89 9.01 9.12 9.24 9.35 9.36 9.48 9.60 9.72 9.84 9.96 10.08 10.20 10.32 10.45 10.57 10.70 10.82 10.95 11.08 11.21 11.34 11.47 11.61 11.74 11.88 12.01 12.15 12.29 12.44 12.58 12.73 12.87 13.02 13.17 13.33 13.48 13.64 13.80 13.96 14.13 14.29 15.17 16.14 17.22 18.44 19.87 21.63 24.05 28.82
28.42 27.09 25.78 24.48 23.19 21.92 20.66 19.41 18.18 17.94 17.70 17.45 17.21 16.97 16.73 16.49 16.25 16.01 15.77 15.53 15.30 15.06 14.84 14.82 14.59 14.35 14.12 13.89 13.66 13.43 13.20 12.97 12.74 12.51 12.28 12.05 11.83 11.60 11.38 11.16 10.93 10.71 10.49 10.27 10.05 9.84 9.62 9.40 9.19 8.97 8.76 8.55 8.34 8.13 7.92 7.71 7.50 7.30 7.09 6.89 5.88 4.91 3.98 3.09 2.24 1.45 0.74 0.15 0.00
–80 –70 –60 –50 –40 –30 –20 –10 0 2 4 6 8 10 12 14 16 18 20 22 24 26 27.82 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 82 84 86 88 90 92 94 96 98 100 110 120 130 140 150 160 170 180 183.68
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 6 Pressure-Enthalpy Diagram for Refrigerant 124
30.12
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.13
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 124 (2-Chloro-1,1,1,2-Tetrafluoroethane) Properties of Saturated Liquid and Saturated Vapor Temp.,* °C
Pressure, MPa
Enthalpy, Density, Volume, kJ/kg kg/m3 m3/kg Liquid Vapor Liquid Vapor
–100 –90 –80 –70 –60 –50 –40 –30 –20 –18 –16 –14 –12 –11.96b –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 95 100 105 110 115 120 122.28c
0.00024 0.00067 0.00169 0.00379 0.00779 0.01482 0.02642 0.04452 0.07145 0.07813 0.08529 0.09296 0.10117 0.10133 0.10993 0.11928 0.12923 0.13983 0.15108 0.16303 0.17570 0.18911 0.20331 0.21830 0.23414 0.25084 0.26844 0.28696 0.30644 0.32692 0.34842 0.37097 0.39462 0.41938 0.44530 0.47241 0.50075 0.53034 0.56123 0.59345 0.62704 0.66202 0.69845 0.73635 0.77577 0.81675 0.85931 0.90350 0.94937 0.99695 1.04630 1.09740 1.15040 1.20520 1.26200 1.32070 1.38150 1.44430 1.50930 1.57640 1.75400 1.94620 2.15420 2.37870 2.62120 2.88310 3.16620 3.47390 3.62430
1714.6 44.37500 98.87 302.29 1688.6 16.53100 108.45 307.68 1662.3 6.96450 118.13 313.21 1635.9 3.25230 127.91 318.86 1609.1 1.65600 137.80 324.62 1582.0 0.90713 147.81 330.46 1554.3 0.52863 157.95 336.36 1526.1 0.32470 168.23 342.29 1497.3 0.20856 178.66 348.23 1491.4 0.19180 180.76 349.42 1485.5 0.17665 182.87 350.61 1479.6 0.16293 184.99 351.79 1473.6 0.15048 187.11 352.97 1473.5 0.15026 187.15 352.99 1467.6 0.13917 189.24 354.15 1461.6 0.12888 191.38 355.33 1455.5 0.11950 193.52 356.51 1449.4 0.11093 195.68 357.68 1443.2 0.10310 197.83 358.86 1437.0 0.09593 200.00 360.02 1430.8 0.08936 202.17 361.19 1424.5 0.08333 204.35 362.35 1418.1 0.07779 206.54 363.51 1411.8 0.07268 208.74 364.67 1405.3 0.06798 210.94 365.82 1398.9 0.06364 213.15 366.97 1392.3 0.05964 215.37 368.11 1385.7 0.05593 217.60 369.25 1379.1 0.05250 219.84 370.38 1372.4 0.04932 222.09 371.51 1365.6 0.04636 224.34 372.63 1358.8 0.04362 226.60 373.75 1351.9 0.04107 228.88 374.85 1345.0 0.03870 231.16 375.96 1337.9 0.03648 233.45 377.05 1330.8 0.03442 235.75 378.14 1323.7 0.03249 238.07 379.22 1316.4 0.03069 240.39 380.29 1309.1 0.02901 242.72 381.36 1301.6 0.02743 245.07 382.41 1294.1 0.02596 247.43 383.45 1286.5 0.02457 249.79 384.49 1278.8 0.02327 252.17 385.51 1271.0 0.02205 254.56 386.52 1263.1 0.02090 256.97 387.53 1255.1 0.01982 259.39 388.51 1247.0 0.01880 261.82 389.49 1238.7 0.01784 264.26 390.45 1230.3 0.01693 266.73 391.39 1221.8 0.01607 269.20 392.33 1213.2 0.01527 271.69 393.24 1204.3 0.01450 274.20 394.14 1195.4 0.01378 276.72 395.01 1186.2 0.01309 279.27 395.87 1176.9 0.01244 281.83 396.71 1167.4 0.01183 284.41 397.52 1157.6 0.01124 287.01 398.31 1147.7 0.01069 289.63 399.08 1137.5 0.01016 292.28 399.81 1127.0 0.00965 294.95 400.52 1099.6 0.00849 301.75 402.14 1070.1 0.00746 308.74 403.51 1037.8 0.00653 315.97 404.56 1001.8 0.00569 323.50 405.20 960.8 0.00492 331.44 405.26 912.1 0.00419 339.99 404.46 849.5 0.00347 349.58 402.13 749.1 0.00267 361.94 395.71 560.0 0.00179 378.79 378.79
*Temperatures on ITS-90 scale
Liquid Vapor
Velocity of Sound, m/s cp /cv Liquid Vapor Vapor Liquid Vapor
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor Liquid Vapor mN/m °C
0.5417 0.5954 0.6469 0.6962 0.7437 0.7896 0.8340 0.8772 0.9191 0.9274 0.9356 0.9438 0.9519 0.9521 0.9600 0.9681 0.9761 0.9841 0.9921 1.0000 1.0079 1.0158 1.0236 1.0314 1.0392 1.0469 1.0546 1.0623 1.0700 1.0776 1.0852 1.0928 1.1004 1.1079 1.1154 1.1229 1.1304 1.1379 1.1453 1.1528 1.1602 1.1676 1.1750 1.1824 1.1897 1.1971 1.2044 1.2118 1.2191 1.2265 1.2338 1.2411 1.2485 1.2558 1.2631 1.2705 1.2779 1.2852 1.2926 1.3001 1.3187 1.3376 1.3568 1.3766 1.3971 1.4188 1.4428 1.4734 1.5156
0.953 0.962 0.972 0.983 0.995 1.007 1.020 1.034 1.049 1.052 1.056 1.059 1.062 1.062 1.065 1.069 1.072 1.076 1.079 1.083 1.087 1.090 1.094 1.098 1.102 1.106 1.110 1.114 1.119 1.123 1.128 1.132 1.137 1.142 1.147 1.152 1.157 1.163 1.168 1.174 1.180 1.187 1.193 1.200 1.207 1.214 1.221 1.229 1.238 1.246 1.255 1.265 1.275 1.286 1.297 1.310 1.323 1.337 1.352 1.368 1.415 1.475 1.554 1.665 1.835 2.135 2.837 6.828
1964.0 1521.0 1215.0 994.1 829.0 701.6 600.9 519.4 452.4 440.4 428.8 417.6 406.8 406.6 396.3 386.2 376.4 366.9 357.7 348.8 340.2 331.8 323.7 315.8 308.2 300.8 293.5 286.5 279.7 273.1 266.6 260.3 254.2 248.2 242.4 236.7 231.2 225.8 220.5 215.4 210.4 205.5 200.7 196.0 191.4 187.0 182.6 178.3 174.1 169.9 165.9 161.9 158.0 154.2 150.4 146.7 143.0 139.4 135.8 132.3 123.7 115.2 106.9 98.4 89.8 80.8 70.7 57.3 —
Entropy, kJ/(kg·K) 1.7165 1.6832 1.6569 1.6362 1.6202 1.6081 1.5993 1.5930 1.5890 1.5884 1.5879 1.5874 1.5870 1.5870 1.5867 1.5864 1.5862 1.5860 1.5859 1.5858 1.5858 1.5858 1.5859 1.5860 1.5861 1.5863 1.5865 1.5868 1.5870 1.5873 1.5876 1.5880 1.5883 1.5887 1.5891 1.5895 1.5900 1.5904 1.5909 1.5913 1.5918 1.5923 1.5928 1.5933 1.5937 1.5942 1.5947 1.5951 1.5956 1.5960 1.5965 1.5969 1.5972 1.5976 1.5979 1.5982 1.5985 1.5987 1.5989 1.5990 1.5990 1.5986 1.5975 1.5955 1.5923 1.5870 1.5782 1.5593 1.5156
Specific Heat cp , kJ/(kg·K) 0.534 0.551 0.569 0.586 0.605 0.624 0.644 0.665 0.688 0.693 0.698 0.703 0.708 0.708 0.713 0.718 0.723 0.729 0.734 0.740 0.746 0.751 0.757 0.763 0.769 0.776 0.782 0.788 0.795 0.802 0.809 0.816 0.823 0.830 0.838 0.845 0.853 0.861 0.870 0.878 0.887 0.897 0.906 0.916 0.926 0.937 0.948 0.959 0.971 0.984 0.997 1.011 1.026 1.042 1.058 1.076 1.095 1.115 1.137 1.160 1.229 1.318 1.437 1.607 1.873 2.357 3.513 9.976
b Normal
1.129 1.125 1.122 1.119 1.117 1.116 1.116 1.118 1.120 1.121 1.122 1.123 1.124 1.124 1.125 1.126 1.127 1.128 1.129 1.131 1.132 1.134 1.135 1.137 1.139 1.141 1.143 1.145 1.148 1.150 1.153 1.156 1.159 1.162 1.165 1.169 1.172 1.176 1.180 1.185 1.189 1.194 1.199 1.205 1.211 1.217 1.224 1.231 1.239 1.247 1.256 1.266 1.276 1.287 1.299 1.313 1.327 1.343 1.360 1.379 1.436 1.513 1.618 1.773 2.022 2.478 3.576 9.676
boiling point
1052 1007 963 919 875 832 790 748 706 698 690 682 673 673 665 657 649 641 633 625 617 609 601 593 585 577 569 561 553 545 537 529 521 513 505 497 489 481 473 465 457 449 441 433 425 417 409 401 393 385 376 368 360 351 343 334 326 317 309 300 278 255 231 207 182 156 127 93 0
109.1 111.9 114.6 117.2 119.6 121.8 123.8 125.5 126.9 127.1 127.3 127.5 127.7 127.7 127.9 128.1 128.2 128.3 128.5 128.6 128.6 128.7 128.7 128.8 128.8 128.7 128.7 128.7 128.6 128.5 128.4 128.2 128.1 127.9 127.7 127.5 127.2 126.9 126.6 126.3 125.9 125.6 125.2 124.7 124.3 123.8 123.3 122.7 122.1 121.5 120.9 120.2 119.5 118.7 117.9 117.1 116.2 115.3 114.4 113.4 110.7 107.8 104.4 100.7 96.6 91.9 86.6 80.2 0.0
Viscosity, Pa·s
6.66 7.05 7.45 7.84 8.23 8.62 9.01 9.39 9.78 9.86 9.93 10.01 10.09 10.09 10.16 10.24 10.32 10.40 10.47 10.55 10.63 10.71 10.79 10.87 10.95 11.03 11.11 11.19 11.28 11.36 11.44 11.53 11.61 11.70 11.79 11.88 11.97 12.06 12.16 12.25 12.35 12.45 12.55 12.66 12.77 12.88 12.99 13.11 13.27 13.40 13.53 13.67 13.82 13.97 14.12 14.28 14.45 14.63 14.82 15.01 15.54 16.16 16.88 17.75 18.85 20.30 22.45 26.73 —
111.8 108.7 105.4 101.9 98.5 94.9 91.3 87.7 84.1 83.4 82.7 82.0 81.2 81.2 80.5 79.8 79.1 78.4 77.7 77.0 76.3 75.6 74.9 74.2 73.5 72.8 72.1 71.4 70.8 70.1 69.4 68.7 68.1 67.4 66.8 66.1 65.5 64.8 64.2 63.5 62.9 62.3 61.7 61.0 60.4 59.8 59.2 58.6 58.0 57.5 56.9 56.3 55.7 55.2 54.6 54.1 53.5 53.0 52.5 51.9 50.6 49.4 48.2 47.0 46.0 45.1 45.0 48.2
4.81 5.26 5.74 6.23 6.74 7.27 7.82 8.39 8.98 9.10 9.22 9.34 9.46 9.46 9.59 9.71 9.84 9.96 10.09 10.22 10.35 10.49 10.62 10.75 10.89 11.03 11.17 11.31 11.46 11.61 11.75 11.91 12.06 12.22 12.38 12.54 12.71 12.88 13.05 13.23 13.41 13.59 13.79 13.98 14.18 14.39 14.61 14.83 15.08 15.32 15.57 15.83 16.10 16.38 16.68 16.98 17.30 17.64 17.99 18.36 19.39 20.58 22.00 23.73 25.94 28.94 33.62 44.96
26.10 24.69 23.30 21.92 20.56 19.21 17.88 16.56 15.26 15.00 14.74 14.49 14.23 14.23 13.98 13.72 13.47 13.22 12.97 12.72 12.47 12.22 11.97 11.72 11.48 11.23 10.99 10.74 10.50 10.26 10.02 9.78 9.54 9.30 9.06 8.83 8.59 8.36 8.13 7.90 7.67 7.44 7.21 6.98 6.76 6.53 6.31 6.09 5.87 5.65 5.43 5.21 5.00 4.79 4.58 4.37 4.16 3.95 3.75 3.54 3.05 2.56 2.09 1.64 1.21 0.80 0.43 0.11 0.00
–100 –90 –80 –70 –60 –50 –40 –30 –20 –18 –16 –14 –12 –11.96 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 95 100 105 110 115 120 122.28
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
PressurePressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 7 Pressure-Enthalpy Diagram for Refrigerant 125
30.14
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.15
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 125 (Pentafluoroethane) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg Temp.,* sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor °C
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
–100.63a –100 –90 –80 –70 –60 –58 –56 –54 –52 –50 –48.09b –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66.02c
0.4902 0.4940 0.5524 0.6082 0.6620 0.7140 0.7242 0.7343 0.7444 0.7544 0.7644 0.7739 0.7743 0.7842 0.7940 0.8037 0.8134 0.8231 0.8327 0.8423 0.8519 0.8614 0.8708 0.8802 0.8896 0.8990 0.9083 0.9176 0.9268 0.9361 0.9453 0.9544 0.9636 0.9727 0.9818 0.9909 1.0000 1.0091 1.0181 1.0272 1.0362 1.0452 1.0542 1.0633 1.0723 1.0813 1.0904 1.0995 1.1085 1.1176 1.1268 1.1359 1.1452 1.1544 1.1637 1.1731 1.1826 1.1921 1.2018 1.2116 1.2216 1.2318 1.2422 1.2529 1.2641 1.2758 1.2884 1.3024 1.3195 1.3696
1.035 0.569 1.035 0.570 1.045 0.592 1.058 0.615 1.074 0.639 1.091 0.664 1.095 0.669 1.099 0.675 1.103 0.680 1.107 0.686 1.111 0.692 1.115 0.697 1.115 0.697 1.119 0.703 1.123 0.709 1.128 0.715 1.132 0.722 1.137 0.728 1.142 0.734 1.146 0.741 1.151 0.748 1.157 0.755 1.162 0.762 1.167 0.769 1.173 0.776 1.178 0.784 1.184 0.791 1.190 0.799 1.196 0.807 1.203 0.815 1.209 0.824 1.216 0.832 1.223 0.841 1.231 0.850 1.238 0.860 1.246 0.870 1.255 0.880 1.263 0.890 1.273 0.902 1.282 0.913 1.292 0.926 1.303 0.939 1.314 0.954 1.326 0.969 1.339 0.986 1.352 1.003 1.367 1.023 1.382 1.044 1.399 1.067 1.417 1.093 1.436 1.121 1.457 1.152 1.481 1.186 1.507 1.224 1.536 1.267 1.568 1.316 1.605 1.372 1.647 1.436 1.697 1.511 1.755 1.600 1.824 1.708 1.910 1.842 2.019 2.014 2.162 2.241 2.359 2.559 2.652 3.036 3.139 3.833 4.120 5.435 7.17 10.29
1.142 1.141 1.138 1.136 1.136 1.138 1.138 1.139 1.140 1.141 1.142 1.143 1.143 1.144 1.145 1.146 1.148 1.150 1.151 1.153 1.155 1.157 1.160 1.162 1.165 1.167 1.170 1.173 1.177 1.181 1.184 1.189 1.193 1.198 1.203 1.209 1.215 1.222 1.229 1.237 1.246 1.255 1.265 1.277 1.289 1.303 1.318 1.334 1.352 1.372 1.395 1.420 1.448 1.480 1.516 1.558 1.606 1.662 1.728 1.808 1.906 2.029 2.187 2.398 2.695 3.142 3.889 5.389 9.90
0.00291 0.00309 0.00729 0.01547 0.03008 0.05432 0.06066 0.06758 0.07511 0.08329 0.09216 0.10132 0.10177 0.11214 0.12332 0.13536 0.14830 0.16218 0.17705 0.19295 0.20994 0.22806 0.24735 0.26787 0.28968 0.31281 0.33733 0.36328 0.39072 0.41970 0.45028 0.48252 0.51646 0.55218 0.58972 0.62915 0.67052 0.71390 0.75935 0.80694 0.85672 0.90875 0.96312 1.0199 1.0791 1.1408 1.2052 1.2722 1.3420 1.4146 1.4901 1.5685 1.6501 1.7347 1.8226 1.9138 2.0085 2.1067 2.2084 2.3140 2.4234 2.5368 2.6544 2.7763 2.9027 3.0339 3.1703 3.3121 3.4602 3.6177
1690.7 1688.7 1656.2 1623.4 1589.9 1555.7 1548.7 1541.7 1534.7 1527.6 1520.5 1513.6 1513.3 1506.0 1498.8 1491.4 1484.0 1476.6 1469.1 1461.5 1453.8 1446.1 1438.4 1430.5 1422.6 1414.5 1406.4 1398.3 1390.0 1381.6 1373.1 1364.5 1355.8 1347.0 1338.1 1329.0 1319.8 1310.5 1301.0 1291.3 1281.5 1271.5 1261.3 1250.9 1240.3 1229.4 1218.3 1206.9 1195.3 1183.3 1171.0 1158.4 1145.4 1131.9 1117.9 1103.5 1088.4 1072.7 1056.2 1038.9 1020.6 1001.1 980.2 957.5 932.6 904.5 872.1 832.4 777.5 573.6
4.0880 3.8709 1.7303 0.85534 0.45942 0.26432 0.23836 0.21542 0.19510 0.17706 0.16100 0.14728 0.14668 0.13387 0.12240 0.11209 0.10283 0.09448 0.08693 0.08011 0.07393 0.06831 0.06321 0.05855 0.05431 0.05043 0.04688 0.04363 0.04064 0.03789 0.03536 0.03303 0.03088 0.02890 0.02706 0.02535 0.02377 0.02230 0.02093 0.01966 0.01848 0.01737 0.01634 0.01537 0.01447 0.01362 0.01283 0.01208 0.01138 0.01072 0.01010 0.00951 0.00895 0.00843 0.00793 0.00746 0.00702 0.00659 0.00619 0.00580 0.00543 0.00507 0.00472 0.00439 0.00406 0.00373 0.00340 0.00305 0.00265 0.00174
*Temperatures on ITS-90 scale
87.13 87.78 98.18 108.70 119.36 130.19 132.38 134.57 136.78 138.99 141.21 143.34 143.44 145.68 147.92 150.18 152.44 154.71 157.00 159.29 161.59 163.90 166.22 168.56 170.90 173.26 175.62 178.00 180.39 182.80 185.21 187.64 190.08 192.54 195.01 197.50 200.00 202.52 205.05 207.60 210.17 212.76 215.37 218.00 220.65 223.32 226.02 228.74 231.49 234.26 237.07 239.91 242.78 245.69 248.64 251.63 254.67 257.76 260.92 264.14 267.44 270.83 274.33 277.95 281.75 285.77 290.10 294.95 300.86 318.06
277.39 277.74 283.36 289.06 294.83 300.60 301.75 302.91 304.06 305.20 306.35 307.44 307.49 308.63 309.77 310.90 312.03 313.16 314.28 315.40 316.51 317.61 318.71 319.80 320.88 321.96 323.03 324.09 325.14 326.19 327.22 328.24 329.25 330.25 331.23 332.20 333.16 334.10 335.02 335.92 336.80 337.66 338.50 339.31 340.10 340.85 341.58 342.28 342.95 343.57 344.16 344.71 345.22 345.67 346.07 346.42 346.69 346.90 347.02 347.05 346.96 346.75 346.38 345.82 345.00 343.85 342.21 339.79 335.77 318.06
1.5931 1.5911 1.5634 1.5421 1.5257 1.5135 1.5114 1.5095 1.5077 1.5060 1.5044 1.5030 1.5029 1.5016 1.5003 1.4991 1.4980 1.4969 1.4960 1.4951 1.4943 1.4935 1.4928 1.4922 1.4916 1.4911 1.4906 1.4901 1.4897 1.4894 1.4890 1.4887 1.4884 1.4882 1.4879 1.4877 1.4875 1.4873 1.4870 1.4868 1.4866 1.4863 1.4860 1.4857 1.4854 1.4850 1.4846 1.4842 1.4836 1.4831 1.4824 1.4817 1.4809 1.4799 1.4789 1.4778 1.4764 1.4750 1.4733 1.4714 1.4692 1.4667 1.4638 1.4604 1.4563 1.4512 1.4448 1.4362 1.4230 1.3696
aTriple
point
Velocity of Sound, m/s Liquid Vapor 933 929 878 827 779 731 721 712 702 693 683 674 674 664 655 645 636 627 617 608 599 589 580 570 561 552 542 533 523 514 505 495 486 476 467 457 448 439 429 420 410 400 391 381 371 362 352 342 332 322 312 302 292 282 272 261 251 240 229 218 207 196 184 172 159 146 132 116 100 0
116.4 116.6 119.4 121.9 124.1 126.0 126.3 126.6 126.9 127.2 127.4 127.6 127.6 127.8 128.0 128.2 128.3 128.5 128.6 128.7 128.7 128.7 128.7 128.7 128.7 128.6 128.5 128.4 128.2 128.0 127.8 127.5 127.3 126.9 126.6 126.2 125.8 125.3 124.9 124.3 123.8 123.2 122.5 121.9 121.1 120.4 119.6 118.7 117.8 116.8 115.8 114.8 113.7 112.5 111.3 109.9 108.6 107.1 105.6 104.0 102.3 100.5 98.6 96.6 94.5 92.2 89.8 87.3 84.4 0.0
bNormal
Liquid Vapor
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C
1152 1134 890.8 720.6 595.2 499.3 482.8 467.0 452.0 437.6 423.8 411.1 410.5 397.9 385.7 374.0 362.8 352.0 341.6 331.6 321.9 312.6 303.6 294.9 286.5 278.4 270.5 262.9 255.5 248.4 241.4 234.7 228.1 221.8 215.6 209.6 203.7 198.0 192.4 187.0 181.7 176.5 171.4 166.5 161.6 156.9 152.2 147.6 143.1 138.7 134.4 130.1 125.9 121.7 117.6 113.5 109.4 105.4 101.3 97.3 93.3 89.2 85.1 80.9 76.5 71.9 67.1 61.6 54.8 —
116.0 115.7 111.0 106.3 101.6 96.8 95.9 95.0 94.0 93.1 92.2 91.3 91.3 90.3 89.4 88.5 87.6 86.7 85.7 84.8 83.9 83.0 82.1 81.2 80.3 79.4 78.5 77.7 76.8 75.9 75.0 74.1 73.3 72.4 71.6 70.7 69.8 69.0 68.1 67.3 66.5 65.6 64.8 64.0 63.1 62.3 61.5 60.7 59.8 59.0 58.2 57.4 56.6 55.8 54.9 54.1 53.3 52.5 51.7 50.9 50.1 49.3 48.5 47.7 47.0 46.3 45.9 46.0 47.8
Viscosity, Pa·s
boiling point
7.43 7.46 7.90 8.34 8.77 9.20 9.29 9.37 9.46 9.54 9.63 9.71 9.71 9.80 9.89 9.97 10.06 10.14 10.23 10.32 10.40 10.49 10.58 10.66 10.75 10.84 10.93 11.02 11.11 11.21 11.30 11.40 11.49 11.59 11.69 11.79 11.89 12.00 12.11 12.22 12.33 12.45 12.57 12.70 12.83 12.96 13.10 13.25 13.40 13.57 13.74 13.92 14.11 14.32 14.54 14.77 15.03 15.31 15.61 15.95 16.32 16.75 17.23 17.79 18.45 19.25 20.27 21.67 23.90 —
5.23 5.28 5.92 6.58 7.25 7.92 8.06 8.20 8.33 8.47 8.61 8.75 8.75 8.89 9.03 9.17 9.32 9.46 9.61 9.75 9.90 10.05 10.20 10.35 10.50 10.65 10.81 10.97 11.12 11.29 11.45 11.62 11.79 11.96 12.14 12.32 12.50 12.69 12.88 13.08 13.29 13.50 13.72 13.95 14.18 14.43 14.69 14.96 15.24 15.55 15.87 16.21 16.57 16.97 17.39 17.86 18.37 18.93 19.55 20.25 21.04 21.95 23.02 24.29 25.83 27.80 30.44 34.37 41.79
21.79 –100.63 21.69 –100 20.08 –90 18.50 –80 16.94 –70 15.41 –60 15.11 –58 14.81 –56 14.51 –54 14.21 –52 13.91 –50 13.63 –48.09 13.61 –48 13.32 –46 13.02 –44 12.73 –42 12.44 –40 12.15 –38 11.86 –36 11.57 –34 11.29 –32 11.00 –30 10.72 –28 10.44 –26 10.15 –24 9.88 –22 9.60 –20 9.32 –18 9.05 –16 8.78 –14 8.50 –12 8.23 –10 7.97 –8 7.70 –6 7.44 –4 7.17 –2 6.91 0 6.65 2 6.40 4 6.14 6 5.89 8 5.64 10 5.39 12 5.14 14 4.90 16 4.66 18 4.42 20 4.18 22 3.95 24 3.72 26 3.49 28 3.26 30 3.04 32 2.82 34 2.60 36 2.39 38 2.18 40 1.97 42 1.77 44 1.57 46 1.38 48 1.19 50 1.01 52 0.84 54 0.67 56 0.51 58 0.36 60 0.22 62 0.09 64 0.00 66.02 cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 8
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 134a
30.16
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.17
Refrigerant 134a (1,1,1,2-Tetrafluoroethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –103.30a –100 –90 –80 –70 –60 –50 –40 –30 –28 –26.07b –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 95 100 101.06c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00039 0.00056 0.00152 0.00367 0.00798 0.01591 0.02945 0.05121 0.08438 0.09270 0.10133 0.10167 0.11130 0.12165 0.13273 0.14460 0.15728 0.17082 0.18524 0.20060 0.21693 0.23428 0.25268 0.27217 0.29280 0.31462 0.33766 0.36198 0.38761 0.41461 0.44301 0.47288 0.50425 0.53718 0.57171 0.60789 0.64578 0.68543 0.72688 0.77020 0.81543 0.86263 0.91185 0.96315 1.0166 1.0722 1.1301 1.1903 1.2529 1.3179 1.3854 1.4555 1.5282 1.6036 1.6818 1.7628 1.8467 1.9337 2.0237 2.1168 2.2132 2.3130 2.4161 2.5228 2.6332 2.9258 3.2442 3.5912 3.9724 4.0593
1591.1 1582.4 1555.8 1529.0 1501.9 1474.3 1446.3 1417.7 1388.4 1382.4 1376.7 1376.5 1370.4 1364.4 1358.3 1352.1 1345.9 1339.7 1333.4 1327.1 1320.8 1314.3 1307.9 1301.4 1294.8 1288.1 1281.4 1274.7 1267.9 1261.0 1254.0 1246.9 1239.8 1232.6 1225.3 1218.0 1210.5 1202.9 1195.2 1187.5 1179.6 1171.6 1163.4 1155.1 1146.7 1138.2 1129.5 1120.6 1111.5 1102.3 1092.9 1083.2 1073.4 1063.2 1052.9 1042.2 1031.2 1020.0 1008.3 996.2 983.8 970.8 957.3 943.1 928.2 887.2 837.8 772.7 651.2 511.9
35.4960 25.1930 9.7698 4.2682 2.0590 1.0790 0.60620 0.36108 0.22594 0.20680 0.19018 0.18958 0.17407 0.16006 0.14739 0.13592 0.12551 0.11605 0.10744 0.09959 0.09242 0.08587 0.07987 0.07436 0.06931 0.06466 0.06039 0.05644 0.05280 0.04944 0.04633 0.04345 0.04078 0.03830 0.03600 0.03385 0.03186 0.03000 0.02826 0.02664 0.02513 0.02371 0.02238 0.02113 0.01997 0.01887 0.01784 0.01687 0.01595 0.01509 0.01428 0.01351 0.01278 0.01209 0.01144 0.01083 0.01024 0.00969 0.00916 0.00865 0.00817 0.00771 0.00727 0.00685 0.00645 0.00550 0.00461 0.00374 0.00268 0.00195
*Temperatures on ITS-90 scale
71.46 75.36 87.23 99.16 111.20 123.36 135.67 148.14 160.79 163.34 165.81 165.90 168.47 171.05 173.64 176.23 178.83 181.44 184.07 186.70 189.34 191.99 194.65 197.32 200.00 202.69 205.40 208.11 210.84 213.58 216.33 219.09 221.87 224.66 227.47 230.29 233.12 235.97 238.84 241.72 244.62 247.54 250.48 253.43 256.41 259.41 262.43 265.47 268.53 271.62 274.74 277.89 281.06 284.27 287.50 290.78 294.09 297.44 300.84 304.28 307.78 311.33 314.94 318.63 322.39 332.22 342.93 355.25 373.30 389.64
334.94 336.85 342.76 348.83 355.02 361.31 367.65 374.00 380.32 381.57 382.78 382.82 384.07 385.32 386.55 387.79 389.02 390.24 391.46 392.66 393.87 395.06 396.25 397.43 398.60 399.77 400.92 402.06 403.20 404.32 405.43 406.53 407.61 408.69 409.75 410.79 411.82 412.84 413.84 414.82 415.78 416.72 417.65 418.55 419.43 420.28 421.11 421.92 422.69 423.44 424.15 424.83 425.47 426.07 426.63 427.14 427.61 428.02 428.36 428.65 428.86 429.00 429.04 428.98 428.81 427.76 425.42 420.67 407.68 389.64
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid
Vapor
0.4126 0.4354 0.5020 0.5654 0.6262 0.6846 0.7410 0.7956 0.8486 0.8591 0.8690 0.8694 0.8798 0.8900 0.9002 0.9104 0.9205 0.9306 0.9407 0.9506 0.9606 0.9705 0.9804 0.9902 1.0000 1.0098 1.0195 1.0292 1.0388 1.0485 1.0581 1.0677 1.0772 1.0867 1.0962 1.1057 1.1152 1.1246 1.1341 1.1435 1.1529 1.1623 1.1717 1.1811 1.1905 1.1999 1.2092 1.2186 1.2280 1.2375 1.2469 1.2563 1.2658 1.2753 1.2848 1.2944 1.3040 1.3137 1.3234 1.3332 1.3430 1.3530 1.3631 1.3733 1.3836 1.4104 1.4390 1.4715 1.5188 1.5621
1.9639 1.184 0.585 1.9456 1.184 0.593 1.8972 1.189 0.617 1.8580 1.198 0.642 1.8264 1.210 0.667 1.8010 1.223 0.692 1.7806 1.238 0.720 1.7643 1.255 0.749 1.7515 1.273 0.781 1.7492 1.277 0.788 1.7472 1.281 0.794 1.7471 1.281 0.794 1.7451 1.285 0.801 1.7432 1.289 0.809 1.7413 1.293 0.816 1.7396 1.297 0.823 1.7379 1.302 0.831 1.7363 1.306 0.838 1.7348 1.311 0.846 1.7334 1.316 0.854 1.7320 1.320 0.863 1.7307 1.325 0.871 1.7294 1.330 0.880 1.7282 1.336 0.888 1.7271 1.341 0.897 1.7260 1.347 0.906 1.7250 1.352 0.916 1.7240 1.358 0.925 1.7230 1.364 0.935 1.7221 1.370 0.945 1.7212 1.377 0.956 1.7204 1.383 0.967 1.7196 1.390 0.978 1.7188 1.397 0.989 1.7180 1.405 1.001 1.7173 1.413 1.013 1.7166 1.421 1.025 1.7159 1.429 1.038 1.7152 1.437 1.052 1.7145 1.446 1.065 1.7138 1.456 1.080 1.7131 1.466 1.095 1.7124 1.476 1.111 1.7118 1.487 1.127 1.7111 1.498 1.145 1.7103 1.510 1.163 1.7096 1.523 1.182 1.7089 1.537 1.202 1.7081 1.551 1.223 1.7072 1.566 1.246 1.7064 1.582 1.270 1.7055 1.600 1.296 1.7045 1.618 1.324 1.7035 1.638 1.354 1.7024 1.660 1.387 1.7013 1.684 1.422 1.7000 1.710 1.461 1.6987 1.738 1.504 1.6972 1.769 1.552 1.6956 1.804 1.605 1.6939 1.843 1.665 1.6920 1.887 1.734 1.6899 1.938 1.812 1.6876 1.996 1.904 1.6850 2.065 2.012 1.6771 2.306 2.397 1.6662 2.756 3.121 1.6492 3.938 5.020 1.6109 17.59 25.35 1.5621 a Triple
Liquid Vapor
point
Thermal Cond., Velocity of Viscosity, Surface Pa·s mW/(m·K) Sound, m/s cp /cv Tension, Temp.,* Vapor Liquid Vapor Liquid Vapor Liquid Vapor mN/m °C 1.164 1.162 1.156 1.151 1.148 1.146 1.146 1.148 1.152 1.153 1.154 1.154 1.155 1.156 1.158 1.159 1.161 1.163 1.165 1.167 1.169 1.171 1.174 1.176 1.179 1.182 1.185 1.189 1.192 1.196 1.200 1.204 1.209 1.214 1.219 1.224 1.230 1.236 1.243 1.249 1.257 1.265 1.273 1.282 1.292 1.303 1.314 1.326 1.339 1.354 1.369 1.386 1.405 1.425 1.448 1.473 1.501 1.532 1.567 1.607 1.653 1.705 1.766 1.838 1.924 2.232 2.820 4.369 20.81
1120 1103 1052 1002 952 903 855 807 760 751 742 742 732 723 714 705 695 686 677 668 658 649 640 631 622 612 603 594 585 576 566 557 548 539 530 520 511 502 493 483 474 465 455 446 436 427 418 408 399 389 379 370 360 350 340 331 321 311 301 290 280 269 259 248 237 207 176 141 101 0
126.8 127.9 131.0 134.0 136.8 139.4 141.7 143.6 145.2 145.4 145.7 145.7 145.9 146.1 146.3 146.4 146.6 146.7 146.8 146.9 146.9 147.0 147.0 147.0 146.9 146.9 146.8 146.7 146.5 146.4 146.2 146.0 145.7 145.5 145.1 144.8 144.5 144.1 143.6 143.2 142.7 142.1 141.6 141.0 140.3 139.7 138.9 138.2 137.4 136.6 135.7 134.7 133.8 132.7 131.7 130.5 129.4 128.1 126.8 125.5 124.0 122.6 121.0 119.4 117.7 113.1 107.9 101.9 94.0 0.0 b Normal
2175.0 1893.0 1339.0 1018.0 809.2 663.1 555.1 472.2 406.4 394.9 384.2 383.8 373.1 362.9 353.0 343.5 334.3 325.4 316.9 308.6 300.6 292.9 285.4 278.1 271.1 264.3 257.6 251.2 244.9 238.8 232.9 227.1 221.5 216.0 210.7 205.5 200.4 195.4 190.5 185.8 181.1 176.6 172.1 167.7 163.4 159.2 155.1 151.0 147.0 143.1 139.2 135.4 131.6 127.9 124.2 120.6 117.0 113.5 109.9 106.4 102.9 99.5 96.0 92.5 89.0 80.2 70.9 60.4 45.1 —
6.46 6.60 7.03 7.46 7.89 8.30 8.72 9.12 9.52 9.60 9.68 9.68 9.77 9.85 9.92 10.01 10.09 10.17 10.25 10.33 10.41 10.49 10.57 10.65 10.73 10.81 10.90 10.98 11.06 11.15 11.23 11.32 11.40 11.49 11.58 11.67 11.76 11.85 11.95 12.04 12.14 12.24 12.34 12.44 12.55 12.65 12.76 12.88 13.00 13.12 13.24 13.37 13.51 13.65 13.79 13.95 14.11 14.28 14.46 14.65 14.85 15.07 15.30 15.56 15.84 16.67 17.81 19.61 24.21 —
boiling point
145.2 143.2 137.3 131.5 126.0 120.7 115.6 110.6 105.8 104.8 103.9 103.9 102.9 102.0 101.1 100.1 99.2 98.3 97.4 96.5 95.6 94.7 93.8 92.9 92.0 91.1 90.2 89.4 88.5 87.6 86.7 85.9 85.0 84.1 83.3 82.4 81.6 80.7 79.8 79.0 78.1 77.3 76.4 75.6 74.7 73.9 73.0 72.1 71.3 70.4 69.6 68.7 67.8 67.0 66.1 65.2 64.3 63.4 62.6 61.7 60.8 59.9 59.0 58.1 57.2 54.9 52.8 51.7 59.9
3.08 3.34 4.15 4.95 5.75 6.56 7.36 8.17 8.99 9.15 9.31 9.32 9.48 9.65 9.82 9.98 10.15 10.32 10.49 10.66 10.83 11.00 11.17 11.34 11.51 11.69 11.86 12.04 12.22 12.40 12.58 12.77 12.95 13.14 13.33 13.53 13.72 13.92 14.13 14.33 14.54 14.76 14.98 15.21 15.44 15.68 15.93 16.18 16.45 16.72 17.01 17.31 17.63 17.96 18.31 18.68 19.07 19.50 19.95 20.45 20.98 21.56 22.21 22.92 23.72 26.22 29.91 36.40 60.58
28.07 –103.30 27.50 –100 25.79 –90 24.10 –80 22.44 –70 20.80 –60 19.18 –50 17.60 –40 16.04 –30 15.73 –28 15.44 –26.07 15.43 –26 15.12 –24 14.82 –22 14.51 –20 14.21 –18 13.91 –16 13.61 –14 13.32 –12 13.02 –10 12.72 –8 12.43 –6 12.14 –4 11.85 –2 11.56 0 11.27 2 10.99 4 10.70 6 10.42 8 10.14 10 9.86 12 9.58 14 9.30 16 9.03 18 8.76 20 8.48 22 8.21 24 7.95 26 7.68 28 7.42 30 7.15 32 6.89 34 6.64 36 6.38 38 6.13 40 5.88 42 5.63 44 5.38 46 5.13 48 4.89 50 4.65 52 4.41 54 4.18 56 3.95 58 3.72 60 3.49 62 3.27 64 3.05 66 2.83 68 2.61 70 2.40 72 2.20 74 1.99 76 1.80 78 1.60 80 1.14 85 0.71 90 0.33 95 0.04 100 0.00 101.06 c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
30.18
2017 ASHRAE Handbook—Fundamentals (SI) Refrigerant 134a Properties of Superheated Vapor Pressure = 0.101325 MPa Saturation temperature = 26.07°C
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kJ/kg kJ/(kg·K) Sound, m/s kg/m3
Saturated Liquid 1374.34 166.07 0.8701 Vapor 5.26 382.90 1.7476 –20 5.11 387.68 1.7667 –10 4.89 395.65 1.7976 0 4.69 403.74 1.8278 10 4.50 411.97 1.8574 20 4.34 420.34 1.8864 30 4.18 428.85 1.9150 40 4.04 437.52 1.9431 50 3.91 446.33 1.9708 60 3.78 455.30 1.9981 70 3.67 464.43 2.0251 80 3.56 473.70 2.0518 90 3.46 483.13 2.0781 100 3.36 492.71 2.1041 110 3.27 502.44 2.1298 120 3.19 512.32 2.1553 130 3.11 522.35 2.1805 140 3.03 532.52 2.2054 150 2.96 542.83 2.2301 Pressure = 0.600 MPa Saturation temperature = 21.58°C Temp.,* °C
747.1 145.7 147.8 151.0 154.2 157.2 160.1 162.9 165.7 168.4 171.0 173.6 176.1 178.6 181.0 183.4 185.7 188.1 190.3 192.6
Density, Enthalpy, Entropy, Vel. of kJ/kg kJ/(kg·K) Sound, m/s kg/m3
Saturated Liquid 1219.08 229.62 1.1035 Vapor 29.13 410.67 1.7178 30 27.79 418.97 1.7455 40 26.41 428.72 1.7772 50 25.21 438.44 1.8077 60 24.16 448.16 1.8374 70 23.22 457.93 1.8662 80 22.37 467.75 1.8944 90 21.59 477.65 1.9221 100 20.88 487.64 1.9492 110 20.22 497.72 1.9759 120 19.61 507.92 2.0022 130 19.04 518.22 2.0280 140 18.51 528.63 2.0536 150 18.01 539.17 2.0787 160 17.54 549.82 2.1036 170 17.10 560.59 2.1282 180 16.68 571.48 2.1525 190 16.29 582.50 2.1766 200 15.91 593.63 2.2003 Pressure = 1.200 MPa Saturation temperature = 46.32°C Temp.,* °C
Pressure = 0.200 MPa Saturation temperature = 10.07°C
524.0 145.0 149.0 153.4 157.4 161.2 164.7 168.0 171.2 174.3 177.3 180.1 182.9 185.6 188.2 190.8 193.3 195.8 198.2 200.6
Density, Enthalpy, Entropy, Vel. of kJ/kg kJ/(kg·K) Sound, m/s kg/m3
Saturated Liquid 1118.89 265.91 Vapor 59.73 422.22 50 58.09 426.51 60 54.32 437.83 70 51.26 448.81 80 48.69 459.61 90 46.49 470.30 100 44.55 480.94 110 42.83 491.58 120 41.28 502.25 130 39.87 512.95 140 38.58 523.72 150 37.39 534.56 160 36.29 545.48 170 35.26 556.50 180 34.31 567.60 190 33.40 578.80 200 32.56 590.11 210 31.76 601.51 220 31.01 613.02 230 30.29 624.64 240 29.61 636.36 250 28.96 648.18 *Temperatures on ITS-90 scale
1.2200 1.7092 1.7226 1.7571 1.7896 1.8206 1.8504 1.8794 1.9075 1.9350 1.9619 1.9882 2.0142 2.0397 2.0648 2.0896 2.1141 2.1382 2.1621 2.1856 2.2090 2.2320 2.2548
405.0 138.2 140.7 146.9 152.3 157.1 161.5 165.6 169.4 173.0 176.4 179.7 182.8 185.8 188.8 191.6 194.4 197.1 199.7 202.3 204.8 207.2 209.7
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid 1325.78 Vapor 10.01 –10 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150
Temp.,* °C
Temp.,* °C
0.9506 1.7337
672.8 146.9 147.0 150.6 154.0 157.3 160.4 163.4 166.3 169.2 171.9 174.6 177.2 179.7 182.2 184.7 187.1 189.4 191.7
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s 243.58 415.58
1.1495 1.7144
36.98 424.61 1.7437 35.03 434.85 1.7758 33.36 444.98 1.8067 31.90 455.08 1.8366 30.62 465.17 1.8656 29.46 475.30 1.8939 28.41 485.49 1.9215 27.46 495.74 1.9486 26.58 506.07 1.9753 25.77 516.50 215 25.01 527.03 2.0272 24.31 537.66 2.0527 23.65 548.40 2.0777 23.03 559.24 2.1025 22.45 570.20 2.1270 21.89 581.28 2.1511 21.37 592.46 2.1750 Pressure = 1.400 MPa Saturation temperature = 52.43°C
477.4 142.9 147.6 152.4 156.8 160.8 164.6 168.1 171.5 174.7 177.8 180.8 183.7 186.4 189.2 191.8 194.4 196.9 199.4
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid 1090.50 Vapor 70.76 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
186.69 392.71
10.01 392.77 1.7339 9.54 401.21 1.7654 9.13 409.73 1.7961 8.76 418.35 1.8260 8.42 427.07 1.8552 8.12 435.90 1.8839 7.83 444.87 1.9121 7.57 453.97 1.9398 7.33 463.20 1.9671 7.11 472.57 1.9940 6.89 482.08 2.0206 6.70 491.74 2.0468 6.51 501.53 2.0727 6.34 511.47 2.0983 6.17 521.55 2.1236 6.01 531.76 2.1486 5.87 542.12 2.1734 Pressure = 0.800 MPa Saturation temperature = 31.33°C
Saturated Liquid 1181.92 Vapor 38.99 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200
Pressure = 0.400 MPa Saturation temperature = 8.94°C
66.61 62.25 58.74 55.79 53.24 51.03 49.05 47.28 45.67 44.19 42.83 41.57 40.41 39.31 38.28 37.32 36.41 35.55 34.73 33.96
275.38 424.50
1.2488 1.7068
375.1 135.6
433.69 445.31 456.56 467.60 478.53 489.39 500.25 511.11 522.02 532.97 544 555.10 566.28 577.55 588.92 600.38 611.94 623.60 635.35 647.22
1.7347 1.7691 1.8014 1.8322 1.8619 1.8906 1.9186 1.9459 1.9726 1.9988 2.0246 2.0499 2.0748 2.0994 2.1237 2.1477 2.1714 2.1948 2.2179 2.2408
141.2 147.5 153.0 158.0 162.5 166.6 170.5 174.2 177.7 181.0 184.2 187.2 190.2 193.1 195.9 198.6 201.3 203.9 206.4 208.9
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid 1263.84 Vapor 19.52
10 20 30 40 50 60 70 80 90 100 110 120 130 140 150
Temp.,* °C
Temp.,* °C
583.8 146.6
147.0 151.2 155.0 158.6 162.0 165.3 168.4 171.4 174.3 177.1 179.8 182.4 185.0 187.5 190.0
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s 255.44 419.31
1.1874 1.7117
48.95 419.99 1.7139 45.86 430.91 1.7482 43.34 441.56 1.7807 41.21 452.05 1.8117 39.36 462.47 1.8416 37.74 472.86 1.8706 36.29 483.26 1.8989 34.99 493.69 1.9265 33.80 504.19 1.9535 32.71 514.75 1.9800 31.70 525.39 2.0061 30.76 536.12 2.0318 29.90 546.95 2.0571 29.08 557.88 2.0820 28.32 568.91 2.1066 27.60 580.05 2.1309 26.92 591.29 2.1550 Pressure = 1.600 MPa Saturation temperature = 57.91°C
438.6 140.6 141.0 146.9 152.0 156.7 160.9 164.9 168.6 172.1 175.4 178.6 181.7 184.6 187.5 190.3 193.0 195.6 198.2
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid 1063.28 Vapor 82.34 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
1.0432 1.7229
19.41 404.78 1.7263 18.45 414.00 1.7583 17.61 423.21 1.7892 16.87 432.46 1.8192 16.20 441.76 1.8485 15.60 451.15 1.8771 15.05 460.63 1.9051 14.54 470.21 1.9326 14.08 479.91 1.9597 13.65 489.72 1.9864 13.24 499.65 2.0126 12.87 509.71 2.0386 12.51 519.90 2.0641 12.18 530.21 2.0894 11.87 540.66 2.1144 Pressure = 1.000 MPa Saturation temperature = 39.39°C
Saturated Liquid 1149.06 Vapor 49.16 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200
212.08 403.80
80.74 74.43 69.61 65.71 62.43 59.62 57.14 54.95 52.98 51.18 49.54 48.03 46.63 45.32 44.10 42.96 41.88 40.87 39.91 39.00
284.11 426.27
1.2748 1.7042
348.1 132.9
428.99 441.47 453.30 464.76 476.01 487.13 498.19 509.23 520.28 531.36 542.49 553.68 564.94 576.29 587.71 599.23 610.84 622.55 634.35 646.25
1.7124 1.7493 1.7833 1.8153 1.8458 1.8753 1.9038 1.9315 1.9586 1.9851 2.0111 2.0366 2.0617 2.0865 2.1109 2.1350 2.1588 2.1823 2.2055 2.2285
134.7 142.3 148.7 154.2 159.2 163.8 168.0 171.9 175.6 179.1 182.5 185.7 188.8 191.8 194.7 197.6 200.3 203.0 205.6 208.2
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.19
Refrigerant 134a Properties of Superheated Vapor (Concluded)
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure = 1.800 MPa Saturation temperature = 62.90°C
Pressure = 2.000 MPa Saturation temperature = 67.49°C
Pressure = 2.200 MPa Saturation temperature = 71.74°C
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid Vapor 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
1036.81 94.53 88.23 81.54 76.38 72.17 68.64 65.60 62.91 60.53 58.37 56.42 54.62 52.97 51.44 50.01 48.68 47.43 46.25 45.14 44.09
Saturated Liquid Vapor 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
1010.74 107.46 104.37 94.85 87.97 82.58 78.17 74.44 71.18 68.33 65.78 63.47 61.37 59.45 57.67 56.02 54.49 53.05 51.70 50.43 49.23
292.26 427.59 437.17 449.76 461.74 473.36 484.78 496.06 507.29 518.50 529.71 540.95 552.24 563.59 575.01 586.50 598.08 609.74 621.50 633.34 645.28
1.2987 1.7014 1.7296 1.7657 1.7992 1.8308 1.8610 1.8900 1.9183 1.9457 1.9725 1.9988 2.0246 2.0499 2.0748 2.0993 2.1236 2.1475 2.1710 2.1944 2.2174
323.2 130.1 136.5 144.0 150.3 155.9 160.8 165.4 169.6 173.5 177.3 180.8 184.2 187.4 190.6 193.6 196.5 199.4 202.1 204.9 207.5
Pressure =2.400 MPa Saturation temperature = 75.70°C Temp.,* °C Saturated Liquid Vapor 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
314.40 429.27 436.42 451.12 464.44 477.04 489.22 501.14 512.90 524.57 536.20 547.82 559.45 571.11 582.82 594.58 606.41 618.31 630.29 642.35
1.3616 1.6908 1.7112 1.7523 1.7885 1.8218 1.8532 1.8831 1.9119 1.9398 1.9670 1.9935 2.0195 2.0449 2.0699 2.0945 2.1188 2.1427 2.1662 2.1895
257.9 121.4 126.9 137.0 144.8 151.5 157.2 162.4 167.2 171.6 175.7 179.6 183.3 186.8 190.2 193.4 196.6 199.6 202.5 205.4
Temp.,* °C Saturated Liquid Vapor 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
Pressure = 3.000 MPa Saturation temperature = 86.20°C Temp.,* °C Saturated Liquid Vapor 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300
334.75 427.47 435.84 453.20 467.93 481.47 494.36 506.86 519.11 531.21 543.21 555.16 567.10 579.05 591.02 603.03 615.10 627.22 639.41 651.66 664 676.41 688.89 701.46
*Temperatures on ITS-90 scale
1.4171 1.6752 1.6983 1.7455 1.7845 1.8194 1.8518 1.8824 1.9117 1.9399 1.9673 1.9940 2.0201 2.0456 2.0706 2.0952 2.1195 2.1433 2.1668 2.1900 2.2130 2.2356 2.2580 2.2801
300.1 127.2 129.9 138.9 146.2 152.4 157.8 162.7 167.2 171.4 175.4 179.1 182.6 186.0 189.3 192.4 195.5 198.4 201.3 204.1 206.8
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s 931.88 152.12 150.48 131.08 119.15 110.50 103.72 98.17 93.46 89.39 85.80 82.59 79.70 77.07 74.65 72.43 70.36 68.44 66.64 64.95
321.29 429.08 430.22 446.81 461.03 474.19 486.75 498.96 510.94 522.79 534.57 546.30 558.04 569.79 581.57 593.40 605.29 617.24 629.27 641.37
1.3806 1.6863 1.6895 1.7359 1.7745 1.8093 1.8417 1.8724 1.9017 1.9301 1.9576 1.9844 2.0106 2.0362 2.0614 2.0861 2.1105 2.1345 2.1581 2.1815
200.4 112.2 119.1 131.8 141.0 148.5 155.0 160.7 165.9 170.6 175.0 179.2 183.1 186.8 190.4 193.8 197.2 200.4 203.4 206.5 209.4 212.2 215.0 217.8
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid Vapor
984.76 121.25
307.32 429.08
1.3417 1.6948
278.4 124.3
80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
110.03 100.70 93.78 88.25 83.70 79.79 76.41 73.40 70.71 68.28 66.06 64.02 62.13 60.38 58.74 57.21 55.77 54.42
441.49 454.98 467.61 479.75 491.59 503.25 514.81 526.32 537.81 549.31 560.84 572.42 584.06 595.76 607.53 619.38 631.31 643.33
1.7303 1.7680 1.8023 1.8344 1.8649 1.8942 1.9226 1.9501 1.9769 2.0032 2.0289 2.0542 2.0790 2.1035 2.1276 2.1514 2.1749 2.1981
133.3 141.8 148.7 154.7 160.0 164.9 169.3 173.5 177.4 181.1 184.6 188.0 191.3 194.4 197.5 200.4 203.3 206.1
Pressure = 2.800 MPa Saturation temperature = 82.90°C
238.2 118.3 119.3 131.7 140.8 148.1 154.4 160.0 165.0 169.7 174.0 178.1 181.9 185.5 189.0 192.4 195.6 198.8 201.8 204.8
Pressure = 4.000 MPa Saturation temperature = 100.35°C
Density, Enthalpy, Entropy, Vel. of kJ/kg kJ/(kg·K) Sound, m/s kg/m3 875.30 189.25 173.82 150.47 136.36 126.23 118.34 111.89 106.45 101.75 97.62 93.94 90.62 87.61 84.84 82.30 79.94 77.74 75.69 73.77 71.96 70.25 68.63 67.10
1.3209 1.6983 1.7091 1.7483 1.7835 1.8164 1.8474 1.8772 1.9059 1.9338 1.9609 1.9875 2.0135 2.0390 2.0641 2.0888 2.1131 2.1371 2.1608 2.1842 2.2073
Pressure = 2.600 MPa Saturation temperature = 79.41°C
Density, Enthalpy, Entropy, Vel. of kJ/kg kJ/(kg·K) Sound, m/s kg/m3 958.58 136.07 127.96 114.90 105.89 99.00 93.44 88.79 84.77 81.27 78.15 75.35 72.81 70.48 68.34 66.36 64.51 62.79 61.18 59.66
299.96 428.52 432.22 445.86 458.49 470.57 482.32 493.86 505.30 516.68 528.03 539.39 550.79 562.23 573.72 585.28 596.92 608.64 620.44 632.33 644.30
Temp.,* °C
Temp.,* °C
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid Vapor
904.29 169.71
328.05 428.50
1.3990 1.6812
219.1 115.3
90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250
150.13 133.85 122.89 114.63 108.00 102.49 97.78 93.66 90.01 86.74 83.78 81.08 78.59 76.29 74.15 72.16 70.30
441.84 457.32 471.16 484.17 496.70 508.93 520.97 532.90 544.77 556.61 568.45 580.31 592.21 604.16 616.17 628.25 640.39
1.7183 1.7603 1.7970 1.8305 1.8620 1.8919 1.9207 1.9486 1.9757 2.0021 2.0279 2.0533 2.0782 2.1027 2.1268 2.1505 2.1740
125.9 136.4 144.6 151.5 157.5 162.9 167.8 172.3 176.5 180.5 184.3 187.9 191.4 194.7 198.0 201.1 204.1
Pressure = 6.00 MPa Saturation temperature = n/a (supercritical)
Density, Enthalpy, Entropy, Vel. of kg/m3 kJ/kg kJ/(kg·K) Sound, m/s
Saturated Liquid Vapor
626.95 396.29
376.48 404.57
1.5272 1.6024
101.3 93.4
110 120 130 140 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300
233.68 199.79 179.83 165.73 154.89 146.10 138.74 132.41 126.88 121.97 117.55 113.56 109.90 106.55 103.44 100.56 97.87 95.35 92.98 90.75
446.28 465.29 481.11 495.51 509.13 522.25 535.07 547.69 560.17 572.58 584.95 597.30 609.66 622.05 634.47 646.93 659.45 672.03 684.67 697.38
1.7131 1.7621 1.8018 1.8371 1.8697 1.9004 1.9296 1.9578 1.9850 2.0115 2.0374 2.0627 2.0875 2.1119 2.1359 2.1595 2.1827 2.2057 2.2283 2.2507
119.8 132.5 142.0 149.7 156.4 162.4 167.8 172.7 177.4 181.7 185.8 189.7 193.4 197.0 200.5 203.8 207.1 210.2 213.3 216.2
Temp.,* °C
110 120 130 140 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300
Density, Enthalpy, Entropy, Vel. of kJ/kg kJ/(kg·K) Sound, m/s kg/m3
762.66 591.77 418.90 333.91 289.37 260.70 239.96 223.87 210.82 199.88 190.50 182.31 175.06 168.56 162.68 157.33 152.41 147.88 143.67 139.75
375.61 405.75 439.87 465.19 484.69 501.52 516.92 531.45 545.43 559.04 572.39 585.57 598.64 611.63 624.57 637.50 650.43 663.38 676.35 689.36
1.5174 1.5950 1.6807 1.7428 1.7894 1.8288 1.8639 1.8963 1.9269 1.9559 1.9839 2.0109 2.0371 2.0626 2.0876 2.1121 2.1361 2.1598 2.1830 2.2059
173.6 127.4 120.4 130.1 139.9 148.3 155.7 162.2 168.1 173.6 178.6 183.4 187.8 192.1 196.2 200.0 203.8 207.4 210.9 214.3
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 9 Pressure-Enthalpy Diagram for Refrigerant 143a
30.20
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.21
Refrigerant 143a (1,1,1-Trifluoroethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –111.81a –110 –100 –90 –80 –70 –60 –50 –48 –47.24b –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 65 70 72.71c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00107 0.00129 0.00333 0.00761 0.01572 0.02991 0.05307 0.08874 0.09773 0.10133 0.10742 0.11786 0.12907 0.14109 0.15398 0.16775 0.18247 0.19816 0.21488 0.23267 0.25156 0.27161 0.29286 0.31535 0.33915 0.36428 0.39081 0.41877 0.44823 0.47923 0.51182 0.54606 0.58199 0.61967 0.65916 0.70051 0.74378 0.78901 0.83628 0.88564 0.93714 0.99085 1.0468 1.1052 1.1659 1.2290 1.2947 1.3630 1.4340 1.5077 1.5842 1.6636 1.7460 1.8314 1.9200 2.0117 2.1068 2.2053 2.3073 2.4130 2.5224 2.6357 2.7530 2.8744 3.1977 3.5527 3.7610
1330.5 14.807 1326.2 12.430 1301.9 5.1127 1277.2 2.3596 1252.2 1.1971 1226.7 0.65675 1200.6 0.38446 1173.9 0.23754 1168.5 0.21695 1166.4 0.20971 1163.0 0.19849 1157.5 0.18191 1152.0 0.16697 1146.4 0.15350 1140.8 0.14133 1135.1 0.13031 1129.4 0.12032 1123.7 0.11124 1117.9 0.10297 1112.1 0.09544 1106.2 0.08857 1100.3 0.08228 1094.3 0.07652 1088.3 0.07125 1082.2 0.06640 1076.0 0.06194 1069.8 0.05784 1063.6 0.05405 1057.2 0.05056 1050.8 0.04733 1044.3 0.04434 1037.7 0.04158 1031.0 0.03901 1024.3 0.03662 1017.4 0.03440 1010.5 0.03234 1003.5 0.03042 996.3 0.02862 989.1 0.02695 981.7 0.02538 974.2 0.02392 966.5 0.02255 958.7 0.02126 950.8 0.02005 942.7 0.01892 934.4 0.01785 926.0 0.01685 917.3 0.01591 908.4 0.01501 899.3 0.01417 890.0 0.01338 880.4 0.01262 870.5 0.01191 860.3 0.01123 849.7 0.01059 838.7 0.00998 827.3 0.00940 815.4 0.00884 803.0 0.00831 789.9 0.00780 776.1 0.00731 761.5 0.00684 745.8 0.00639 728.9 0.00594 678.3 0.00486 600.8 0.00370 431.0 0.00232
*Temperatures on ITS-90 scale
52.52 54.71 66.87 79.13 91.55 104.16 116.99 130.05 132.69 133.70 135.35 138.01 140.69 143.38 146.08 148.79 151.52 154.25 157.00 159.77 162.54 165.33 168.13 170.95 173.78 176.63 179.49 182.37 185.27 188.18 191.11 194.05 197.02 200.00 203.00 206.03 209.07 212.13 215.22 218.33 221.47 224.63 227.81 231.02 234.27 237.54 240.84 244.18 247.56 250.97 254.42 257.91 261.45 265.04 268.68 272.39 276.15 279.98 283.90 287.90 292.00 296.22 300.57 305.09 317.45 333.19 358.91
319.59 320.68 326.81 333.06 339.40 345.80 352.21 358.58 359.85 360.33 361.11 362.37 363.62 364.86 366.10 367.34 368.56 369.78 370.99 372.19 373.39 374.57 375.74 376.91 378.06 379.20 380.33 381.44 382.54 383.63 384.70 385.75 386.79 387.81 388.81 389.79 390.75 391.68 392.60 393.48 394.35 395.18 395.98 396.76 397.50 398.20 398.87 399.49 400.07 400.61 401.09 401.52 401.89 402.19 402.42 402.56 402.62 402.58 402.43 402.15 401.72 401.12 400.31 399.24 394.94 385.42 358.91
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid
Vapor
Liquid Vapor
cp /cv Vapor
0.3142 0.3277 0.4000 0.4688 0.5348 0.5984 0.6599 0.7197 0.7314 0.7359 0.7431 0.7548 0.7664 0.7779 0.7894 0.8008 0.8122 0.8236 0.8348 0.8461 0.8573 0.8685 0.8796 0.8907 0.9018 0.9128 0.9238 0.9347 0.9457 0.9566 0.9675 0.9783 0.9892 1.0000 1.0108 1.0216 1.0324 1.0432 1.0539 1.0647 1.0755 1.0863 1.0970 1.1078 1.1186 1.1295 1.1403 1.1512 1.1621 1.1730 1.1840 1.1951 1.2062 1.2174 1.2286 1.2400 1.2515 1.2631 1.2748 1.2868 1.2989 1.3113 1.3240 1.3371 1.3726 1.4172 1.4906
1.9695 1.9579 1.9012 1.8553 1.8180 1.7879 1.7635 1.7438 1.7403 1.7391 1.7370 1.7339 1.7308 1.7279 1.7251 1.7224 1.7198 1.7173 1.7149 1.7126 1.7104 1.7083 1.7062 1.7043 1.7024 1.7005 1.6987 1.6970 1.6953 1.6937 1.6921 1.6906 1.6890 1.6876 1.6861 1.6846 1.6832 1.6818 1.6804 1.6790 1.6775 1.6761 1.6747 1.6732 1.6717 1.6701 1.6685 1.6669 1.6652 1.6634 1.6616 1.6596 1.6575 1.6553 1.6530 1.6505 1.6478 1.6448 1.6416 1.6381 1.6343 1.6300 1.6252 1.6197 1.6018 1.5694 1.4906
1.211 0.630 1.212 0.635 1.220 0.664 1.233 0.694 1.250 0.726 1.270 0.759 1.293 0.794 1.318 0.833 1.323 0.841 1.325 0.844 1.328 0.850 1.334 0.858 1.339 0.867 1.345 0.876 1.351 0.885 1.357 0.894 1.363 0.904 1.369 0.913 1.375 0.923 1.382 0.933 1.388 0.944 1.395 0.955 1.402 0.966 1.409 0.977 1.417 0.988 1.424 1.000 1.432 1.012 1.440 1.025 1.449 1.038 1.457 1.051 1.466 1.065 1.476 1.079 1.485 1.093 1.495 1.109 1.505 1.124 1.516 1.141 1.528 1.158 1.539 1.176 1.552 1.194 1.565 1.214 1.578 1.234 1.593 1.256 1.608 1.278 1.624 1.302 1.641 1.328 1.659 1.355 1.679 1.384 1.699 1.416 1.722 1.449 1.746 1.486 1.772 1.526 1.801 1.570 1.832 1.618 1.867 1.671 1.906 1.732 1.949 1.799 1.998 1.877 2.054 1.966 2.118 2.070 2.194 2.194 2.285 2.343 2.395 2.528 2.534 2.762 2.714 3.069 3.564 4.532 7.720 11.500
1.192 1.191 1.185 1.181 1.178 1.177 1.178 1.182 1.184 1.184 1.185 1.186 1.188 1.189 1.191 1.193 1.195 1.198 1.200 1.203 1.206 1.209 1.212 1.216 1.219 1.223 1.227 1.232 1.237 1.242 1.247 1.253 1.260 1.266 1.273 1.281 1.289 1.298 1.307 1.317 1.328 1.340 1.353 1.366 1.381 1.398 1.416 1.435 1.457 1.480 1.507 1.536 1.569 1.606 1.648 1.696 1.752 1.817 1.894 1.985 2.097 2.236 2.413 2.647 3.763 9.040
a Triple
point
Velocity of Sound, m/s Liquid Vapor 1058 1049 1002 954 907 859 811 764 754 751 745 735 726 716 707 697 688 678 669 659 650 640 630 621 611 602 592 582 573 563 553 544 534 524 515 505 495 485 475 465 455 445 435 425 415 405 394 384 374 363 352 342 331 320 309 298 286 275 263 251 238 226 213 200 164 122 0
137.6 138.2 141.7 144.9 147.8 150.4 152.5 154.2 154.5 154.6 154.7 154.9 155.1 155.3 155.5 155.6 155.7 155.8 155.8 155.9 155.9 155.8 155.8 155.7 155.6 155.4 155.2 155.0 154.8 154.5 154.2 153.9 153.5 153.1 152.6 152.1 151.6 151.0 150.4 149.8 149.1 148.4 147.6 146.8 145.9 145.0 144.0 143.0 141.9 140.8 139.6 138.4 137.1 135.7 134.3 132.8 131.2 129.6 127.9 126.1 124.2 122.2 120.1 117.9 111.8 104.2 0.0 b Normal
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor Liquid Vapor mN/m °C Viscosity, Pa·s
912.1 867.9 680.1 553.4 462.1 393.0 338.8 294.9 287.1 284.2 279.6 272.3 265.3 258.6 252.0 245.7 239.6 233.6 227.8 222.3 216.8 211.6 206.4 201.4 196.6 191.8 187.2 182.7 178.4 174.1 169.9 165.8 161.8 157.9 154.1 150.3 146.6 143.0 139.5 136.0 132.6 129.3 126.0 122.7 119.5 116.4 113.3 110.2 107.2 104.2 101.2 98.3 95.4 92.5 89.7 86.8 84.0 81.1 78.3 75.4 72.5 69.6 66.6 63.6 55.4 45.1 —
5.91 5.97 6.34 6.72 7.08 7.45 7.82 8.19 8.26 8.29 8.33 8.41 8.48 8.56 8.63 8.70 8.78 8.85 8.93 9.18 9.26 9.33 9.41 9.48 9.56 9.64 9.71 9.79 9.87 9.95 10.04 10.12 10.21 10.29 10.38 10.47 10.57 10.66 10.76 10.86 10.96 11.07 11.18 11.29 11.40 11.52 11.64 11.77 11.91 12.04 12.19 12.34 12.50 12.67 12.85 13.03 13.24 13.45 13.69 13.94 14.23 14.54 14.89 15.29 16.64 19.30 —
boiling point
137.0 135.8 129.5 123.5 117.8 112.5 107.4 102.5 101.6 101.2 100.6 99.7 98.8 97.8 96.9 96.0 95.1 94.2 93.3 92.4 91.6 90.7 89.8 89.0 88.1 87.2 86.4 85.5 84.7 83.9 83.0 82.2 81.4 80.5 79.7 78.9 78.1 77.2 76.4 75.6 74.8 74.0 73.2 72.4 71.5 70.7 69.9 69.1 68.3 67.5 66.6 65.8 65.0 64.2 63.3 62.5 61.7 60.8 60.0 59.1 58.2 57.4 56.5 55.6 53.6 53.2
4.90 5.00 5.53 6.10 6.70 7.34 8.02 8.73 8.88 8.94 9.03 9.18 9.33 9.49 9.65 9.81 9.97 10.13 10.30 10.49 10.66 10.83 11.01 11.19 11.37 11.55 11.74 11.94 12.13 12.33 12.53 12.74 12.96 13.17 13.40 13.63 13.87 14.12 14.38 14.64 14.92 15.21 15.51 15.82 16.15 16.50 16.87 17.26 17.67 18.11 18.58 19.09 19.63 20.22 20.87 21.57 22.35 23.20 24.16 25.23 26.45 27.84 29.45 31.37 38.41 55.97
13.72 –111.81 13.75 –110 13.80 –100 13.71 –90 13.48 –80 13.11 –70 12.61 –60 12.00 –50 11.86 –48 11.81 –47.24 11.72 –46 11.57 –44 11.42 –42 11.27 –40 11.11 –38 10.95 –36 10.79 –34 10.62 –32 10.44 –30 10.27 –28 10.09 –26 9.90 –24 9.72 –22 9.53 –20 9.33 –18 9.14 –16 8.94 –14 8.74 –12 8.53 –10 8.32 –8 8.12 –6 7.90 –4 7.69 –2 7.47 0 7.25 2 7.03 4 6.81 6 6.59 8 6.36 10 6.14 12 5.91 14 5.68 16 5.45 18 5.22 20 4.99 22 4.76 24 4.53 26 4.30 28 4.07 30 3.84 32 3.61 34 3.38 36 3.16 38 2.93 40 2.71 42 2.49 44 2.27 46 2.06 48 1.85 50 1.64 52 1.44 54 1.24 56 1.05 58 0.87 60 0.45 65 0.11 70 0.00 72.71 c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 10
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 152a
30.22
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.23
Refrigerant 152a (1,1-Difluoroethane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –118.59a –110 –100 –90 –80 –70 –60 –50 –40 –30 –28 –26 –24.02b –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 90 100 110 113.26c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00006 0.00019 0.00058 0.00153 0.00359 0.00765 0.01500 0.02742 0.04721 0.07718 0.08469 0.09276 0.10133 0.10142 0.11072 0.12068 0.13133 0.14271 0.15484 0.16777 0.18152 0.19614 0.21166 0.22812 0.24555 0.26399 0.28349 0.30407 0.32578 0.34867 0.37277 0.39812 0.42476 0.45275 0.48211 0.51291 0.54517 0.57894 0.61428 0.65122 0.68982 0.73012 0.77216 0.81600 0.86169 0.90927 0.95879 1.01030 1.06390 1.11960 1.17740 1.23740 1.29970 1.36430 1.43130 1.50070 1.57260 1.64710 1.72420 1.80390 1.88640 1.97170 2.05990 2.15100 2.24520 2.34240 2.87800 3.50500 4.24320 4.51680
1192.9 1177.1 1158.7 1140.1 1121.3 1102.4 1083.2 1063.7 1043.8 1023.5 1019.4 1015.3 1011.2 1011.1 1006.9 1002.7 998.5 994.2 989.9 985.6 981.3 976.9 972.5 968.1 963.6 959.1 954.6 950.0 945.4 940.8 936.1 931.3 926.6 921.8 916.9 912.0 907.0 902.0 896.9 891.8 886.6 881.4 876.0 870.7 865.2 859.7 854.1 848.4 842.6 836.7 830.8 824.7 818.6 812.3 805.9 799.4 792.7 785.9 779.0 771.9 764.6 757.2 749.5 741.6 733.5 725.2 678.5 618.5 517.4 368.0
303.290 107.740 37.6170 15.0520 6.74380 3.31970 1.76820 1.00640 0.60583 0.38242 0.35056 0.32186 0.29622 0.29595 0.27253 0.25131 0.23206 0.21457 0.19865 0.18414 0.17090 0.15879 0.14770 0.13754 0.12821 0.11963 0.11174 0.10447 0.09776 0.09156 0.08583 0.08052 0.07560 0.07104 0.06680 0.06286 0.05919 0.05577 0.05258 0.04960 0.04682 0.04422 0.04179 0.03951 0.03737 0.03536 0.03348 0.03170 0.03004 0.02846 0.02699 0.02559 0.02427 0.02303 0.02185 0.02074 0.01968 0.01868 0.01774 0.01684 0.01598 0.01517 0.01440 0.01366 0.01296 0.01228 0.00933 0.00686 0.00446 0.00272
*Temperatures on ITS-90 scale
13.79 26.62 41.75 56.96 72.23 87.57 103.02 118.62 134.40 150.39 153.62 156.86 160.07 160.11 163.37 166.64 169.92 173.21 176.52 179.83 183.16 186.50 189.86 193.22 196.61 200.00 203.41 206.83 210.27 213.72 217.19 220.67 224.17 227.69 231.22 234.77 238.34 241.93 245.53 249.16 252.80 256.47 260.16 263.86 267.60 271.35 275.13 278.93 282.76 286.62 290.50 294.41 298.35 302.33 306.34 310.38 314.45 318.57 322.72 326.92 331.16 335.45 339.79 344.18 348.63 353.15 376.87 403.59 439.22 477.55
419.32 425.38 432.59 439.97 447.48 455.08 462.74 470.40 478.02 485.55 487.04 488.52 489.98 490.00 491.47 492.94 494.40 495.85 497.29 498.72 500.15 501.56 502.96 504.36 505.74 507.11 508.47 509.82 511.16 512.48 513.78 515.08 516.36 517.62 518.86 520.09 521.30 522.50 523.67 524.83 525.96 527.07 528.16 529.23 530.27 531.28 532.27 533.23 534.16 535.06 535.93 536.77 537.56 538.32 539.04 539.72 540.35 540.94 541.47 541.95 542.37 542.73 543.02 543.24 543.38 543.43 542.06 536.28 517.31 477.55
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
0.1119 0.1927 0.2827 0.3681 0.4492 0.5266 0.6009 0.6723 0.7414 0.8085 0.8216 0.8348 0.8477 0.8478 0.8608 0.8737 0.8866 0.8994 0.9122 0.9249 0.9375 0.9501 0.9627 0.9752 0.9876 1.0000 1.0124 1.0247 1.0370 1.0492 1.0614 1.0736 1.0857 1.0978 1.1098 1.1219 1.1339 1.1459 1.1578 1.1698 1.1817 1.1936 1.2055 1.2174 1.2292 1.2411 1.2529 1.2648 1.2766 1.2884 1.3003 1.3121 1.3240 1.3358 1.3477 1.3596 1.3716 1.3835 1.3955 1.4076 1.4196 1.4318 1.4440 1.4562 1.4686 1.4810 1.5451 1.6151 1.7058 1.8037
1.477 1.505 1.518 1.524 1.530 1.539 1.551 1.567 1.587 1.610 1.615 1.620 1.625 1.625 1.630 1.635 1.641 1.647 1.653 1.658 1.665 1.671 1.677 1.684 1.690 1.697 1.704 1.711 1.719 1.726 1.734 1.742 1.750 1.759 1.768 1.776 1.786 1.795 1.805 1.815 1.826 1.837 1.848 1.860 1.872 1.885 1.898 1.912 1.926 1.941 1.957 1.974 1.992 2.010 2.030 2.051 2.073 2.097 2.122 2.150 2.179 2.211 2.245 2.283 2.324 2.370 2.703 3.495 9.260
1.220 1.214 1.207 1.201 1.196 1.192 1.190 1.189 1.190 1.193 1.194 1.195 1.196 1.196 1.197 1.199 1.200 1.202 1.203 1.205 1.207 1.209 1.211 1.213 1.215 1.218 1.221 1.224 1.227 1.230 1.233 1.237 1.240 1.244 1.249 1.253 1.258 1.263 1.268 1.274 1.280 1.286 1.293 1.300 1.307 1.315 1.324 1.333 1.342 1.353 1.364 1.375 1.388 1.402 1.416 1.432 1.450 1.468 1.488 1.511 1.535 1.562 1.591 1.624 1.661 1.702 2.016 2.805 8.530
2.7357 2.6368 2.5399 2.4593 2.3920 2.3357 2.2885 2.2487 2.2152 2.1868 2.1817 2.1767 2.1719 2.1719 2.1672 2.1627 2.1583 2.1541 2.1500 2.1460 2.1421 2.1383 2.1347 2.1311 2.1277 2.1243 2.1211 2.1179 2.1148 2.1118 2.1089 2.1060 2.1032 2.1005 2.0978 2.0952 2.0926 2.0901 2.0876 2.0852 2.0828 2.0804 2.0780 2.0757 2.0734 2.0711 2.0689 2.0666 2.0643 2.0620 2.0598 2.0575 2.0552 2.0528 2.0504 2.0480 2.0456 2.0431 2.0405 2.0379 2.0351 2.0323 2.0294 2.0264 2.0232 2.0198 2.0000 1.9707 1.9096 1.8037
a Triple
point
0.699 0.717 0.740 0.763 0.789 0.816 0.845 0.877 0.913 0.952 0.960 0.968 0.977 0.977 0.985 0.994 1.003 1.013 1.022 1.032 1.041 1.051 1.062 1.072 1.083 1.094 1.105 1.116 1.128 1.139 1.152 1.164 1.177 1.190 1.203 1.217 1.231 1.246 1.261 1.277 1.293 1.309 1.326 1.344 1.362 1.381 1.401 1.421 1.443 1.465 1.489 1.513 1.539 1.566 1.595 1.626 1.658 1.693 1.730 1.769 1.812 1.859 1.909 1.964 2.025 2.092 2.586 3.776 12.220
Velocity of Sound, m/s Liquid Vapor 1401 1337 1275 1219 1166 1115 1066 1016 967 919 909 899 889 889 880 870 860 850 841 831 821 811 801 792 782 772 762 752 742 733 723 713 703 693 683 673 663 653 643 633 623 613 602 592 582 572 561 551 541 530 520 509 499 488 478 467 456 445 435 424 413 402 390 379 368 356 297 233 158 0
154.0 157.8 162.0 166.0 169.8 173.4 176.6 179.5 182.1 184.2 184.5 184.9 185.2 185.2 185.5 185.8 186.1 186.3 186.5 186.7 186.9 187.0 187.1 187.2 187.3 187.4 187.4 187.4 187.4 187.3 187.2 187.1 187.0 186.8 186.6 186.4 186.1 185.8 185.5 185.1 184.7 184.3 183.9 183.4 182.9 182.3 181.7 181.1 180.4 179.7 178.9 178.2 177.3 176.4 175.5 174.6 173.6 172.5 171.4 170.2 169.0 167.8 166.5 165.1 163.7 162.2 153.7 143.1 129.1 0.0 b Normal
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor Liquid Vapor mN/m °C Viscosity, Pa·s
2025.0 1528.0 1153.0 904.0 730.2 603.5 507.4 432.5 372.5 323.6 314.8 306.4 298.4 298.3 290.5 282.9 275.6 268.6 261.7 255.1 248.7 242.6 236.6 230.8 225.1 219.7 214.4 209.2 204.2 199.4 194.7 190.1 185.6 181.3 177.1 173.0 169.0 165.1 161.3 157.6 154.0 150.4 147.0 143.6 140.3 137.1 134.0 130.9 127.9 125.0 122.1 119.2 116.5 113.7 111.1 108.4 105.8 103.3 100.8 98.3 95.9 93.5 91.1 88.7 86.4 84.1 72.6 60.7 45.6 —
5.20 5.48 5.81 6.14 6.47 6.80 7.14 7.47 7.80 8.14 8.21 8.27 8.34 8.34 8.41 8.48 8.54 8.61 8.68 8.75 8.82 8.88 8.95 9.02 9.09 9.16 9.23 9.30 9.37 9.44 9.51 9.58 9.65 9.73 9.80 9.87 9.95 10.02 10.10 10.18 10.26 10.34 10.42 10.50 10.58 10.66 10.75 10.84 11.21 11.32 11.42 11.53 11.64 11.76 11.88 12.00 12.12 12.26 12.39 12.53 12.68 12.83 12.99 13.16 13.34 13.52 14.64 16.35 20.19 —
boiling point
176.3 170.2 163.4 156.9 150.7 144.7 139.0 133.5 128.2 123.1 122.1 121.2 120.2 120.2 119.2 118.2 117.3 116.3 115.4 114.5 113.5 112.6 111.7 110.8 109.8 108.9 108.0 107.1 106.2 105.4 104.5 103.6 102.7 101.8 101.0 100.1 99.3 98.4 97.5 96.7 95.9 95.0 94.2 93.3 92.5 91.7 90.8 90.0 89.2 88.4 87.5 86.7 85.9 85.1 84.2 83.4 82.6 81.8 80.9 80.1 79.3 78.5 77.6 76.8 76.0 75.2 71.0 67.0 67.5
0.10 0.94 1.91 2.89 3.86 4.84 5.82 6.81 7.80 8.80 9.01 9.21 9.41 9.41 9.62 9.82 10.03 10.23 10.44 10.65 10.86 11.07 11.28 11.50 11.71 11.93 12.14 12.36 12.58 12.80 13.03 13.25 13.48 13.71 13.95 14.18 14.42 14.66 14.91 15.16 15.41 15.67 15.93 16.20 16.47 16.74 17.03 17.32 17.60 17.90 18.22 18.54 18.87 19.21 19.56 19.92 20.30 20.70 21.11 21.54 22.00 22.48 22.98 23.51 24.08 24.69 28.54 35.03 53.48
31.65 30.23 28.58 26.96 25.34 23.75 22.18 20.63 19.09 17.58 17.29 16.99 16.69 16.69 16.39 16.10 15.80 15.51 15.22 14.93 14.64 14.35 14.06 13.77 13.48 13.20 12.91 12.63 12.35 12.07 11.79 11.51 11.23 10.96 10.68 10.41 10.14 9.87 9.60 9.33 9.06 8.80 8.54 8.27 8.01 7.75 7.50 7.24 6.99 6.73 6.48 6.23 5.98 5.74 5.50 5.25 5.01 4.78 4.54 4.31 4.08 3.85 3.62 3.40 3.18 2.96 1.91 0.96 0.17 0.00
–118.59 –110 –100 –90 –80 –70 –60 –50 –40 –30 –28 –26 –24.02 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 90 100 110 113.26
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 11
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 245fa
30.24
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.25
Refrigerant 245fa (1,1,1,3,3-pentafluoroethane) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp., °C –60 –50 –40 –30 –20 –10 0 2 4 6 8 10 12 14 15.05b 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 82 84 86 88 90 92 94 96 98 100 105 110 115 120 125 130 135 140 145 150 153.86c bNormal
0.00131 0.00289 0.00584 0.01103 0.01961 0.03307 0.05327 0.05831 0.06372 0.06953 0.07576 0.08242 0.08954 0.09714 0.10132 0.10525 0.11388 0.12306 0.13282 0.14317 0.15415 0.16578 0.17808 0.19108 0.20481 0.21930 0.23457 0.25065 0.26757 0.28535 0.30403 0.32364 0.34421 0.36576 0.38833 0.41195 0.43665 0.46246 0.48941 0.51754 0.54689 0.57747 0.60933 0.64250 0.67702 0.71292 0.75024 0.78901 0.82927 0.87105 0.91440 0.95935 1.0059 1.0542 1.1042 1.1559 1.2095 1.2649 1.4117 1.5711 1.7437 1.9304 2.1320 2.3495 2.5841 2.8370 3.1099 3.4050 3.6510
1548.2 1525.0 1501.5 1477.7 1453.5 1429.0 1403.9 1398.8 1393.7 1388.6 1383.4 1378.3 1373.0 1367.8 1365.1 1362.5 1357.2 1351.9 1346.6 1341.2 1335.7 1330.3 1324.8 1319.3 1313.7 1308.1 1302.4 1296.7 1291.0 1285.2 1279.3 1273.4 1267.5 1261.5 1255.4 1249.3 1243.2 1236.9 1230.6 1224.3 1217.8 1211.3 1204.7 1198.0 1191.3 1184.4 1177.5 1170.5 1163.4 1156.1 1148.8 1141.3 1133.7 1126.0 1118.2 1110.2 1102.0 1093.7 1072.1 1049.3 1024.9 998.6 970.0 938.4 902.7 861.1 810.0 738.4 519.4
boiling point
10.041 4.7815 2.4619 1.3552 0.79014 0.48405 0.30947 0.28429 0.26153 0.24093 0.22226 0.20530 0.18987 0.17582 0.16896 0.16301 0.15131 0.14061 0.13080 0.12182 0.11356 0.10598 0.09899 0.09256 0.08662 0.08113 0.07606 0.07137 0.06702 0.06298 0.05923 0.05575 0.05251 0.04949 0.04667 0.04405 0.04159 0.03930 0.03715 0.03514 0.03326 0.03149 0.02984 0.02828 0.02681 0.02543 0.02414 0.02291 0.02176 0.02067 0.01964 0.01867 0.01774 0.01687 0.01605 0.01526 0.01452 0.01381 0.01220 0.01077 0.00950 0.00836 0.00734 0.00641 0.00555 0.00476 0.00399 0.00321 0.00193
127.42 139.14 150.99 162.99 175.14 187.48 200.00 202.53 205.07 207.61 210.17 212.73 215.30 217.88 219.24 220.47 223.07 225.68 228.29 230.92 233.56 236.20 238.86 241.52 244.20 246.89 249.58 252.29 255.01 257.74 260.48 263.23 266.00 268.77 271.56 274.36 277.17 279.99 282.83 285.68 288.54 291.42 294.31 297.22 300.13 303.07 306.02 308.98 311.96 314.96 317.97 321.01 324.06 327.12 330.21 333.32 336.45 339.60 347.57 355.71 364.02 372.54 381.31 390.39 399.87 409.90 420.80 433.55 459.00
361.18 368.14 375.25 382.49 389.83 397.25 404.72 406.22 407.72 409.22 410.72 412.22 413.72 415.22 416.01 416.72 418.22 419.72 421.22 422.71 424.21 425.70 427.19 428.68 430.17 431.65 433.13 434.61 436.08 437.55 439.01 440.47 441.93 443.38 444.82 446.26 447.69 449.11 450.53 451.94 453.34 454.73 456.11 457.48 458.84 460.19 461.53 462.86 464.17 465.46 466.74 468.01 469.26 470.48 471.69 472.88 474.04 475.19 477.92 480.46 482.77 484.79 486.45 487.64 488.22 487.91 486.22 481.69 459.00
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
0.7006 0.7543 0.8063 0.8566 0.9056 0.9533 1.0000 1.0092 1.0184 1.0275 1.0366 1.0457 1.0547 1.0637 1.0684 1.0727 1.0816 1.0905 1.0994 1.1082 1.1170 1.1258 1.1346 1.1433 1.1520 1.1607 1.1694 1.1780 1.1866 1.1952 1.2038 1.2123 1.2209 1.2294 1.2379 1.2463 1.2548 1.2632 1.2717 1.2801 1.2885 1.2968 1.3052 1.3136 1.3219 1.3303 1.3386 1.3469 1.3552 1.3636 1.3719 1.3802 1.3885 1.3968 1.4051 1.4134 1.4217 1.4301 1.4509 1.4719 1.4930 1.5144 1.5360 1.5581 1.5809 1.6046 1.6300 1.6594 1.7183
1.166 1.178 1.192 1.207 1.224 1.242 1.261 1.265 1.270 1.274 1.278 1.282 1.287 1.291 1.293 1.296 1.300 1.305 1.309 1.314 1.319 1.324 1.329 1.334 1.339 1.344 1.349 1.355 1.360 1.366 1.371 1.377 1.383 1.389 1.395 1.401 1.407 1.414 1.420 1.427 1.434 1.441 1.448 1.456 1.464 1.472 1.480 1.488 1.497 1.506 1.516 1.525 1.536 1.546 1.557 1.569 1.582 1.595 1.630 1.673 1.723 1.786 1.867 1.977 2.137 2.398 2.927 4.734
1.099 1.097 1.095 1.094 1.093 1.094 1.095 1.096 1.096 1.097 1.097 1.098 1.098 1.099 1.100 1.100 1.101 1.102 1.102 1.103 1.104 1.106 1.107 1.108 1.109 1.111 1.112 1.114 1.115 1.117 1.119 1.121 1.123 1.125 1.128 1.130 1.133 1.135 1.138 1.141 1.145 1.148 1.152 1.156 1.160 1.165 1.169 1.175 1.180 1.186 1.192 1.199 1.206 1.214 1.222 1.231 1.241 1.252 1.283 1.323 1.374 1.442 1.535 1.670 1.877 2.232 2.961 5.361
1.7973 1.7805 1.7681 1.7594 1.7537 1.7505 1.7495 1.7495 1.7496 1.7497 1.7500 1.7502 1.7506 1.7510 1.7512 1.7514 1.7519 1.7524 1.7530 1.7537 1.7544 1.7551 1.7559 1.7567 1.7575 1.7584 1.7593 1.7602 1.7612 1.7622 1.7632 1.7642 1.7653 1.7664 1.7675 1.7686 1.7697 1.7709 1.7720 1.7732 1.7744 1.7755 1.7767 1.7779 1.7791 1.7803 1.7815 1.7826 1.7838 1.7850 1.7861 1.7872 1.7883 1.7894 1.7905 1.7915 1.7925 1.7934 1.7956 1.7975 1.7990 1.7999 1.8001 1.7993 1.7973 1.7934 1.7865 1.7732 1.7183
0.695 0.716 0.738 0.760 0.783 0.807 0.833 0.838 0.843 0.849 0.854 0.860 0.865 0.871 0.874 0.877 0.883 0.888 0.894 0.900 0.907 0.913 0.919 0.926 0.932 0.939 0.946 0.952 0.959 0.967 0.974 0.981 0.989 0.996 1.004 1.012 1.021 1.029 1.038 1.046 1.056 1.065 1.075 1.084 1.095 1.105 1.116 1.128 1.140 1.152 1.165 1.178 1.192 1.207 1.223 1.239 1.257 1.275 1.327 1.389 1.466 1.563 1.693 1.875 2.148 2.606 3.536 6.570
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
1043 996 950 904 859 816 773 764 756 748 739 731 723 714 710 706 698 690 682 673 665 657 649 641 633 625 617 609 601 593 585 577 569 560 552 544 536 528 520 512 504 496 487 479 471 463 454 446 438 429 421 412 404 395 387 378 369 360 338 316 292 269 245 220 194 167 139 108 0
120.3 122.8 125.2 127.3 129.3 131.0 132.5 132.8 133.0 133.2 133.5 133.7 133.9 134.0 134.1 134.2 134.4 134.5 134.6 134.7 134.8 134.9 134.9 135.0 135.0 135.0 135.0 134.9 134.9 134.8 134.7 134.6 134.4 134.3 134.1 133.9 133.7 133.4 133.1 132.8 132.5 132.2 131.8 131.4 131.0 130.5 130.0 129.5 129.0 128.4 127.8 127.1 126.4 125.7 125.0 124.2 123.4 122.5 120.2 117.5 114.6 111.4 107.9 103.9 99.6 94.9 89.6 83.4 0.0
1879.0 1495.0 1205.0 984.2 814.4 682.5 578.8 560.8 543.5 527.1 511.3 496.2 481.7 467.9 460.8 454.5 441.8 429.5 417.7 406.3 395.4 384.9 374.8 365.0 355.6 346.4 337.6 329.1 320.9 312.9 305.2 297.7 290.4 283.3 276.5 269.8 263.3 256.9 250.8 244.7 238.9 233.1 227.5 222.0 216.7 211.4 206.3 201.2 196.2 191.4 186.6 181.9 177.3 172.7 168.2 163.8 159.4 155.1 144.5 134.3 124.2 114.3 104.6 94.8 85.0 75.0 64.6 52.6 —
7.17 7.57 7.96 8.35 8.74 9.11 9.49 9.56 9.63 9.71 9.78 9.85 9.93 10.00 10.04 10.07 10.15 10.22 10.29 10.36 10.44 10.51 10.58 10.65 10.72 10.80 10.87 10.94 11.01 11.09 11.16 11.23 11.31 11.38 11.45 11.53 11.60 11.68 11.75 11.83 11.90 11.98 12.06 12.14 12.21 12.29 12.37 12.46 12.54 12.62 12.71 12.79 12.88 12.97 13.06 13.15 13.25 13.34 13.60 13.87 14.17 14.50 14.88 15.33 15.88 16.61 17.67 19.55 —
Thermal Cond., Surface mW/(m·K) Tension, Temp., Liquid Vapor mN/m °C 114.9 112.0 109.0 106.0 102.8 99.6 96.3 95.7 95.0 94.4 93.7 93.1 92.4 91.8 91.4 91.1 90.5 89.8 89.1 88.5 87.8 87.2 86.5 85.9 85.2 84.6 83.9 83.3 82.6 82.0 81.4 80.7 80.1 79.4 78.8 78.1 77.5 76.9 76.2 75.6 75.0 74.3 73.7 73.1 72.5 71.8 71.2 70.6 70.0 69.3 68.7 68.1 67.5 66.9 66.3 65.7 65.1 64.4 62.9 61.4 59.9 58.4 57.0 55.5 54.1 52.8 52.0 53.2
4.97 5.87 6.78 7.68 8.58 9.48 10.37 10.55 10.73 10.91 11.09 11.27 11.45 11.62 11.72 11.80 11.98 12.16 12.34 12.52 12.69 12.87 13.05 13.23 13.41 13.59 13.77 13.95 14.13 14.31 14.49 14.67 14.85 15.03 15.21 15.40 15.58 15.77 15.96 16.14 16.33 16.52 16.72 16.91 17.11 17.31 17.51 17.72 17.93 18.14 18.36 18.58 18.80 19.03 19.27 19.52 19.77 20.03 20.73 21.51 22.41 23.47 24.75 26.38 28.56 31.65 36.54 46.38
25.13 23.73 22.33 20.95 19.59 18.23 16.89 16.63 16.36 16.10 15.83 15.57 15.31 15.05 14.91 14.79 14.53 14.27 14.01 13.75 13.49 13.23 12.98 12.72 12.47 12.22 11.96 11.71 11.46 11.21 10.96 10.71 10.46 10.22 9.97 9.73 9.48 9.24 9.00 8.76 8.52 8.28 8.04 7.81 7.57 7.34 7.11 6.88 6.65 6.42 6.19 5.96 5.74 5.52 5.30 5.08 4.86 4.64 4.11 3.59 3.08 2.59 2.11 1.66 1.22 0.82 0.46 0.16 0.00
–60 –50 –40 –30 –20 –10 0 2 4 6 8 10 12 14 15.05 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 82 84 86 88 90 92 94 96 98 100 105 110 115 120 125 130 135 140 145 150 153.86
cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 12
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant R-1233zd(E)
30.26
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.27
Refrigerant 1233zd(E) (trans-1-chloro-3,3,3-trifluoroprop-1-ene) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp., °C
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor
–50 0.00289 –45 0.00410 –40 0.00570 –35 0.00781 –30 0.01052 –25 0.01398 –20 0.01833 –18 0.02035 –16 0.02256 –14 0.02495 –12 0.02755 –10 0.03036 –8 0.03340 –6 0.03668 –4 0.04021 –2 0.04402 0 0.04811 2 0.05250 4 0.05720 6 0.06224 8 0.06762 10 0.07337 12 0.07949 14 0.08601 16 0.09295 18 0.10032 18.26b 0.10132 20 0.10815 22 0.11644 24 0.12522 26 0.13452 28 0.14434 30 0.15471 32 0.16566 34 0.17719 36 0.18934 38 0.20212 40 0.21555 42 0.22966 44 0.24447 46 0.26000 48 0.27628 50 0.29332 52 0.31116 54 0.32981 56 0.34929 58 0.36964 60 0.39088 62 0.41303 64 0.43612 66 0.46017 68 0.48521 70 0.51126 72 0.53835 74 0.56651 76 0.59576 78 0.62613 80 0.65765 90 0.83345 100 1.0423 110 1.2880 120 1.5750 130 1.9080 140 2.2928 150 2.7364 160 3.2485 166.45c 3.6237 bNormal
1429.8 1419.3 1408.7 1398.1 1387.4 1376.6 1365.7 1361.3 1357.0 1352.6 1348.1 1343.7 1339.3 1334.8 1330.3 1325.8 1321.3 1316.7 1312.2 1307.6 1303.0 1298.3 1293.7 1289.0 1284.3 1279.6 1278.9 1274.8 1270.0 1265.2 1260.4 1255.5 1250.6 1245.7 1240.7 1235.7 1230.7 1225.6 1220.5 1215.4 1210.2 1205.0 1199.7 1194.4 1189.1 1183.7 1178.3 1172.8 1167.2 1161.7 1156.0 1150.3 1144.6 1138.8 1132.9 1127.0 1121.0 1114.9 1083.4 1049.7 1013.1 972.6 926.9 873.1 805.4 704.0 480.2
boiling point
4.9048 3.5348 2.5918 1.9310 1.4601 1.1193 0.86907 0.78804 0.71589 0.65150 0.59392 0.54235 0.49606 0.45444 0.41695 0.38313 0.35257 0.32490 0.29981 0.27702 0.25630 0.23743 0.22022 0.20449 0.19011 0.17694 0.17529 0.16486 0.15376 0.14356 0.13417 0.12552 0.11753 0.11015 0.10333 0.09702 0.09116 0.08573 0.08069 0.07600 0.07163 0.06757 0.06378 0.06024 0.05693 0.05385 0.05096 0.04825 0.04571 0.04333 0.04110 0.03900 0.03702 0.03517 0.03342 0.03177 0.03021 0.02875 0.02253 0.01778 0.01408 0.01116 0.00881 0.00687 0.00522 0.00368 0.00208
142.87 148.43 154.02 159.65 165.31 171.00 176.73 179.03 181.34 183.65 185.97 188.29 190.62 192.96 195.30 197.65 200.00 202.36 204.72 207.10 209.47 211.86 214.25 216.64 219.04 221.45 221.77 223.87 226.29 228.72 231.15 233.59 236.04 238.50 240.96 243.43 245.91 248.39 250.88 253.38 255.89 258.40 260.93 263.46 266.00 268.55 271.10 273.67 276.24 278.83 281.42 284.03 286.64 289.26 291.90 294.54 297.20 299.86 313.38 327.24 341.51 356.29 371.74 388.12 406.03 427.43 458.27
369.48 372.76 376.08 379.43 382.82 386.24 389.68 391.06 392.45 393.83 395.22 396.62 398.01 399.41 400.80 402.20 403.60 405.00 406.40 407.80 409.21 410.61 412.01 413.41 414.81 416.20 416.39 417.60 419.00 420.39 421.78 423.17 424.56 425.94 427.32 428.70 430.08 431.45 432.81 434.18 435.54 436.89 438.24 439.59 440.93 442.26 443.59 444.91 446.23 447.54 448.84 450.14 451.43 452.70 453.98 455.24 456.49 457.73 463.78 469.49 474.77 479.44 483.27 485.81 486.16 481.49 458.27
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
0.7694 0.7940 0.8182 0.8421 0.8656 0.8888 0.9116 0.9207 0.9297 0.9386 0.9475 0.9564 0.9652 0.9740 0.9827 0.9914 1.0000 1.0086 1.0171 1.0257 1.0341 1.0426 1.0510 1.0593 1.0676 1.0759 1.0770 1.0841 1.0924 1.1005 1.1087 1.1168 1.1249 1.1329 1.1409 1.1489 1.1569 1.1648 1.1727 1.1805 1.1884 1.1962 1.2040 1.2117 1.2195 1.2272 1.2349 1.2426 1.2502 1.2578 1.2655 1.2730 1.2806 1.2882 1.2957 1.3032 1.3107 1.3182 1.3555 1.3926 1.4297 1.4671 1.5050 1.5441 1.5856 1.6340 1.7032
1.109 1.115 1.122 1.128 1.135 1.142 1.149 1.152 1.154 1.157 1.160 1.163 1.166 1.169 1.171 1.174 1.177 1.180 1.183 1.186 1.189 1.192 1.195 1.198 1.201 1.205 1.205 1.208 1.211 1.214 1.218 1.221 1.224 1.228 1.231 1.235 1.238 1.242 1.246 1.250 1.254 1.258 1.262 1.266 1.270 1.274 1.279 1.283 1.288 1.293 1.298 1.303 1.308 1.313 1.319 1.324 1.330 1.336 1.370 1.412 1.465 1.538 1.643 1.819 2.193 3.751
1.108 1.107 1.105 1.104 1.104 1.103 1.103 1.102 1.102 1.102 1.102 1.103 1.103 1.103 1.103 1.103 1.104 1.104 1.104 1.105 1.105 1.106 1.106 1.107 1.108 1.108 1.108 1.109 1.110 1.111 1.112 1.113 1.114 1.115 1.116 1.117 1.119 1.120 1.122 1.123 1.125 1.127 1.128 1.130 1.132 1.134 1.137 1.139 1.142 1.144 1.147 1.150 1.153 1.156 1.159 1.163 1.167 1.171 1.195 1.227 1.273 1.342 1.455 1.661 2.141 4.191
1.7849 1.7773 1.7707 1.7650 1.7602 1.7561 1.7528 1.7517 1.7506 1.7497 1.7488 1.7480 1.7473 1.7467 1.7462 1.7458 1.7454 1.7451 1.7448 1.7447 1.7445 1.7445 1.7445 1.7445 1.7447 1.7448 1.7448 1.7450 1.7453 1.7456 1.7459 1.7463 1.7467 1.7472 1.7477 1.7482 1.7488 1.7493 1.7500 1.7506 1.7513 1.7520 1.7527 1.7534 1.7542 1.7550 1.7558 1.7566 1.7574 1.7582 1.7591 1.7600 1.7608 1.7617 1.7626 1.7635 1.7644 1.7653 1.7697 1.7739 1.7775 1.7803 1.7816 1.7805 1.7750 1.7588 1.7032
0.660 0.672 0.683 0.694 0.704 0.715 0.726 0.730 0.734 0.739 0.743 0.747 0.751 0.756 0.760 0.764 0.769 0.773 0.778 0.782 0.787 0.791 0.796 0.801 0.805 0.810 0.811 0.815 0.820 0.825 0.830 0.835 0.840 0.845 0.850 0.856 0.861 0.866 0.872 0.878 0.884 0.889 0.895 0.902 0.908 0.914 0.921 0.927 0.934 0.941 0.948 0.955 0.963 0.970 0.978 0.986 0.995 1.003 1.050 1.107 1.179 1.277 1.423 1.675 2.232 4.545
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
992 971 949 928 907 886 865 857 849 841 833 825 817 809 800 792 784 776 769 761 753 745 737 729 721 713 712 705 698 690 682 674 666 659 651 643 635 627 620 612 604 597 589 581 573 566 558 550 542 535 527 519 512 504 496 488 481 473 434 394 354 312 270 225 177 123 0
125.2 126.3 127.5 128.6 129.7 130.7 131.7 132.1 132.4 132.8 133.1 133.5 133.8 134.1 134.4 134.7 135.0 135.3 135.6 135.8 136.1 136.3 136.5 136.7 136.9 137.1 137.1 137.2 137.4 137.5 137.6 137.7 137.8 137.9 138.0 138.0 138.0 138.0 138.0 138.0 138.0 137.9 137.8 137.7 137.6 137.5 137.3 137.2 137.0 136.7 136.5 136.3 136.0 135.7 135.3 135.0 134.6 134.2 131.8 128.7 124.8 120.0 114.2 107.1 98.7 88.04 0.00
2041.0 1735.0 1504.0 1322.0 1175.0 1054.0 953.2 917.2 883.3 851.5 821.4 793.1 766.2 740.8 716.7 693.8 672.1 651.3 631.6 612.7 594.7 577.4 560.9 545.1 529.9 515.3 513.5 501.3 487.9 474.9 462.4 450.4 438.8 427.6 416.8 406.3 396.2 386.4 376.9 367.7 358.8 350.2 341.8 333.7 325.8 318.1 310.6 303.3 296.2 289.3 282.6 276.0 269.6 263.4 257.3 251.3 245.5 239.8 213.1 188.9 166.8 146.4 127.1 108.5 89.7 68.4 —
8.31 8.50 8.69 8.87 9.06 9.25 9.43 9.51 9.58 9.66 9.73 9.80 9.88 9.95 10.02 10.10 10.17 10.24 10.32 10.39 10.46 10.54 10.61 10.68 10.75 10.83 10.84 10.90 10.97 11.05 11.12 11.19 11.26 11.34 11.41 11.48 11.55 11.63 11.70 11.78 11.85 11.92 12.00 12.07 12.15 12.22 12.30 12.38 12.45 12.53 12.61 12.69 12.77 12.85 12.93 13.02 13.10 13.19 13.64 14.15 14.74 15.45 16.37 17.65 19.67 23.83 —
Thermal Cond., Surface mW/(m·K) Tension, Temp., Liquid Vapor mN/m °C 106.3 104.6 102.9 101.3 99.7 98.1 96.5 95.8 95.2 94.6 94.0 93.3 92.7 92.1 91.4 90.8 90.2 89.6 89.0 88.3 87.7 87.1 86.5 85.9 85.3 84.7 84.6 84.1 83.5 82.9 82.3 81.7 81.1 80.5 79.9 79.3 78.7 78.1 77.5 76.9 76.3 75.8 75.2 74.6 74.0 73.5 72.9 72.3 71.8 71.2 70.6 70.1 69.5 69.0 68.4 67.9 67.3 66.8 64.1 61.6 59.0 56.6 54.3 52.0 49.8 48.6
6.08 6.39 6.70 7.01 7.33 7.65 7.97 8.10 8.23 8.36 8.49 8.63 8.76 8.89 9.03 9.16 9.30 9.44 9.58 9.72 9.86 10.00 10.14 10.28 10.43 10.57 10.59 10.72 10.87 11.02 11.17 11.32 11.48 11.63 11.79 11.95 12.11 12.27 12.43 12.60 12.76 12.93 13.10 13.27 13.45 13.62 13.80 13.98 14.16 14.34 14.53 14.72 14.91 15.10 15.29 15.49 15.69 15.89 16.94 18.07 19.30 20.69 22.34 24.52 27.90 35.43
25.07 24.33 23.60 22.87 22.15 21.43 20.72 20.44 20.15 19.87 19.59 19.31 19.03 18.75 18.48 18.20 17.92 17.65 17.38 17.10 16.83 16.56 16.29 16.02 15.75 15.49 15.45 15.22 14.96 14.69 14.43 14.17 13.91 13.65 13.39 13.13 12.87 12.62 12.36 12.11 11.86 11.61 11.36 11.11 10.86 10.62 10.37 10.13 9.89 9.64 9.40 9.17 8.93 8.69 8.46 8.23 7.99 7.76 6.64 5.55 4.51 3.51 2.58 1.71 0.93 0.28 0.00
–50 –45 –40 –35 –30 –25 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 18.26 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 90 100 110 120 130 140 150 160 166.45
cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 13 Pressure-Enthalpy Diagram for Refrigerant 1234yf
30.28
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.29
Refrigerant 1234yf (2,3,3,3-tetrafluoroprop-1-ene) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp., °C –50 –45 –40 –35 –30 –29.49b –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 94.70c bNormal
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.03742 0.04862 0.06237 0.07904 0.09906 0.10132 0.10810 0.11778 0.12812 0.13916 0.15092 0.16344 0.17676 0.19090 0.20590 0.22178 0.23860 0.25637 0.27514 0.29495 0.31582 0.33780 0.36092 0.38523 0.41075 0.43753 0.46561 0.49503 0.52583 0.55804 0.59172 0.62690 0.66363 0.70194 0.74189 0.78351 0.82686 0.87197 0.91890 0.96769 1.0184 1.0711 1.1257 1.1825 1.2413 1.3023 1.3656 1.4311 1.4989 1.5692 1.6419 1.7171 1.7949 1.8754 1.9586 2.0445 2.1334 2.2252 2.3201 2.4181 2.5194 2.7879 3.0803 3.3822
1318.4 1305.2 1291.9 1278.3 1264.5 1263.1 1259.0 1253.4 1247.7 1242.0 1236.3 1230.5 1224.7 1218.8 1212.9 1207.0 1200.9 1194.9 1188.7 1182.5 1176.3 1170.0 1163.6 1157.2 1150.6 1144.0 1137.4 1130.6 1123.8 1116.9 1109.9 1102.8 1095.5 1088.2 1080.8 1073.3 1065.7 1057.9 1050.0 1042.0 1033.8 1025.5 1017.0 1008.3 999.4 990.4 981.1 971.6 961.8 951.7 941.3 930.6 919.5 907.9 895.8 883.2 870.0 856.1 841.4 825.7 809.0 760.4 694.1 475.6
boiling point
0.42472 0.33259 0.26354 0.21110 0.17079 0.16719 0.15732 0.14512 0.13404 0.12398 0.11482 0.10647 0.09884 0.09187 0.08548 0.07962 0.07424 0.06930 0.06474 0.06054 0.05667 0.05309 0.04977 0.04670 0.04385 0.04121 0.03875 0.03646 0.03433 0.03235 0.03049 0.02876 0.02714 0.02562 0.02420 0.02287 0.02162 0.02044 0.01934 0.01830 0.01732 0.01639 0.01552 0.01469 0.01392 0.01318 0.01248 0.01182 0.01119 0.01059 0.01002 0.00948 0.00897 0.00848 0.00801 0.00756 0.00712 0.00671 0.00631 0.00592 0.00555 0.00464 0.00372 0.00210
139.63 145.31 151.07 156.90 162.81 163.42 165.20 167.60 170.01 172.43 174.87 177.32 179.79 182.27 184.76 187.26 189.78 192.31 194.86 197.42 200.00 202.59 205.20 207.82 210.45 213.10 215.77 218.45 221.15 223.87 226.60 229.34 232.11 234.89 237.69 240.51 243.35 246.21 249.09 251.98 254.90 257.84 260.81 263.80 266.81 269.85 272.92 276.02 279.15 282.32 285.53 288.77 292.06 295.39 298.78 302.22 305.72 309.30 312.95 316.69 320.54 330.81 342.79 369.55
329.85 333.21 336.58 339.95 343.32 343.67 344.67 346.02 347.36 348.71 350.05 351.39 352.73 354.06 355.39 356.72 358.04 359.36 360.68 361.99 363.29 364.59 365.88 367.16 368.44 369.70 370.96 372.21 373.45 374.67 375.89 377.09 378.28 379.45 380.61 381.75 382.87 383.98 385.06 386.13 387.17 388.19 389.18 390.14 391.08 391.98 392.85 393.68 394.48 395.23 395.93 396.58 397.18 397.71 398.18 398.57 398.87 399.07 399.16 399.11 398.90 397.40 393.32 369.55
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
0.7573 0.7825 0.8074 0.8321 0.8566 0.8591 0.8663 0.8760 0.8857 0.8954 0.9050 0.9146 0.9242 0.9338 0.9433 0.9528 0.9623 0.9717 0.9812 0.9906 1.0000 1.0094 1.0187 1.0281 1.0374 1.0467 1.0560 1.0653 1.0746 1.0838 1.0931 1.1023 1.1115 1.1208 1.1300 1.1392 1.1484 1.1576 1.1668 1.1759 1.1851 1.1943 1.2035 1.2128 1.2220 1.2312 1.2405 1.2498 1.2592 1.2685 1.2779 1.2874 1.2969 1.3065 1.3162 1.3260 1.3359 1.3459 1.3561 1.3664 1.3770 1.4049 1.4370 1.5087
1.128 1.143 1.157 1.173 1.188 1.190 1.195 1.201 1.207 1.214 1.220 1.227 1.234 1.240 1.247 1.254 1.261 1.268 1.275 1.282 1.289 1.297 1.304 1.312 1.320 1.327 1.335 1.344 1.352 1.361 1.369 1.378 1.387 1.397 1.407 1.417 1.427 1.438 1.449 1.461 1.473 1.486 1.500 1.515 1.531 1.548 1.566 1.586 1.607 1.631 1.656 1.685 1.717 1.752 1.792 1.837 1.890 1.951 2.025 2.114 2.227 2.702 4.186
1.124 1.125 1.126 1.127 1.129 1.129 1.130 1.131 1.132 1.133 1.135 1.136 1.137 1.139 1.141 1.142 1.144 1.146 1.149 1.151 1.153 1.156 1.159 1.162 1.165 1.168 1.172 1.176 1.180 1.185 1.190 1.195 1.201 1.207 1.213 1.220 1.228 1.236 1.245 1.254 1.265 1.276 1.289 1.302 1.317 1.333 1.351 1.371 1.393 1.418 1.445 1.477 1.512 1.553 1.600 1.655 1.719 1.797 1.891 2.009 2.158 2.803 4.820
1.6098 1.6060 1.6031 1.6007 1.5990 1.5988 1.5984 1.5980 1.5976 1.5973 1.5970 1.5968 1.5967 1.5967 1.5967 1.5968 1.5969 1.5970 1.5973 1.5975 1.5978 1.5981 1.5985 1.5989 1.5993 1.5998 1.6003 1.6008 1.6013 1.6018 1.6024 1.6029 1.6034 1.6040 1.6045 1.6051 1.6056 1.6061 1.6066 1.6071 1.6075 1.6079 1.6083 1.6087 1.6089 1.6092 1.6094 1.6095 1.6095 1.6095 1.6093 1.6091 1.6087 1.6082 1.6076 1.6068 1.6058 1.6045 1.6030 1.6011 1.5989 1.5909 1.5762 1.5087
0.746 0.762 0.778 0.794 0.811 0.813 0.818 0.825 0.832 0.839 0.847 0.854 0.862 0.869 0.877 0.885 0.893 0.901 0.909 0.918 0.926 0.935 0.944 0.953 0.962 0.972 0.982 0.992 1.002 1.013 1.024 1.035 1.047 1.060 1.073 1.086 1.101 1.116 1.132 1.149 1.167 1.186 1.207 1.229 1.252 1.277 1.305 1.335 1.367 1.403 1.442 1.485 1.534 1.589 1.651 1.724 1.808 1.907 2.028 2.176 2.364 3.168 5.688
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
783 759 736 713 690 688 681 672 664 655 646 637 628 620 611 602 594 585 576 568 559 551 542 533 525 516 508 499 490 482 473 464 456 447 438 429 421 412 403 394 385 376 367 358 349 340 331 322 312 303 293 283 272 262 251 240 229 217 206 194 182 150 116 0
132.1 132.9 133.7 134.3 134.9 134.9 135.0 135.2 135.4 135.5 135.6 135.7 135.7 135.8 135.8 135.8 135.7 135.7 135.6 135.5 135.4 135.2 135.1 134.8 134.6 134.3 134.1 133.7 133.4 133.0 132.6 132.1 131.7 131.1 130.6 130.0 129.4 128.7 128.1 127.3 126.6 125.7 124.9 124.0 123.1 122.1 121.0 120.0 118.8 117.6 116.4 115.1 113.7 112.3 110.8 109.3 107.6 105.9 104.2 102.3 100.3 94.9 88.4 0.00
369.8 347.6 327.0 307.9 290.1 288.3 283.3 276.7 270.3 264.0 258.0 252.0 246.2 240.6 235.1 229.7 224.5 219.4 214.4 209.5 204.7 200.1 195.5 191.0 186.7 182.4 178.2 174.1 170.1 166.1 162.3 158.5 154.7 151.1 147.5 144.0 140.5 137.1 133.8 130.5 127.2 124.0 120.9 117.8 114.7 111.7 108.7 105.8 102.8 99.9 97.0 94.2 91.3 88.4 85.6 82.7 79.8 76.9 74.0 71.1 68.1 60.2 51.0 —
7.68 7.96 8.23 8.49 8.75 8.78 8.85 8.96 9.06 9.16 9.25 9.35 9.45 9.55 9.64 9.74 9.83 9.93 10.02 10.12 10.21 10.31 10.40 10.50 10.59 10.69 10.78 10.88 10.98 11.07 11.17 11.27 11.37 11.48 11.58 11.69 11.79 11.90 12.01 12.13 12.25 12.37 12.49 12.62 12.75 12.89 13.03 13.18 13.34 13.50 13.67 13.85 14.05 14.25 14.47 14.70 14.96 15.23 15.54 15.88 16.26 17.47 19.49 —
Thermal Cond., Surface mW/(m·K) Tension, Temp., Liquid Vapor mN/m °C 89.0 87.1 85.3 83.5 81.7 81.5 81.0 80.3 79.6 78.9 78.2 77.5 76.8 76.2 75.5 74.8 74.1 73.5 72.8 72.1 71.5 70.8 70.2 69.5 68.9 68.2 67.6 67.0 66.3 65.7 65.1 64.5 63.8 63.2 62.6 62.0 61.4 60.8 60.2 59.6 59.0 58.5 57.9 57.3 56.7 56.1 55.6 55.0 54.4 53.8 53.3 52.7 52.1 51.5 51.0 50.4 49.8 49.2 48.7 48.1 47.6 46.4 46.6
7.64 8.04 8.43 8.83 9.22 9.26 9.38 9.54 9.70 9.86 10.02 10.17 10.33 10.49 10.65 10.82 10.98 11.14 11.30 11.47 11.63 11.80 11.97 12.14 12.31 12.48 12.65 12.83 13.01 13.19 13.38 13.57 13.76 13.96 14.16 14.37 14.58 14.80 15.03 15.26 15.50 15.76 16.02 16.29 16.57 16.87 17.19 17.52 17.87 18.25 18.65 19.08 19.54 20.04 20.59 21.19 21.86 22.60 23.43 24.38 25.48 29.23 36.40
17.09 16.27 15.46 14.67 13.89 13.81 13.58 13.27 12.97 12.66 12.36 12.06 11.76 11.47 11.18 10.88 10.60 10.31 10.02 9.74 9.46 9.19 8.91 8.64 8.37 8.10 7.83 7.57 7.31 7.05 6.80 6.55 6.30 6.05 5.81 5.56 5.33 5.09 4.86 4.63 4.40 4.18 3.96 3.74 3.53 3.32 3.12 2.92 2.72 2.52 2.33 2.15 1.97 1.79 1.62 1.45 1.29 1.14 0.99 0.84 0.71 0.39 0.14 0.00
–50 –45 –40 –35 –30 –29.49 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 94.70
cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 14 Pressure-Enthalpy Diagram for Refrigerant 1234ze(E)
30.30
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.31
Refrigerant 1234ze(E) (Trans-1,3,3,3-Tetrafluoropropene) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp., °C –50 –45 –40 –35 –30 –25 –20 –18.97b –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 95 100 105 109.36c bNormal
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.02092 0.02793 0.03675 0.04768 0.06109 0.07735 0.09687 0.10132 0.10569 0.11512 0.12520 0.13596 0.14744 0.15965 0.17263 0.18642 0.20105 0.21655 0.23296 0.25031 0.26863 0.28797 0.30836 0.32983 0.35243 0.37619 0.40114 0.42734 0.45482 0.48362 0.51377 0.54533 0.57833 0.61281 0.64882 0.68640 0.72560 0.76645 0.80901 0.85332 0.89943 0.94738 0.99722 1.0490 1.1028 1.1586 1.2165 1.2766 1.3388 1.4033 1.4702 1.5393 1.6110 1.6851 1.7617 1.8410 1.9230 2.0077 2.2321 2.4755 2.7395 3.0260 3.3378 3.6349
1375.3 1362.4 1349.5 1336.5 1323.3 1309.9 1296.4 1293.6 1290.9 1285.4 1279.9 1274.3 1268.7 1263.1 1257.4 1251.7 1245.9 1240.1 1234.3 1228.4 1222.4 1216.4 1210.4 1204.3 1198.1 1191.9 1185.6 1179.3 1172.8 1166.4 1159.8 1153.2 1146.4 1139.6 1132.8 1125.8 1118.7 1111.5 1104.2 1096.8 1089.3 1081.6 1073.8 1065.9 1057.8 1049.5 1041.1 1032.5 1023.7 1014.7 1005.5 996.0 986.2 976.2 965.9 955.2 944.1 932.7 901.9 867.2 826.9 777.3 708.1 489.2
boiling point
0.76762 0.58586 0.45337 0.35533 0.28176 0.22584 0.18282 0.17525 0.16843 0.15540 0.14357 0.13281 0.12302 0.11409 0.10593 0.09847 0.09163 0.08537 0.07961 0.07431 0.06944 0.06494 0.06079 0.05695 0.05340 0.05012 0.04707 0.04423 0.04160 0.03915 0.03687 0.03475 0.03276 0.03091 0.02918 0.02755 0.02603 0.02461 0.02327 0.02202 0.02084 0.01972 0.01868 0.01769 0.01676 0.01588 0.01505 0.01427 0.01353 0.01282 0.01216 0.01152 0.01092 0.01035 0.00980 0.00929 0.00879 0.00832 0.00722 0.00623 0.00531 0.00444 0.00356 0.00204
136.20 142.42 148.66 154.94 161.25 167.60 173.99 175.31 176.56 179.13 181.71 184.30 186.90 189.50 192.11 194.73 197.36 200.00 202.65 205.30 207.97 210.64 213.32 216.02 218.72 221.44 224.16 226.90 229.65 232.41 235.19 237.98 240.78 243.59 246.42 249.27 252.13 255.00 257.89 260.80 263.73 266.68 269.64 272.63 275.63 278.66 281.71 284.78 287.88 291.01 294.16 297.34 300.56 303.80 307.09 310.41 313.77 317.19 325.95 335.12 344.89 355.59 368.23 395.54
348.81 352.37 355.94 359.51 363.09 366.65 370.20 370.93 371.62 373.03 374.44 375.85 377.25 378.64 380.04 381.42 382.80 384.18 385.55 386.91 388.27 389.62 390.96 392.29 393.61 394.93 396.23 397.53 398.81 400.09 401.35 402.60 403.83 405.06 406.27 407.46 408.64 409.80 410.95 412.07 413.18 414.26 415.33 416.37 417.38 418.37 419.33 420.26 421.16 422.03 422.86 423.65 424.40 425.10 425.76 426.36 426.90 427.38 428.25 428.52 427.93 425.95 421.12 395.54
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
0.7429 0.7704 0.7975 0.8241 0.8503 0.8761 0.9015 0.9067 0.9116 0.9216 0.9316 0.9415 0.9513 0.9612 0.9709 0.9807 0.9904 1.0000 1.0096 1.0192 1.0287 1.0382 1.0476 1.0571 1.0664 1.0758 1.0851 1.0944 1.1037 1.1129 1.1221 1.1313 1.1405 1.1497 1.1588 1.1679 1.1770 1.1861 1.1952 1.2043 1.2134 1.2224 1.2315 1.2405 1.2496 1.2587 1.2677 1.2768 1.2859 1.2950 1.3041 1.3133 1.3225 1.3317 1.3410 1.3503 1.3596 1.3691 1.3930 1.4177 1.4435 1.4715 1.5040 1.5744
1.240 1.246 1.251 1.258 1.265 1.272 1.280 1.282 1.284 1.287 1.291 1.294 1.298 1.302 1.306 1.310 1.314 1.318 1.323 1.327 1.332 1.337 1.342 1.347 1.353 1.358 1.364 1.370 1.376 1.382 1.389 1.396 1.403 1.410 1.418 1.426 1.434 1.443 1.452 1.462 1.472 1.482 1.493 1.505 1.517 1.530 1.544 1.558 1.574 1.591 1.609 1.628 1.649 1.672 1.698 1.726 1.757 1.792 1.901 2.062 2.327 2.859 4.573
1.115 1.115 1.115 1.116 1.117 1.118 1.120 1.121 1.121 1.122 1.123 1.125 1.126 1.127 1.128 1.130 1.132 1.133 1.135 1.137 1.139 1.141 1.144 1.146 1.149 1.152 1.155 1.158 1.161 1.165 1.169 1.173 1.177 1.181 1.186 1.192 1.197 1.203 1.210 1.217 1.224 1.232 1.241 1.250 1.260 1.272 1.284 1.297 1.312 1.328 1.345 1.365 1.386 1.411 1.438 1.468 1.503 1.542 1.671 1.868 2.206 2.905 5.179
1.6956 1.6907 1.6865 1.6831 1.6803 1.6782 1.6766 1.6763 1.6761 1.6756 1.6752 1.6749 1.6747 1.6745 1.6744 1.6743 1.6743 1.6743 1.6743 1.6744 1.6746 1.6748 1.6750 1.6752 1.6755 1.6758 1.6761 1.6765 1.6768 1.6772 1.6776 1.6780 1.6784 1.6788 1.6792 1.6796 1.6801 1.6805 1.6809 1.6813 1.6816 1.6820 1.6823 1.6826 1.6829 1.6831 1.6833 1.6835 1.6836 1.6836 1.6836 1.6835 1.6834 1.6831 1.6828 1.6824 1.6818 1.6811 1.6787 1.6749 1.6691 1.6600 1.6438 1.5744
0.755 0.766 0.777 0.789 0.801 0.814 0.826 0.829 0.832 0.837 0.843 0.848 0.854 0.860 0.865 0.871 0.877 0.884 0.890 0.897 0.903 0.910 0.917 0.924 0.932 0.939 0.947 0.955 0.963 0.971 0.980 0.989 0.999 1.008 1.018 1.029 1.039 1.051 1.063 1.075 1.088 1.101 1.116 1.131 1.147 1.165 1.184 1.204 1.225 1.249 1.274 1.302 1.333 1.367 1.404 1.445 1.492 1.545 1.712 1.963 2.386 3.258 6.122
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
867 845 822 800 777 755 733 728 724 715 706 697 688 679 671 662 653 644 635 626 618 609 600 591 582 574 565 556 547 538 530 521 512 503 495 486 477 468 459 450 441 432 423 414 404 395 386 376 367 357 347 337 327 317 307 296 286 275 248 219 189 157 121 0
132.9 133.9 134.9 135.8 136.6 137.3 137.9 138.0 138.1 138.3 138.4 138.6 138.7 138.8 138.9 139.0 139.0 139.1 139.1 139.1 139.0 139.0 138.9 138.8 138.6 138.5 138.3 138.1 137.8 137.6 137.3 137.0 136.6 136.2 135.8 135.4 134.9 134.4 133.9 133.3 132.7 132.0 131.4 130.6 129.9 129.1 128.3 127.4 126.5 125.5 124.5 123.4 122.3 121.2 119.9 118.7 117.3 116.0 112.2 108.0 103.3 97.9 91.7 0.0
515.9 481.3 449.3 419.5 391.9 366.2 342.3 337.7 333.3 324.5 315.9 307.6 299.6 291.7 284.1 276.7 269.6 262.6 255.8 249.3 242.9 236.7 230.6 224.8 219.1 213.5 208.2 202.9 197.9 192.9 188.1 183.4 178.9 174.5 170.2 166.0 161.9 157.9 154.0 150.2 146.5 142.9 139.4 136.0 132.6 129.3 126.1 122.9 119.8 116.8 113.8 110.8 107.9 105.1 102.2 99.4 96.6 93.8 86.9 79.8 72.4 64.2 54.3 —
7.66 7.98 8.29 8.60 8.91 9.21 9.52 9.58 9.64 9.76 9.88 10.00 10.12 10.24 10.36 10.48 10.60 10.72 10.84 10.96 11.08 11.20 11.32 11.44 11.56 11.68 11.80 11.92 12.05 12.17 12.29 12.41 12.53 12.66 12.78 12.90 13.03 13.15 13.28 13.41 13.53 13.66 13.79 13.92 14.05 14.18 14.32 14.45 14.59 14.73 14.87 15.02 15.16 15.32 15.47 15.63 15.79 15.96 16.42 16.96 17.63 18.58 20.27 —
Thermal Cond., Surface mW/(m·K) Tension, Temp., Liquid Vapor mN/m °C 102.5 100.4 98.4 96.4 94.5 92.5 90.6 90.2 89.8 89.0 88.3 87.5 86.8 86.0 85.3 84.5 83.8 83.1 82.3 81.6 80.9 80.2 79.4 78.7 78.0 77.3 76.6 75.9 75.2 74.6 73.9 73.2 72.5 71.9 71.2 70.5 69.9 69.2 68.6 67.9 67.3 66.6 66.0 65.4 64.7 64.1 63.5 62.8 62.2 61.6 61.0 60.4 59.8 59.2 58.6 58.0 57.4 56.8 55.3 53.9 52.6 51.6 52.2
7.70 8.09 8.48 8.88 9.27 9.66 10.04 10.12 10.20 10.35 10.50 10.66 10.81 10.97 11.12 11.28 11.43 11.58 11.74 11.90 12.05 12.21 12.37 12.53 12.69 12.85 13.01 13.17 13.34 13.51 13.67 13.85 14.02 14.20 14.38 14.57 14.76 14.95 15.15 15.35 15.56 15.78 16.00 16.24 16.48 16.73 17.00 17.27 17.56 17.87 18.19 18.54 18.90 19.30 19.72 20.17 20.66 21.20 22.78 24.85 27.76 32.26 41.13
20.06 19.26 18.46 17.67 16.89 16.12 15.36 15.20 15.05 14.75 14.45 14.15 13.85 13.56 13.26 12.97 12.67 12.38 12.09 11.81 11.52 11.24 10.95 10.67 10.39 10.11 9.84 9.56 9.29 9.02 8.75 8.48 8.21 7.95 7.69 7.43 7.17 6.91 6.66 6.40 6.15 5.91 5.66 5.42 5.18 4.94 4.70 4.47 4.24 4.01 3.79 3.56 3.34 3.13 2.92 2.71 2.50 2.30 1.81 1.35 0.92 0.53 0.20 0.00
–50 –45 –40 –35 –30 –25 –20 –18.97 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80 85 90 95 100 105 109.36
cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 15 Pressure-Enthalpy Diagram for Refrigerant 404A
30.32
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.33
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 404A [R-125/143a/134a (44/52/4)] Properties of Liquid on Bubble Line and Vapor on Dew Line Pressure, MPa
Temperature,* Enthalpy, Density, Volume, °C kJ/kg kg/m3 m3/kg Bubble Dew Liquid Vapor Liquid Vapor
Velocity of Sound, m/s cp /cv Liquid Vapor Liquid Vapor Vapor Liquid Vapor
Thermal Cond., Surface PresmW/(m · K) Tension, sure, Liquid Vapor Liquid Vapor mN/m MPa
0.005 0.006 0.007 0.008 0.009 0.01 0.02 0.04 0.06 0.08 0.1 0.10132b 0.12 0.14 0.16 0.18 0.2 0.22 0.24 0.26 0.28 0.3 0.32 0.34 0.36 0.38 0.4 0.42 0.44 0.46 0.48 0.5 0.55 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 2.1 2.2 2.3 2.4 2.5 2.6 2.7 2.8 2.9 3.0 3.2 3.4 3.729c
–93.70 –91.48 –89.56 –87.86 –86.32 –84.93 –75.05 –63.85 –56.57 –51.03 –46.50 –46.22 –42.63 –39.24 –36.20 –33.45 –30.93 –28.59 –26.42 –24.37 –22.45 –20.62 –18.89 –17.24 –15.66 –14.15 –12.69 –11.29 –9.94 –8.64 –7.37 –6.15 –3.24 –0.53 2.02 4.42 6.70 8.87 10.94 12.92 14.81 16.64 20.09 23.32 26.35 29.22 31.93 34.51 36.97 39.33 41.58 43.75 45.84 47.85 49.80 51.68 53.50 55.26 56.97 58.63 60.24 61.81 64.82 67.67 72.05
0.4716 0.4865 0.4993 0.5106 0.5206 0.5296 0.5917 0.6587 0.7007 0.7320 0.7571 0.7586 0.7783 0.7967 0.8130 0.8277 0.8411 0.8534 0.8649 0.8755 0.8855 0.8950 0.9039 0.9125 0.9206 0.9283 0.9358 0.9429 0.9498 0.9564 0.9628 0.9690 0.9837 0.9973 1.0101 1.0222 1.0336 1.0444 1.0547 1.0646 1.0741 1.0832 1.1005 1.1166 1.1318 1.1462 1.1599 1.1730 1.1856 1.1977 1.2095 1.2208 1.2319 1.2427 1.2532 1.2635 1.2737 1.2837 1.2937 1.3036 1.3135 1.3234 1.3438 1.3657 1.4455
764.9 727.8 697.9 673.0 651.7 633.3 523.7 431.3 383.8 352.7 329.8 328.5 311.9 297.3 285.0 274.4 265.1 256.9 249.5 242.8 236.7 231.1 225.9 221.1 216.6 212.4 208.4 204.7 201.2 197.8 194.6 191.6 184.6 178.2 172.5 167.2 162.4 157.9 153.6 149.7 146.0 142.5 136.1 130.2 124.9 119.9 115.3 111.0 107.0 103.2 99.5 96.1 92.7 89.5 86.5 83.5 80.5 77.7 74.9 72.1 69.3 66.5 60.9 54.7 —
–92.50 –90.32 –88.42 –86.74 –85.22 –83.84 –74.08 –62.97 –55.75 –50.25 –45.74 –45.47 –41.90 –38.53 –35.51 –32.78 –30.27 –27.94 –25.78 –23.75 –21.83 –20.02 –18.29 –16.65 –15.08 –13.57 –12.12 –10.73 –9.39 –8.09 –6.83 –5.61 –2.72 –0.02 2.52 4.91 7.18 9.34 11.40 13.37 15.26 17.08 20.52 23.73 26.75 29.60 32.30 34.87 37.32 39.67 41.91 44.07 46.15 48.15 50.08 51.95 53.76 55.51 57.21 58.86 60.46 62.01 64.99 67.81 72.05
1447.1 1440.6 1434.9 1429.9 1425.4 1421.3 1392.4 1359.4 1337.7 1321.0 1307.1 1306.3 1295.1 1284.5 1275.0 1266.2 1258.0 1250.4 1243.3 1236.5 1230.1 1223.9 1218.0 1212.4 1206.9 1201.6 1196.5 1191.6 1186.7 1182.0 1177.5 1173.0 1162.3 1152.0 1142.3 1132.9 1123.8 1115.1 1106.5 1098.2 1090.2 1082.2 1066.9 1052.0 1037.5 1023.4 1009.5 995.7 982.1 968.6 955.1 941.6 928.1 914.4 900.6 886.5 872.2 857.5 842.4 826.8 810.5 793.4 755.6 709.8 486.5
*Temperatures on ITS-90 scale
3.05794 2.57690 2.22992 1.96748 1.76182 1.59620 0.83425 0.43619 0.29837 0.22779 0.18467 0.18240 0.15551 0.13443 0.11846 0.10592 0.09581 0.08748 0.08049 0.07454 0.06941 0.06494 0.06101 0.05752 0.05441 0.05162 0.04909 0.04680 0.04471 0.04279 0.04103 0.03940 0.03584 0.03284 0.03029 0.02809 0.02618 0.02449 0.02300 0.02166 0.02046 0.01937 0.01749 0.01590 0.01455 0.01338 0.01236 0.01146 0.01066 0.00994 0.00930 0.00871 0.00817 0.00768 0.00723 0.00680 0.00641 0.00604 0.00569 0.00536 0.00505 0.00475 0.00417 0.00361 0.00206
81.16 83.85 86.19 88.26 90.13 91.83 103.81 117.48 126.44 133.31 138.97 139.31 143.83 148.12 151.97 155.49 158.73 161.75 164.57 167.23 169.75 172.14 174.43 176.61 178.71 180.73 182.68 184.56 186.38 188.15 189.86 191.53 195.51 199.26 202.81 206.18 209.41 212.49 215.46 218.32 221.09 223.77 228.89 233.75 238.37 242.81 247.07 251.19 255.17 259.05 262.83 266.52 270.14 273.70 277.20 280.66 284.09 287.50 290.89 294.29 297.70 301.15 308.25 315.97 343.92
311.61 312.92 314.06 315.07 315.99 316.83 322.78 329.58 334.00 337.36 340.08 340.25 342.40 344.41 346.20 347.81 349.28 350.63 351.88 353.04 354.13 355.15 356.12 357.03 357.90 358.72 359.51 360.26 360.98 361.67 362.33 362.96 364.45 365.81 367.06 368.21 369.28 370.27 371.19 372.05 372.85 373.59 374.94 376.12 377.14 378.02 378.78 379.42 379.95 380.38 380.70 380.92 381.05 381.08 381.01 380.83 380.55 380.15 379.62 378.96 378.14 377.15 374.49 370.45 343.92
Entropy, kJ/(kg · K)
1.7532 1.7450 1.7382 1.7324 1.7273 1.7229 1.6953 1.6707 1.6578 1.6494 1.6434 1.6430 1.6387 1.6349 1.6318 1.6292 1.6270 1.6250 1.6233 1.6217 1.6203 1.6190 1.6179 1.6168 1.6158 1.6149 1.6141 1.6133 1.6125 1.6118 1.6112 1.6105 1.6091 1.6078 1.6066 1.6055 1.6044 1.6035 1.6025 1.6016 1.6007 1.5999 1.5982 1.5965 1.5949 1.5932 1.5914 1.5896 1.5878 1.5858 1.5838 1.5817 1.5794 1.5770 1.5745 1.5718 1.5689 1.5658 1.5624 1.5587 1.5547 1.5503 1.5397 1.5255 1.4455 b Bubble
Specific Heat cp , kJ/(kg · K) 1.220 1.218 1.216 1.215 1.214 1.214 1.215 1.225 1.234 1.243 1.251 1.252 1.259 1.266 1.273 1.279 1.285 1.291 1.297 1.303 1.308 1.313 1.319 1.324 1.329 1.334 1.339 1.344 1.349 1.353 1.358 1.363 1.374 1.386 1.397 1.409 1.420 1.432 1.443 1.455 1.466 1.478 1.503 1.528 1.554 1.582 1.611 1.643 1.676 1.712 1.751 1.794 1.841 1.893 1.952 2.019 2.095 2.183 2.288 2.414 2.569 2.765 3.381 4.771 —
0.640 0.646 0.651 0.655 0.660 0.663 0.691 0.725 0.749 0.767 0.784 0.785 0.798 0.811 0.823 0.834 0.844 0.855 0.864 0.873 0.882 0.891 0.899 0.907 0.915 0.923 0.931 0.938 0.946 0.953 0.960 0.967 0.984 1.001 1.018 1.034 1.051 1.067 1.084 1.100 1.117 1.134 1.169 1.206 1.244 1.285 1.329 1.376 1.426 1.481 1.541 1.607 1.681 1.763 1.856 1.962 2.085 2.229 2.401 2.609 2.868 3.197 4.233 6.536 —
1.163 1.162 1.161 1.161 1.160 1.160 1.159 1.159 1.161 1.163 1.166 1.166 1.169 1.171 1.174 1.177 1.179 1.182 1.185 1.188 1.190 1.193 1.196 1.199 1.202 1.205 1.208 1.211 1.214 1.217 1.220 1.223 1.231 1.239 1.247 1.256 1.264 1.274 1.283 1.293 1.303 1.313 1.336 1.360 1.386 1.414 1.445 1.478 1.515 1.556 1.601 1.652 1.709 1.774 1.847 1.932 2.032 2.149 2.289 2.459 2.672 2.944 3.797 5.689 —
998 980 966 953 942 933 870 807 770 742 719 718 700 684 669 656 644 633 623 613 604 595 587 579 572 565 558 551 545 538 532 527 513 500 488 476 465 455 445 435 426 417 400 384 368 354 340 327 314 301 289 277 266 254 243 232 222 211 200 190 179 169 148 126 —
132.9 133.6 134.1 134.6 135.0 135.4 137.9 140.4 141.7 142.6 143.2 143.2 143.6 143.9 144.1 144.2 144.3 144.3 144.4 144.3 144.3 144.2 144.1 144.0 143.9 143.8 143.7 143.5 143.4 143.2 143.0 142.8 142.4 141.9 141.3 140.8 140.2 139.6 139.0 138.3 137.7 137.1 135.7 134.4 133.0 131.6 130.1 128.7 127.2 125.7 124.1 122.6 121.0 119.4 117.8 116.2 114.5 112.9 111.2 109.5 107.7 106.0 102.3 98.5 —
and dew points at one standard atmosphere
Viscosity, Pa · s
7.32 7.41 7.48 7.55 7.61 7.66 8.04 8.47 8.74 8.95 9.12 9.13 9.26 9.39 9.50 9.60 9.69 9.78 9.86 9.94 10.01 10.08 10.15 10.21 10.27 10.33 10.39 10.44 10.49 10.55 10.60 10.65 10.77 10.88 10.99 11.10 11.20 11.30 11.40 11.50 11.59 11.69 11.88 12.07 12.26 12.45 12.65 12.84 13.05 13.25 13.47 13.70 13.93 14.18 14.44 14.72 15.02 15.34 15.69 16.07 16.49 16.95 18.09 19.68 —
122.5 121.2 120.1 119.2 118.3 117.5 112.2 106.4 102.8 100.1 98.0 97.8 96.2 94.6 93.2 91.9 90.8 89.7 88.7 87.8 87.0 86.2 85.4 84.7 84.0 83.3 82.7 82.1 81.5 81.0 80.4 79.9 78.7 77.5 76.5 75.5 74.5 73.6 72.8 72.0 71.2 70.4 69.0 67.7 66.5 65.3 64.2 63.1 62.1 61.2 60.2 59.3 58.5 57.6 56.8 56.0 55.3 54.5 53.8 53.2 52.6 52.0 51.2 51.3 —
6.15 6.28 6.40 6.50 6.60 6.68 7.31 8.05 8.55 8.93 9.25 9.27 9.53 9.78 10.01 10.21 10.40 10.58 10.75 10.91 11.06 11.21 11.34 11.48 11.61 11.73 11.85 11.97 12.08 12.19 12.30 12.41 12.66 12.91 13.16 13.41 13.65 13.89 14.12 14.35 14.59 14.82 15.29 15.76 16.23 16.71 17.21 17.72 18.24 18.80 19.37 19.98 20.62 21.31 22.04 22.83 23.69 24.62 25.65 26.79 28.06 29.51 33.17 38.73 —
17.78 17.58 17.40 17.24 17.09 16.96 16.00 14.85 14.08 13.48 12.98 12.95 12.55 12.17 11.82 11.51 11.21 10.94 10.69 10.45 10.22 10.01 9.81 9.61 9.42 9.24 9.07 8.90 8.74 8.58 8.43 8.28 7.93 7.61 7.30 7.01 6.74 6.48 6.23 5.99 5.76 5.54 5.13 4.75 4.39 4.06 3.75 3.45 3.17 2.91 2.66 2.43 2.21 2.00 1.80 1.61 1.43 1.26 1.10 0.94 0.80 0.67 0.43 0.23 0.00
0.005 0.006 0.007 0.008 0.009 0.01 0.02 0.04 0.06 0.08 0.1 0.10132 0.12 0.14 0.16 0.18 0.2 0.22 0.24 0.26 0.28 0.3 0.32 0.34 0.36 0.38 0.4 0.42 0.44 0.46 0.48 0.5 0.55 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 2.1 2.2 2.3 2.4 2.5 2.6 2.7 2.8 2.9 3.0 3.2 3.4 3.729
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 16 Pressure-Enthalpy Diagram for Refrigerant 407C
30.34
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.35
Refrigerant 407C [R-32/125/134a (23/25/52)] Properties of Liquid on Bubble Line and Vapor on Dew Line
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure, MPa 0.01 0.02 0.04 0.06 0.08 0.1 0.10132b 0.12 0.14 0.16 0.18 0.2 0.22 0.24 0.26 0.28 0.3 0.32 0.34 0.36 0.38 0.4 0.42 0.44 0.46 0.48 0.5 0.55 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 2.1 2.2 2.3 2.4 2.5 2.6 2.7 2.8 2.9 3.0 3.2 3.4 3.6 3.8 4.0 4.2 4.63c
Thermal Cond., mW/(m · K)
–82.45 –72.50 –61.25 –53.96 –48.42 –43.90 –43.63 –40.05 –36.67 –33.65 –30.92 –28.41 –26.09 –23.93 –21.90 –19.99 –18.19 –16.47 –14.83 –13.27 –11.77 –10.33 –8.94 –7.61 –6.31 –5.06 –3.85 –0.98 1.70 4.22 6.60 8.85 11.00 13.04 15.00 16.88 18.69 22.11 25.30 28.30 31.14 33.83 36.39 38.84 41.18 43.43 45.59 47.67 49.68 51.63 53.51 55.34 57.11 58.83 60.51 62.14 63.73 66.80 69.73 72.53 75.22 77.82 80.32 86.03
0.5259 0.5910 0.6612 0.7050 0.7374 0.7635 0.7650 0.7854 0.8043 0.8211 0.8362 0.8499 0.8625 0.8742 0.8851 0.8954 0.9050 0.9141 0.9228 0.9310 0.9389 0.9465 0.9537 0.9607 0.9674 0.9739 0.9801 0.9950 1.0087 1.0216 1.0338 1.0452 1.0561 1.0665 1.0764 1.0859 1.0950 1.1122 1.1283 1.1434 1.1577 1.1713 1.1843 1.1967 1.2086 1.2200 1.2311 1.2418 1.2522 1.2624 1.2723 1.2819 1.2914 1.3006 1.3097 1.3187 1.3276 1.3450 1.3622 1.3795 1.3970 1.4152 1.4348 1.5384
1495.5 1466.7 1433.7 1412.0 1395.3 1381.5 1380.7 1369.7 1359.1 1349.7 1341.0 1333.0 1325.5 1318.4 1311.8 1305.5 1299.5 1293.7 1288.2 1282.9 1277.8 1272.8 1268.0 1263.4 1258.8 1254.4 1250.1 1239.8 1230.0 1220.7 1211.7 1203.1 1194.9 1186.8 1179.1 1171.5 1164.1 1149.9 1136.2 1123.0 1110.2 1097.7 1085.5 1073.5 1061.7 1050.0 1038.5 1027.1 1015.7 1004.4 993.1 981.8 970.5 959.0 947.5 935.9 924.1 899.9 874.6 847.6 818.1 785.1 746.0 484.2
*Temperatures on ITS-90 scale
1.89703 0.99017 0.51705 0.35346 0.26975 0.21865 0.21595 0.18411 0.15916 0.14025 0.12542 0.11347 0.10362 0.09536 0.08833 0.08227 0.07699 0.07235 0.06824 0.06457 0.06127 0.05830 0.05559 0.05313 0.05087 0.04879 0.04687 0.04267 0.03915 0.03615 0.03356 0.03131 0.02933 0.02757 0.02600 0.02460 0.02332 0.02111 0.01926 0.01768 0.01631 0.01512 0.01408 0.01315 0.01231 0.01157 0.01089 0.01027 0.00971 0.00919 0.00871 0.00827 0.00786 0.00747 0.00711 0.00677 0.00645 0.00587 0.00533 0.00484 0.00439 0.00395 0.00352 0.00207
90.48 103.24 117.72 127.17 134.39 140.31 140.67 145.39 149.86 153.86 157.51 160.87 163.99 166.91 169.65 172.24 174.71 177.06 179.30 181.45 183.52 185.52 187.44 189.30 191.11 192.86 194.56 198.61 202.42 206.02 209.44 212.71 215.83 218.83 221.71 224.50 227.19 232.34 237.20 241.82 246.24 250.48 254.57 258.51 262.33 266.05 269.66 273.19 276.64 280.02 283.34 286.60 289.82 292.99 296.12 299.23 302.31 308.43 314.54 320.71 327.02 333.64 340.83 378.48
366.78 372.75 379.47 383.77 386.99 389.59 389.75 391.78 393.68 395.36 396.86 398.22 399.47 400.62 401.69 402.69 403.62 404.49 405.32 406.10 406.85 407.55 408.23 408.87 409.48 410.07 410.64 411.95 413.15 414.25 415.25 416.18 417.03 417.83 418.57 419.25 419.89 421.03 422.03 422.89 423.63 424.27 424.80 425.25 425.61 425.89 426.10 426.23 426.29 426.28 426.20 426.06 425.85 425.57 425.21 424.79 424.29 423.06 421.46 419.45 416.91 413.66 409.34 378.48
1.9471 1.9104 1.8761 1.8573 1.8445 1.8349 1.8343 1.8273 1.8210 1.8156 1.8110 1.8069 1.8033 1.8000 1.7970 1.7942 1.7917 1.7894 1.7872 1.7851 1.7832 1.7814 1.7796 1.7780 1.7764 1.7750 1.7735 1.7702 1.7672 1.7644 1.7618 1.7594 1.7571 1.7550 1.7529 1.7509 1.7491 1.7455 1.7421 1.7389 1.7358 1.7328 1.7298 1.7269 1.7241 1.7212 1.7184 1.7155 1.7126 1.7097 1.7068 1.7038 1.7007 1.6976 1.6944 1.6911 1.6877 1.6805 1.6726 1.6639 1.6540 1.6424 1.6281 1.5384 b Bubble
Specific Heat cp , kJ/(kg · K)
Viscosity, Pa · s
Surface Prescp /cv Tension, sure, Liquid Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor mN/m MPa
–74.81 –65.02 –53.95 –46.79 –41.34 –36.90 –36.63 –33.11 –29.79 –26.83 –24.15 –21.69 –19.41 –17.29 –15.31 –13.43 –11.66 –9.98 –8.38 –6.85 –5.38 –3.97 –2.61 –1.31 –0.04 1.18 2.36 5.17 7.79 10.24 12.56 14.76 16.85 18.84 20.74 22.56 24.32 27.63 30.73 33.63 36.37 38.97 41.43 43.78 46.03 48.18 50.25 52.24 54.15 56.00 57.79 59.51 61.19 62.81 64.38 65.91 67.40 70.25 72.94 75.50 77.92 80.21 82.37 86.03
Entropy, kJ/(kg · K)
Velocity of Sound, m/s
Enthalpy, Temperature,* Density, Volume, kJ/kg °C kg/m3 m3/kg Liquid Vapor Liquid Vapor Bubble Dew
1.281 1.283 1.291 1.299 1.306 1.312 1.312 1.318 1.324 1.329 1.334 1.339 1.344 1.349 1.354 1.358 1.362 1.367 1.371 1.375 1.379 1.383 1.387 1.391 1.395 1.399 1.403 1.413 1.422 1.432 1.441 1.451 1.460 1.469 1.479 1.488 1.498 1.517 1.537 1.557 1.578 1.600 1.622 1.645 1.669 1.695 1.722 1.750 1.780 1.813 1.847 1.884 1.924 1.968 2.016 2.069 2.128 2.268 2.451 2.701 3.065 3.647 4.726 —
0.668 0.694 0.727 0.750 0.769 0.786 0.787 0.800 0.813 0.825 0.837 0.848 0.858 0.868 0.877 0.886 0.895 0.903 0.911 0.919 0.927 0.934 0.942 0.949 0.956 0.963 0.970 0.987 1.004 1.020 1.036 1.052 1.067 1.082 1.098 1.113 1.128 1.159 1.190 1.222 1.255 1.289 1.324 1.361 1.400 1.440 1.483 1.529 1.577 1.629 1.684 1.744 1.810 1.881 1.958 2.044 2.139 2.365 2.657 3.050 3.613 4.486 6.029 —
1.182 1.181 1.182 1.184 1.187 1.190 1.190 1.193 1.196 1.199 1.201 1.204 1.207 1.210 1.213 1.216 1.219 1.222 1.224 1.227 1.230 1.233 1.236 1.239 1.242 1.245 1.248 1.255 1.262 1.270 1.278 1.286 1.294 1.302 1.310 1.319 1.327 1.346 1.365 1.385 1.406 1.428 1.452 1.477 1.504 1.533 1.564 1.597 1.633 1.671 1.713 1.758 1.808 1.863 1.923 1.990 2.065 2.243 2.475 2.789 3.239 3.935 5.159 —
1008 953 893 856 828 806 804 787 770 755 742 730 719 708 698 689 680 672 664 656 649 642 635 629 622 616 610 596 583 571 559 548 537 527 518 508 499 482 466 451 436 423 409 397 385 373 361 350 339 329 318 308 298 288 279 269 259 240 222 203 184 165 146 —
149.1 151.8 154.6 156.1 157.1 157.8 157.8 158.3 158.7 159.0 159.3 159.5 159.6 159.7 159.8 159.8 159.8 159.8 159.8 159.8 159.7 159.7 159.6 159.5 159.4 159.3 159.2 158.9 158.6 158.2 157.8 157.4 157.0 156.6 156.1 155.6 155.2 154.2 153.2 152.1 151.0 150.0 148.8 147.7 146.6 145.4 144.2 143.0 141.8 140.6 139.4 138.2 136.9 135.6 134.4 133.1 131.7 129.1 126.4 123.6 120.7 117.7 114.6 —
and dew points at one standard atmosphere
779.8 632.8 513.1 453.1 414.4 386.2 384.6 364.3 346.6 331.8 319.1 308.0 298.2 289.5 281.6 274.4 267.8 261.8 256.1 250.9 246.0 241.4 237.1 233.0 229.1 225.4 221.9 213.9 206.7 200.1 194.1 188.6 183.6 178.8 174.4 170.3 166.4 159.2 152.8 146.9 141.5 136.5 131.8 127.5 123.4 119.6 115.9 112.5 109.2 106.0 103.0 100.0 97.2 94.5 91.9 89.3 86.8 81.9 77.1 72.5 67.7 62.9 57.6 —
8.43 8.83 9.28 9.57 9.79 9.97 9.98 10.12 10.25 10.37 10.48 10.57 10.66 10.75 10.83 10.90 10.97 11.04 11.11 11.17 11.23 11.28 11.34 11.39 11.45 11.50 11.54 11.66 11.77 11.88 11.98 12.08 12.17 12.26 12.35 12.44 12.52 12.68 12.84 13.01 13.15 13.31 13.47 13.62 13.78 13.94 14.10 14.27 14.44 14.62 14.79 14.98 15.17 15.37 15.58 15.80 16.03 16.53 17.09 17.75 18.52 19.48 20.75 —
151.5 145.4 138.5 134.1 130.7 128.1 127.9 125.8 123.8 122.0 120.4 119.0 117.6 116.4 115.2 114.2 113.1 112.2 111.2 110.4 109.5 108.7 107.9 107.2 106.5 105.8 105.1 103.5 102.1 100.7 99.4 98.2 97.1 96.0 94.9 93.9 93.0 91.1 89.5 87.9 86.4 85.0 83.7 82.4 81.2 80.1 78.9 77.9 76.8 75.8 74.9 73.9 73.0 72.1 71.3 70.4 69.6 68.1 66.6 65.2 64.0 63.1 62.7 —
6.94 7.52 8.19 8.64 8.99 9.28 9.29 9.52 9.75 9.94 10.13 10.29 10.45 10.60 10.74 10.87 10.99 11.11 11.23 11.35 11.46 11.57 11.68 11.78 11.88 11.98 12.08 12.31 12.54 12.75 12.96 13.17 13.37 13.58 13.78 13.98 14.18 14.59 14.99 15.39 15.80 16.22 16.64 17.07 17.52 17.98 18.45 18.94 19.45 19.98 20.54 21.12 21.73 22.38 23.06 23.79 24.56 26.26 28.22 30.54 33.33 36.86 41.59 —
24.75 22.93 20.91 19.62 18.65 17.87 17.82 17.21 16.63 16.12 15.66 15.24 14.86 14.50 14.16 13.85 13.56 13.28 13.01 12.76 12.52 12.29 12.07 11.85 11.65 11.45 11.26 10.81 10.40 10.01 9.64 9.30 8.98 8.67 8.38 8.11 7.84 7.35 6.89 6.47 6.07 5.70 5.35 5.02 4.71 4.42 4.14 3.87 3.62 3.38 3.15 2.93 2.72 2.52 2.33 2.14 1.96 1.63 1.33 1.05 0.79 0.56 0.36 0.00
0.01 0.02 0.04 0.06 0.08 0.1 0.10132 0.12 0.14 0.16 0.18 0.2 0.22 0.24 0.26 0.28 0.3 0.32 0.34 0.36 0.38 0.4 0.42 0.44 0.46 0.48 0.5 0.55 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 1.0 1.1 1.2 1.3 1.4 1.5 1.6 1.7 1.8 1.9 2.0 2.1 2.2 2.3 2.4 2.5 2.6 2.7 2.8 2.9 3.0 3.2 3.4 3.6 3.8 4.0 4.2 4.63
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 17 Pressure-Enthalpy Diagram for Refrigerant 410A
30.36
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.37
Refrigerant 410A [R-32/125 (50/50)] Properties of Liquid on Bubble Line and Vapor on Dew Line
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure, MPa
Temperature,* Enthalpy, Density, Volume, °C kJ/kg kg/m3 m3/kg Bubble Dew Liquid Vapor Liquid Vapor
0.01 –88.23 0.02 –78.79 0.04 –68.12 0.06 –61.22 0.08 –55.98 0.1 –51.70 0.10132b –51.44 0.12 –48.06 0.14 –44.87 0.16 –42.02 0.18 –39.44 0.2 –37.07 0.22 –34.89 0.24 –32.85 0.26 –30.94 0.28 –29.14 0.3 –27.44 0.32 –25.82 0.34 –24.28 0.36 –22.81 0.38 –21.40 0.4 –20.04 0.42 –18.74 0.44 –17.48 0.46 –16.27 0.48 –15.10 0.5 –13.96 0.55 –11.26 0.6 –8.74 0.65 –6.38 0.7 –4.15 0.75 –2.04 0.8 –0.03 0.85 1.89 0.9 3.72 0.95 5.48 1.0 7.17 1.1 10.36 1.2 13.34 1.3 16.15 1.4 18.79 1.5 21.30 1.6 23.68 1.7 25.96 1.8 28.13 1.9 30.22 2.0 32.22 2.1 34.16 2.2 36.02 2.3 37.82 2.4 39.56 2.5 41.25 2.6 42.89 2.7 44.48 2.8 46.02 2.9 47.53 3.0 48.99 3.2 51.81 3.4 54.49 3.6 57.05 3.8 59.50 4.0 61.85 4.2 64.10 71.36 4.903c
–88.14 –78.70 –68.04 –61.14 –55.90 –51.62 –51.36 –47.98 –44.79 –41.94 –39.36 –36.99 –34.80 –32.76 –30.85 –29.05 –27.35 –25.73 –24.19 –22.72 –21.31 –19.95 –18.65 –17.39 –16.18 –15.00 –13.86 –11.16 –8.64 –6.28 –4.05 –1.93 0.08 1.99 3.83 5.58 7.27 10.47 13.46 16.26 18.91 21.41 23.80 26.07 28.25 30.34 32.34 34.28 36.14 37.94 39.68 41.37 43.00 44.59 46.14 47.64 49.10 51.91 54.59 57.15 59.59 61.93 64.17 71.36
1460.6 1432.9 1401.1 1380.0 1363.9 1350.5 1349.7 1339.0 1328.8 1319.6 1311.2 1303.4 1296.2 1289.4 1283.0 1276.9 1271.1 1265.5 1260.2 1255.0 1250.1 1245.3 1240.6 1236.1 1231.8 1227.5 1223.3 1213.4 1203.9 1194.9 1186.3 1178.1 1170.1 1162.4 1154.9 1147.6 1140.5 1126.8 1113.7 1101.0 1088.8 1076.9 1065.2 1053.8 1042.6 1031.6 1020.7 1009.9 999.2 988.6 978.0 967.5 957.0 946.4 935.8 925.2 914.5 892.6 870.0 846.3 821.0 793.5 762.6 459.5
*Temperatures on ITS-90 scale
2.09888 1.09659 0.57309 0.39193 0.29918 0.24256 0.23957 0.20427 0.17661 0.15565 0.13921 0.12595 0.11503 0.10587 0.09807 0.09135 0.08550 0.08035 0.07579 0.07172 0.06806 0.06476 0.06176 0.05902 0.05652 0.05421 0.05209 0.04743 0.04352 0.04019 0.03732 0.03482 0.03262 0.03068 0.02894 0.02738 0.02596 0.02351 0.02145 0.01970 0.01819 0.01687 0.01571 0.01468 0.01376 0.01293 0.01218 0.0115 0.01088 0.01031 0.00978 0.00929 0.00883 0.00841 0.00802 0.00764 0.00729 0.00665 0.00607 0.00555 0.00506 0.00460 0.00417 0.00218
76.56 89.26 103.64 113.00 120.14 125.99 126.34 130.99 135.39 139.34 142.93 146.23 149.29 152.15 154.84 157.38 159.80 162.10 164.29 166.40 168.43 170.38 172.26 174.08 175.84 177.55 179.21 183.17 186.89 190.40 193.74 196.92 199.96 202.88 205.69 208.40 211.02 216.03 220.76 225.26 229.56 233.68 237.65 241.48 245.19 248.79 252.29 255.71 259.05 262.32 265.52 268.67 271.77 274.82 277.84 280.82 283.78 289.62 295.43 301.26 307.16 313.24 319.65 368.55
378.76 384.25 390.29 394.10 396.92 399.17 399.31 401.05 402.67 404.09 405.36 406.50 407.53 408.49 409.36 410.18 410.94 411.65 412.32 412.95 413.54 414.10 414.64 415.14 415.63 416.09 416.53 417.54 418.46 419.28 420.03 420.71 421.33 421.89 422.41 422.88 423.31 424.07 424.68 425.19 425.59 425.89 426.11 426.25 426.31 426.31 426.24 426.10 425.90 425.64 425.33 424.95 424.51 424.02 423.47 422.85 422.18 420.62 418.78 416.60 414.03 410.97 407.24 368.55
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
Thermal Cond., Velocity of Viscosity, Surface PresPa·s mW/(m·K) Sound, m/s cp /cv Tension, sure, Vapor Liquid Vapor Liquid Vapor Liquid Vapor mN/m MPa
0.4588 0.5258 0.5978 0.6426 0.6758 0.7024 0.7040 0.7247 0.7441 0.7612 0.7766 0.7905 0.8034 0.8153 0.8264 0.8368 0.8466 0.8558 0.8646 0.8703 0.8810 0.8887 0.8960 0.9031 0.9099 0.9165 0.9228 0.9379 0.9518 0.9649 0.9772 0.9888 0.9998 1.0103 1.0204 1.0300 1.0392 1.0567 1.0730 1.0883 1.1027 1.1165 1.1296 1.1421 1.1542 1.1657 1.1769 1.1878 1.1983 1.2085 1.2185 1.2282 1.2377 1.2470 1.2561 1.2651 1.2740 1.2913 1.3085 1.3254 1.3425 1.3600 1.3783 1.5181
1.344 1.345 1.351 1.358 1.364 1.369 1.370 1.375 1.380 1.385 1.390 1.395 1.399 1.404 1.408 1.413 1.417 1.421 1.426 1.430 1.434 1.438 1.443 1.447 1.451 1.455 1.459 1.469 1.479 1.489 1.499 1.509 1.519 1.529 1.540 1.550 1.560 1.581 1.603 1.624 1.647 1.670 1.694 1.719 1.745 1.772 1.800 1.830 1.861 1.894 1.929 1.967 2.008 2.052 2.100 2.153 2.211 2.348 2.522 2.752 3.070 3.541 4.306 —
1.227 1.228 1.231 1.235 1.239 1.243 1.244 1.247 1.251 1.255 1.259 1.263 1.266 1.270 1.274 1.277 1.281 1.285 1.288 1.292 1.295 1.299 1.303 1.306 1.310 1.313 1.317 1.326 1.335 1.344 1.354 1.363 1.373 1.382 1.392 1.402 1.413 1.434 1.457 1.481 1.506 1.532 1.560 1.590 1.621 1.655 1.690 1.728 1.769 1.813 1.860 1.911 1.966 2.026 2.091 2.163 2.243 2.429 2.663 2.970 3.386 3.987 4.929 —
2.0927 2.0432 1.9956 1.9687 1.9500 1.9358 1.9350 1.9243 1.9147 1.9065 1.8993 1.8928 1.8871 1.8818 1.8770 1.8726 1.8685 1.8647 1.8611 1.8577 1.8545 1.8514 1.8486 1.8458 1.8432 1.8407 1.8383 1.8326 1.8275 1.8227 1.8183 1.8141 1.8102 1.8065 1.8030 1.7996 1.7964 1.7903 1.7846 1.7792 1.7741 1.7691 1.7644 1.7597 1.7552 1.7508 1.7464 1.7421 1.7379 1.7336 1.7294 1.7251 1.7209 1.7166 1.7123 1.7079 1.7035 1.6944 1.6849 1.6747 1.6638 1.6517 1.6380 1.5181 b Bubble
0.668 0.696 0.734 0.762 0.785 0.805 0.807 0.823 0.839 0.854 0.868 0.881 0.893 0.904 0.916 0.926 0.936 0.946 0.956 0.965 0.975 0.983 0.992 1.001 1.009 1.017 1.025 1.045 1.064 1.083 1.101 1.119 1.136 1.154 1.171 1.188 1.205 1.239 1.274 1.31 1.347 1.385 1.424 1.465 1.509 1.555 1.603 1.655 1.709 1.768 1.831 1.898 1.971 2.050 2.136 2.230 2.333 2.575 2.879 3.276 3.815 4.596 5.826 —
1004 958 906 872 847 826 824 808 792 778 765 753 743 732 723 714 705 697 689 682 675 668 661 655 649 643 637 623 610 597 586 574 564 554 544 535 525 508 492 477 462 448 435 422 410 398 386 375 364 353 343 332 322 313 303 293 284 265 247 229 210 192 173 —
159.7 162.8 165.8 167.5 168.7 169.5 169.5 170.1 170.6 170.9 171.2 171.5 171.6 171.8 171.9 172.0 172.0 172.0 172.1 172.1 172.0 172.0 172.0 171.9 171.8 171.8 171.7 171.4 171.2 170.9 170.5 170.2 169.8 169.4 169.0 168.6 168.1 167.2 166.3 165.4 164.4 163.4 162.4 161.4 160.3 159.3 158.2 157.1 156.0 154.9 153.8 152.6 151.5 150.3 149.1 147.9 146.7 144.2 141.7 139.0 136.3 133.4 130.4 —
and dew points at one standard atmosphere
669.9 552.9 454.8 404.6 371.8 347.8 346.4 329.0 313.8 300.9 289.9 280.3 271.8 264.2 257.2 251.0 245.2 239.8 234.9 230.3 226.0 221.9 218.1 214.5 211.1 207.8 204.7 197.6 191.2 185.3 180.0 175.1 170.6 166.4 162.4 158.7 155.3 148.8 143.1 137.8 133.0 128.5 124.3 120.4 116.8 113.3 110.1 107.0 104.0 101.2 98.5 95.9 93.4 91.0 88.6 86.3 84.1 79.9 75.7 71.7 67.7 63.7 59.4 —
8.29 8.71 9.17 9.47 9.70 9.88 9.90 10.04 10.18 10.30 10.41 10.51 10.61 10.70 10.78 10.86 10.93 11.00 11.07 11.13 11.19 11.25 11.31 11.36 11.42 11.47 11.52 11.64 11.75 11.86 11.96 12.06 12.15 12.24 12.33 12.41 12.49 12.65 12.81 12.95 13.12 13.23 13.38 13.52 13.66 13.81 13.95 14.10 14.25 14.40 14.55 14.71 14.87 15.03 15.21 15.38 15.57 15.96 16.39 16.87 17.43 18.08 18.87 —
177.3 170.8 163.3 158.3 154.6 151.5 151.3 148.9 146.6 144.6 142.8 141.1 139.5 138.1 136.7 135.5 134.3 133.10 132.10 131.00 130.10 129.10 128.20 127.30 126.50 125.70 124.90 123.10 121.40 119.70 118.20 116.80 115.50 114.20 113.00 111.80 110.70 108.60 106.70 104.80 103.10 101.50 99.98 98.53 97.15 95.82 94.56 93.34 92.17 91.05 89.96 88.91 87.89 86.91 85.96 85.04 84.14 82.42 80.81 79.29 77.90 76.68 75.77 —
7.44 7.79 8.21 8.50 8.73 8.93 8.94 9.11 9.26 9.40 9.53 9.66 9.77 9.88 9.98 10.08 10.18 10.27 10.36 10.46 10.55 10.64 10.73 10.82 10.91 10.99 11.08 11.28 11.48 11.68 11.88 12.07 12.26 12.45 12.64 12.82 13.01 13.39 13.79 14.19 14.60 15.03 15.46 15.91 16.38 16.86 17.36 17.88 18.42 18.99 19.58 20.21 20.87 21.56 22.29 23.07 23.89 25.70 27.77 30.17 33.02 36.48 40.86 —
24.72 22.91 20.90 19.62 18.66 17.88 17.84 17.23 16.65 16.15 15.69 15.27 14.89 14.54 14.21 13.90 13.60 13.33 13.06 12.81 12.57 12.35 12.13 11.92 11.71 11.52 11.33 10.89 10.47 10.09 9.73 9.39 9.07 8.77 8.48 8.21 7.95 7.46 7.01 6.59 6.20 5.83 5.49 5.16 4.86 4.57 4.29 4.03 3.78 3.54 3.31 3.10 2.89 2.69 2.50 2.31 2.14 1.81 1.50 1.22 0.97 0.74 0.53 0.00
0.01 0.02 0.04 0.06 0.08 0.1 0.10132 0.12 0.14 0.16 0.18 0.2 0.22 0.24 0.26 0.28 0.30 0.32 0.34 0.36 0.38 0.40 0.42 0.44 0.46 0.48 0.50 0.55 0.60 0.65 0.70 0.75 0.80 0.85 0.90 0.95 1.00 1.10 1.20 1.30 1.40 1.50 1.60 1.70 1.80 1.90 2.00 2.10 2.20 2.30 2.40 2.50 2.60 2.70 2.80 2.90 3.00 3.20 3.40 3.60 3.80 4.00 4.20 4.903
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 18 Pressure-Enthalpy Diagram for Refrigerant 507A
30.38
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.39
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 507A [R-125/143a (50/50)] Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg Temp.,* sure,** kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor °C
Liquid Vapor
–100 –95 –90 –85 –80 –75 –70 –65 –60 –55 –50 –48 –46.74b –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 55 60 65 70 70.62c
0.4323 0.4669 0.5003 0.5327 0.5642 0.5950 0.6250 0.6545 0.6833 0.7116 0.7395 0.7505 0.7574 0.7615 0.7724 0.7832 0.7940 0.8047 0.8153 0.8260 0.8365 0.8470 0.8575 0.8679 0.8783 0.8886 0.8989 0.9091 0.9193 0.9295 0.9397 0.9498 0.9599 0.9699 0.9800 0.9900 1.0000 1.0100 1.0199 1.0299 1.0398 1.0498 1.0597 1.0696 1.0796 1.0895 1.0995 1.1094 1.1194 1.1294 1.1394 1.1495 1.1595 1.1697 1.1799 1.1901 1.2004 1.2108 1.2213 1.2320 1.2427 1.2536 1.2818 1.3120 1.3465 1.4007 1.4358
0.00295 0.00458 0.00693 0.01019 0.01464 0.02058 0.02836 0.03837 0.05105 0.06688 0.08638 0.09533 0.10132 0.10499 0.11541 0.12662 0.13867 0.15159 0.16542 0.18022 0.19602 0.21287 0.23081 0.24989 0.27016 0.29167 0.31446 0.33858 0.36408 0.39102 0.41945 0.44941 0.48096 0.51416 0.54906 0.58571 0.62417 0.66450 0.70676 0.75099 0.79728 0.84566 0.89622 0.94900 1.00410 1.06150 1.12140 1.18370 1.24860 1.31610 1.38640 1.45940 1.53520 1.61400 1.69580 1.78070 1.86880 1.96020 2.05490 2.15310 2.25480 2.36030 2.64090 2.94760 3.28380 3.65570 3.70500
1476.9 1461.7 1446.8 1431.9 1417.1 1402.3 1387.4 1372.5 1357.4 1342.3 1326.9 1320.7 1316.8 1314.5 1308.2 1301.9 1295.6 1289.2 1282.8 1276.3 1269.7 1263.2 1256.5 1249.8 1243.1 1236.3 1229.4 1222.5 1215.4 1208.4 1201.2 1193.9 1186.6 1179.2 1171.7 1164.0 1156.3 1148.5 1140.5 1132.4 1124.2 1115.9 1107.4 1098.7 1089.9 1080.9 1071.7 1062.4 1052.8 1043.0 1032.9 1022.6 1011.9 1001.0 989.7 978.1 966.0 953.5 940.5 926.9 912.7 897.7 856.2 806.1 739.1 599.6 490.8
4.92920 3.25360 2.20850 1.53750 1.09510 0.79638 0.59012 0.44482 0.34056 0.26444 0.20801 0.18960 0.17902 0.17313 0.15836 0.14510 0.13317 0.12240 0.11268 0.10388 0.09590 0.08865 0.08205 0.07604 0.07055 0.06553 0.06094 0.05673 0.05286 0.04931 0.04603 0.04301 0.04023 0.03765 0.03527 0.03306 0.03101 0.02910 0.02733 0.02568 0.02415 0.02271 0.02138 0.02012 0.01895 0.01785 0.01683 0.01586 0.01495 0.01410 0.01329 0.01253 0.01182 0.01114 0.01050 0.00989 0.00932 0.00877 0.00825 0.00776 0.00728 0.00683 0.00578 0.00480 0.00384 0.00260 0.00204
*Temperatures on ITS-90 scale
74.41 80.48 86.51 92.53 98.54 104.57 110.60 116.66 122.74 128.87 135.03 137.51 139.07 139.99 142.48 144.99 147.49 150.01 152.54 155.08 157.63 160.18 162.75 165.33 167.92 170.52 173.13 175.76 178.39 181.04 183.71 186.39 189.08 191.78 194.51 197.25 200.00 202.77 205.56 208.37 211.20 214.04 216.91 219.80 222.71 225.65 228.61 231.60 234.61 237.66 240.73 243.84 246.98 250.16 253.39 256.65 259.96 263.33 266.74 270.23 273.78 277.41 286.91 297.28 309.30 328.32 340.45
303.90 306.85 309.83 312.83 315.85 318.88 321.92 324.96 328.00 331.03 334.05 335.25 336.01 336.45 337.65 338.84 340.03 341.21 342.38 343.55 344.72 345.88 347.03 348.17 349.30 350.43 351.54 352.65 353.75 354.83 355.91 356.97 358.02 359.06 360.08 361.08 362.07 363.05 364.00 364.94 365.85 366.75 367.61 368.46 369.28 370.07 370.83 371.55 372.25 372.91 373.52 374.10 374.63 375.11 375.54 375.91 376.22 376.46 376.61 376.68 376.66 376.52 375.54 373.26 368.44 353.47 340.45
Entropy, kJ/(kg·K) 1.7579 1.7377 1.7197 1.7036 1.6893 1.6766 1.6652 1.6552 1.6463 1.6384 1.6314 1.6288 1.6273 1.6264 1.6241 1.6219 1.6198 1.6178 1.6159 1.6141 1.6123 1.6107 1.6092 1.6077 1.6063 1.6049 1.6037 1.6024 1.6013 1.6001 1.5991 1.5980 1.5971 1.5961 1.5952 1.5943 1.5934 1.5925 1.5917 1.5908 1.5900 1.5891 1.5883 1.5874 1.5865 1.5856 1.5846 1.5836 1.5826 1.5815 1.5804 1.5792 1.5779 1.5765 1.5750 1.5734 1.5717 1.5698 1.5678 1.5655 1.5631 1.5603 1.5519 1.5401 1.5215 1.4740 1.4358
Specific Heat cp , kJ/(kg·K) Liquid Vapor 1.219 0.618 1.210 0.631 1.205 0.644 1.203 0.658 1.203 0.672 1.205 0.686 1.208 0.701 1.213 0.716 1.220 0.732 1.227 0.749 1.235 0.766 1.239 0.773 1.241 0.777 1.243 0.780 1.247 0.787 1.251 0.795 1.255 0.803 1.259 0.810 1.264 0.818 1.269 0.826 1.274 0.835 1.279 0.843 1.284 0.852 1.289 0.861 1.295 0.870 1.301 0.879 1.307 0.888 1.313 0.898 1.319 0.908 1.326 0.918 1.333 0.929 1.340 0.940 1.348 0.951 1.355 0.962 1.363 0.974 1.372 0.987 1.381 0.999 1.390 1.012 1.399 1.026 1.410 1.040 1.420 1.055 1.431 1.071 1.443 1.088 1.455 1.105 1.468 1.124 1.482 1.144 1.497 1.165 1.513 1.188 1.530 1.212 1.548 1.239 1.568 1.268 1.589 1.299 1.612 1.333 1.637 1.371 1.664 1.413 1.695 1.459 1.729 1.511 1.767 1.570 1.811 1.638 1.860 1.716 1.918 1.807 1.985 1.915 2.225 2.304 2.677 3.060 3.940 5.190 31.960 44.630
cp /cv Vapor 1.164 1.162 1.161 1.159 1.159 1.158 1.158 1.159 1.160 1.161 1.164 1.165 1.166 1.166 1.167 1.169 1.170 1.172 1.174 1.176 1.178 1.180 1.183 1.186 1.188 1.191 1.195 1.198 1.202 1.206 1.210 1.214 1.219 1.224 1.230 1.236 1.242 1.249 1.256 1.264 1.272 1.282 1.291 1.302 1.314 1.327 1.341 1.356 1.372 1.391 1.411 1.433 1.458 1.485 1.516 1.551 1.591 1.636 1.689 1.750 1.823 1.910 2.228 2.855 4.625 36.780
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C
— 784.2 701.5 631.9 572.7 521.7 477.4 438.5 404.3 373.8 346.5 336.4 330.2 326.7 317.4 308.4 299.8 291.4 283.4 275.7 268.3 261.1 254.1 247.4 240.9 234.5 228.4 222.5 216.8 211.2 205.7 200.5 195.3 190.3 185.5 180.7 176.1 171.6 167.2 162.9 158.7 154.5 150.5 146.6 142.7 138.9 135.1 131.5 127.9 124.3 120.8 117.4 114.0 110.6 107.3 104.0 100.7 97.5 94.3 91.0 87.8 84.6 76.4 67.7 57.9 42.0 —
124.6 121.7 118.8 116.1 113.4 110.8 108.2 105.7 103.2 100.8 98.4 97.4 96.8 96.5 95.5 94.6 93.7 92.7 91.8 90.9 90.0 89.1 88.2 87.3 86.5 85.6 84.7 83.8 83.0 82.1 81.2 80.4 79.5 78.7 77.8 77.0 76.2 75.3 74.5 73.7 72.8 72.0 71.2 70.4 69.6 68.8 67.9 67.1 66.3 65.5 64.7 63.9 63.1 62.2 61.4 60.6 59.8 59.0 58.1 57.3 56.5 55.7 53.6 51.7 50.7 63.2 —
1046 1000 960 925 892 862 833 806 779 754 729 719 713 709 699 690 680 670 661 651 642 632 622 613 603 594 584 575 566 556 547 537 528 518 508 499 489 480 470 460 451 441 431 422 412 402 392 382 372 362 352 341 331 321 310 300 289 278 267 256 245 233 203 171 135 92 0
129.6 131.2 132.6 134.0 135.4 136.6 137.8 138.9 139.8 140.7 141.4 141.6 141.8 141.9 142.1 142.3 142.5 142.6 142.7 142.8 142.9 142.9 143.0 143.0 142.9 142.9 142.8 142.7 142.5 142.3 142.1 141.9 141.6 141.3 141.0 140.6 140.2 139.8 139.3 138.8 138.2 137.6 137.0 136.3 135.6 134.9 134.1 133.2 132.3 131.4 130.4 129.3 128.2 127.1 125.9 124.6 123.2 121.8 120.4 118.8 117.2 115.5 110.8 105.5 99.3 90.7 0.0
**Small deviations from azeotropic behavior occur at some conditions; tabulated pressures are average of bubble and dew-point pressures
— 7.29 7.49 7.68 7.88 8.07 8.27 8.46 8.65 8.84 9.02 9.10 9.15 9.17 9.25 9.32 9.40 9.47 9.55 9.62 9.70 9.77 9.85 9.93 10.00 10.08 10.15 10.23 10.31 10.39 10.47 10.55 10.63 10.71 10.79 10.88 10.97 11.05 11.14 11.23 11.33 11.43 11.52 11.63 11.73 11.86 11.97 12.09 12.22 12.35 12.48 12.62 12.77 12.93 13.10 13.28 13.47 13.68 13.90 14.14 14.41 14.70 15.59 16.86 18.93 25.07 — b Normal
5.77 6.06 6.36 6.67 6.99 7.31 7.63 7.96 8.30 8.65 9.00 9.14 9.23 9.28 9.42 9.57 9.71 9.86 10.01 10.16 10.31 10.46 10.61 10.77 10.93 11.08 11.24 11.40 11.56 11.73 11.89 12.06 12.23 12.41 12.58 12.76 12.96 13.16 13.36 13.57 13.79 14.01 14.24 14.49 14.75 15.01 15.29 15.58 15.89 16.21 16.54 16.90 17.28 17.68 18.12 18.58 19.09 19.63 20.23 20.89 21.62 22.43 24.98 28.77 35.54 67.27 —
boiling point
18.35 17.88 17.41 16.92 16.43 15.92 15.40 14.88 14.34 13.80 13.24 13.02 12.88 12.80 12.57 12.34 12.12 11.89 11.66 11.42 11.19 10.96 10.72 10.49 10.25 10.02 9.78 9.54 9.30 9.06 8.82 8.58 8.34 8.10 7.86 7.62 7.37 7.13 6.89 6.65 6.41 6.17 5.93 5.69 5.45 5.21 4.97 4.74 4.50 4.27 4.04 3.81 3.58 3.35 3.12 2.90 2.68 2.47 2.25 2.04 1.83 1.63 1.15 0.70 0.31 0.02 0.00
–100 –95 –90 –85 –80 –75 –70 –65 –60 –55 –50 –48 –46.74 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 55 60 65 70 70.62
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 19 Pressure-Enthalpy Diagram for Refrigerant 717 (Ammonia)
30.40
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.41
Refrigerant 717 (Ammonia) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor
–77.65a 0.00609 732.9 –70 0.01094 724.7 –60 0.02189 713.6 –50 0.04084 702.1 –40 0.07169 690.2 –38 0.07971 687.7 –36 0.08845 685.3 –34 0.09795 682.8 –33.33b 0.10133 682.0 –32 0.10826 680.3 –30 0.11943 677.8 –28 0.13151 675.3 –26 0.14457 672.8 –24 0.15864 670.3 –22 0.17379 667.7 –20 0.19008 665.1 –18 0.20756 662.6 –16 0.22630 660.0 –14 0.24637 657.3 –12 0.26782 654.7 –10 0.29071 652.1 –8 0.31513 649.4 –6 0.34114 646.7 –4 0.36880 644.0 –2 0.39819 641.3 0 0.42938 638.6 2 0.46246 635.8 4 0.49748 633.1 6 0.53453 630.3 8 0.57370 627.5 10 0.61505 624.6 12 0.65866 621.8 14 0.70463 618.9 16 0.75303 616.0 18 0.80395 613.1 20 0.85748 610.2 22 0.91369 607.2 24 0.97268 604.3 26 1.03450 601.3 28 1.09930 598.2 30 1.16720 595.2 32 1.23820 592.1 34 1.31240 589.0 36 1.39000 585.8 38 1.47090 582.6 40 1.55540 579.4 42 1.64350 576.2 44 1.73530 572.9 46 1.83100 569.6 48 1.93050 566.3 50 2.03400 562.9 55 2.31110 554.2 60 2.61560 545.2 65 2.94910 536.0 70 3.31350 526.3 75 3.71050 516.2 80 4.14200 505.7 85 4.61000 494.5 90 5.11670 482.8 95 5.66430 470.2 100 6.25530 456.6 105 6.89230 441.9 110 7.57830 425.6 115 8.31700 407.2 120 9.11250 385.5 125 9.97002 357.8 130 10.89770 312.3 132.25c 11.33300 225.0
15.602 –143.15 9.0079 –110.81 4.7057 –68.06 2.6277 –24.73 1.5533 19.17 1.4068 28.01 1.2765 36.88 1.1604 45.77 1.1242 48.76 1.0567 54.67 0.96396 63.60 0.88082 72.55 0.80614 81.52 0.73896 90.51 0.67840 99.52 0.62373 108.55 0.57428 117.60 0.52949 126.67 0.48885 135.76 0.45192 144.88 0.41830 154.01 0.38767 163.16 0.35970 172.34 0.33414 181.54 0.31074 190.76 0.28930 200.00 0.26962 209.27 0.25153 218.55 0.23489 227.87 0.21956 237.20 0.20543 246.57 0.19237 255.95 0.18031 265.37 0.16914 274.81 0.15879 284.28 0.14920 293.78 0.14029 303.31 0.13201 312.87 0.12431 322.47 0.11714 332.09 0.11046 341.76 0.10422 351.45 0.09840 361.19 0.09296 370.96 0.08787 380.78 0.08310 390.64 0.07863 400.54 0.07445 410.48 0.07052 420.48 0.06682 430.52 0.06335 440.62 0.05554 466.10 0.04880 491.97 0.04296 518.26 0.03787 545.04 0.03342 572.37 0.02951 600.34 0.02606 629.04 0.02300 658.61 0.02027 689.19 0.01782 721.00 0.01561 754.35 0.01360 789.68 0.01174 827.74 0.00999 869.92 0.00828 919.68 0.00638 992.02 0.00444 1119.22
*Temperatures on ITS-90 scale
1341.23 1355.55 1373.73 1391.19 1407.76 1410.96 1414.11 1417.23 1418.26 1420.29 1423.31 1426.28 1429.21 1432.08 1434.91 1437.68 1440.39 1443.06 1445.66 1448.21 1450.70 1453.14 1455.51 1457.81 1460.06 1462.24 1464.35 1466.40 1468.37 1470.28 1472.11 1473.88 1475.56 1477.17 1478.70 1480.16 1481.53 1482.82 1484.02 1485.14 1486.17 1487.11 1487.95 1488.70 1489.36 1489.91 1490.36 1490.70 1490.94 1491.06 1491.07 1490.57 1489.27 1487.09 1483.94 1479.72 1474.31 1467.53 1459.19 1449.01 1436.63 1421.57 1403.08 1379.99 1350.23 1309.12 1239.32 1119.22
Liquid
Vapor
Velocity of Sound, m/s cp /cv Liquid Vapor Vapor Liquid Vapor
–0.4716 –0.3094 –0.1040 0.0945 0.2867 0.3245 0.3619 0.3992 0.4117 0.4362 0.4730 0.5096 0.5460 0.5821 0.6180 0.6538 0.6893 0.7246 0.7597 0.7946 0.8293 0.8638 0.8981 0.9323 0.9662 1.0000 1.0336 1.0670 1.1003 1.1334 1.1664 1.1992 1.2318 1.2643 1.2967 1.3289 1.3610 1.3929 1.4248 1.4565 1.4881 1.5196 1.5509 1.5822 1.6134 1.6446 1.6756 1.7065 1.7374 1.7683 1.7990 1.8758 1.9523 2.0288 2.1054 2.1823 2.2596 2.3377 2.4168 2.4973 2.5797 2.6647 2.7533 2.8474 2.9502 3.0702 3.2437 3.5542
7.1213 6.9088 6.6602 6.4396 6.2425 6.2056 6.1694 6.1339 6.1221 6.0992 6.0651 6.0317 5.9989 5.9667 5.9351 5.9041 5.8736 5.8437 5.8143 5.7853 5.7569 5.7289 5.7013 5.6741 5.6474 5.6210 5.5951 5.5695 5.5442 5.5192 5.4946 5.4703 5.4463 5.4226 5.3991 5.3759 5.3529 5.3301 5.3076 5.2853 5.2631 5.2412 5.2194 5.1978 5.1763 5.1549 5.1337 5.1126 5.0915 5.0706 5.0497 4.9977 4.9458 4.8939 4.8415 4.7885 4.7344 4.6789 4.6213 4.5612 4.4975 4.4291 4.3542 4.2702 4.1719 4.0483 3.8571 3.5542
4.202 4.245 4.303 4.360 4.414 4.424 4.434 4.444 4.448 4.455 4.465 4.474 4.484 4.494 4.504 4.514 4.524 4.534 4.543 4.553 4.564 4.574 4.584 4.595 4.606 4.617 4.628 4.639 4.651 4.663 4.676 4.689 4.702 4.716 4.730 4.745 4.760 4.776 4.793 4.810 4.828 4.847 4.867 4.888 4.909 4.932 4.956 4.981 5.007 5.034 5.064 5.143 5.235 5.341 5.465 5.610 5.784 5.993 6.250 6.573 6.991 7.555 8.360 9.630 11.940 17.660 54.210
Entropy, kJ/(kg·K)
a Triple
Viscosity, Pa·s
Specific Heat cp , kJ/(kg·K)
point
2.063 1.325 2.086 1.327 2.125 1.330 2.178 1.335 2.244 1.342 2.259 1.343 2.275 1.345 2.291 1.347 2.297 1.348 2.308 1.349 2.326 1.351 2.344 1.353 2.363 1.355 2.383 1.358 2.403 1.360 2.425 1.363 2.446 1.365 2.469 1.368 2.493 1.371 2.517 1.375 2.542 1.378 2.568 1.382 2.594 1.385 2.622 1.389 2.651 1.393 2.680 1.398 2.710 1.402 2.742 1.407 2.774 1.412 2.807 1.417 2.841 1.422 2.877 1.428 2.913 1.434 2.951 1.440 2.990 1.446 3.030 1.453 3.071 1.460 3.113 1.468 3.158 1.475 3.203 1.484 3.250 1.492 3.299 1.501 3.349 1.510 3.401 1.520 3.455 1.530 3.510 1.541 3.568 1.553 3.628 1.565 3.691 1.577 3.756 1.591 3.823 1.605 4.005 1.643 4.208 1.687 4.438 1.739 4.699 1.799 5.001 1.870 5.355 1.955 5.777 2.058 6.291 2.187 6.933 2.349 7.762 2.562 8.877 2.851 10.46 3.26 12.91 3.91 17.21 5.04 27.00 7.62 76.49 20.66
2124 2051 1967 1890 1816 1802 1787 1773 1768 1759 1744 1730 1716 1702 1687 1673 1659 1645 1631 1616 1602 1588 1574 1559 1545 1531 1516 1502 1487 1473 1458 1443 1429 1414 1399 1384 1370 1355 1340 1324 1309 1294 1279 1263 1248 1232 1216 1201 1185 1169 1153 1112 1070 1028 984 940 895 848 800 751 701 649 594 538 477 411 334 0
354.1 360.5 368.4 375.6 382.2 383.4 384.6 385.8 386.2 387.0 388.1 389.2 390.2 391.2 392.2 393.2 394.1 395.0 395.8 396.7 397.5 398.2 398.9 399.6 400.2 400.8 401.4 401.9 402.4 402.8 403.2 403.6 403.9 404.2 404.4 404.6 404.8 404.9 404.9 405.0 404.9 404.8 404.7 404.5 404.3 404.0 403.7 403.3 402.9 402.4 401.9 400.3 398.3 396.0 393.3 390.1 386.5 382.5 377.9 372.7 367.0 360.5 353.3 345.0 335.4 323.6 306.6 0.0
Liquid Vapor 559.6 475.0 391.3 328.9 281.2 273.1 265.3 257.9 255.5 250.8 244.1 237.6 231.4 225.5 219.8 214.4 209.2 204.2 199.3 194.7 190.2 185.9 181.7 177.7 173.8 170.1 166.5 162.9 159.5 156.2 153.0 149.9 146.9 144.0 141.1 138.3 135.6 133.0 130.4 127.9 125.5 123.1 120.7 118.4 116.2 114.0 111.9 109.8 107.8 105.8 103.8 99.0 94.5 90.1 85.9 81.9 78.0 74.2 70.5 66.8 63.2 59.6 56.0 52.3 48.3 43.8 37.3 —
b Normal
6.84 7.03 7.30 7.57 7.86 7.92 7.98 8.03 8.05 8.09 8.15 8.21 8.27 8.33 8.39 8.45 8.51 8.57 8.63 8.69 8.75 8.81 8.87 8.93 8.99 9.06 9.12 9.18 9.24 9.30 9.36 9.43 9.49 9.55 9.61 9.68 9.74 9.80 9.87 9.93 10.00 10.06 10.13 10.19 10.26 10.33 10.39 10.46 10.53 10.60 10.67 10.86 11.05 11.25 11.47 11.70 11.95 12.23 12.55 12.91 13.32 13.82 14.42 15.19 16.21 17.73 20.63 —
boiling point
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 819.0 792.1 757.0 722.3 688.1 681.4 674.6 667.9 665.7 661.3 654.6 648.0 641.5 634.9 628.4 622.0 615.5 609.1 602.8 596.4 590.1 583.9 577.7 571.5 565.3 559.2 553.1 547.1 541.1 535.1 529.1 523.2 517.3 511.5 505.6 499.9 494.1 488.4 482.7 477.0 471.4 465.7 460.1 454.6 449.1 443.5 438.0 432.6 427.1 421.7 416.3 402.9 389.6 376.4 363.2 350.2 337.1 324.1 311.0 297.9 284.8 271.5 258.1 244.6 231.2 219.1 221.9
19.64 19.73 19.93 20.24 20.64 20.73 20.83 20.93 20.97 21.04 21.15 21.26 21.38 21.51 21.63 21.77 21.90 22.05 22.19 22.35 22.50 22.67 22.83 23.00 23.18 23.37 23.55 23.75 23.95 24.15 24.37 24.58 24.81 25.04 25.27 25.52 25.77 26.03 26.29 26.57 26.85 27.14 27.43 27.74 28.05 28.38 28.71 29.06 29.41 29.78 30.16 31.16 32.26 33.47 34.80 36.30 38.00 39.95 42.24 44.99 48.36 52.65 58.33 66.28 78.40 100.01 160.39
62.26 59.10 55.05 51.11 47.26 46.51 45.75 45.00 44.75 44.26 43.52 42.78 42.05 41.32 40.60 39.88 39.16 38.45 37.74 37.04 36.34 35.65 34.96 34.27 33.59 32.91 32.24 31.57 30.91 30.24 29.59 28.94 28.29 27.65 27.01 26.38 25.75 25.12 24.50 23.89 23.28 22.67 22.07 21.47 20.88 20.29 19.71 19.13 18.56 17.99 17.43 16.04 14.69 13.37 12.08 10.83 9.61 8.44 7.30 6.20 5.15 4.15 3.20 2.31 1.50 0.77 0.18 0.00
–77.65 –70 –60 –50 –40 –38 –36 –34 –33.33 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125 130 132.25
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 20 Pressure-Enthalpy Diagram for Refrigerant 718 (Water/Steam)
30.42
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.43
Refrigerant 718 (Water/Steam) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Enthalpy, Density, Volume, kJ/kg kg/m3 m3/kg Temp.,* Pres°C sure, MPa Liquid Vapor Liquid Vapor 0.01a 0.00061 999.8 205.990 5 0.00087 999.9 147.010 10 0.00123 999.7 106.300 15 0.00171 999.1 77.8750 20 0.00234 998.2 57.7570 25 0.00317 997.0 43.3370 30 0.00425 995.6 32.8780 35 0.00563 994.0 25.2050 40 0.00738 992.2 19.5150 45 0.00959 990.2 15.2520 50 0.01235 988.0 12.0270 55 0.01576 985.7 9.5643 60 0.01995 983.2 7.6672 65 0.02504 980.5 6.1935 70 0.03120 977.7 5.0395 75 0.03860 974.8 4.1289 80 0.04741 971.8 3.4052 85 0.05787 968.6 2.8258 90 0.07018 965.3 2.3591 95 0.08461 961.9 1.9806 99.97b 0.10133 958.4 1.6732 100 0.10142 958.3 1.6718 105 0.12090 954.7 1.4184 110 0.14338 950.9 1.2093 115 0.16918 947.1 1.0358 120 0.19867 943.1 0.89121 125 0.23224 939.0 0.77003 130 0.27028 934.8 0.66800 135 0.31323 930.5 0.58173 140 0.36154 926.1 0.50845 145 0.41568 921.6 0.44596 150 0.47616 917.0 0.39245 155 0.54350 912.3 0.34646 160 0.61823 907.4 0.30678 165 0.70093 902.5 0.27243 170 0.79219 897.5 0.24259 175 0.89260 892.3 0.21658 180 1.00280 887.0 0.19384 185 1.12350 881.6 0.17390 190 1.25520 876.1 0.15636 195 1.39880 870.4 0.14089 200 1.55490 864.7 0.12721 205 1.72430 858.8 0.11508 210 1.90770 852.7 0.10429 215 2.10580 846.5 0.09468 220 2.31960 840.2 0.08609 225 2.54970 833.7 0.07840 230 2.79710 827.1 0.07150 235 3.06250 820.3 0.06530 240 3.34690 813.4 0.05970 245 3.65120 806.2 0.05465 250 3.97620 798.9 0.05008 255 4.32290 791.4 0.04594 260 4.69230 783.6 0.04217 265 5.08530 775.7 0.03875 270 5.50300 767.5 0.03562 275 5.94640 759.0 0.03277 280 6.41660 750.3 0.03015 285 6.91470 741.3 0.02776 290 7.44180 731.9 0.02555 295 7.99910 722.2 0.02353 300 8.58790 712.1 0.02166 310 9.86510 690.7 0.01833 320 11.28430 667.1 0.01547 330 12.85810 640.8 0.01298 340 14.60070 610.7 0.01078 350 16.52940 574.7 0.00880 360 18.66600 527.6 0.00695 370 21.04360 451.4 0.00495 373.95c 22.06400 322.0 0.00311 *Temperatures on ITS-90 scale
0.00 21.02 42.02 62.98 83.91 104.83 125.73 146.63 167.53 188.43 209.34 230.26 251.18 272.12 293.07 314.03 335.01 356.01 377.04 398.09 419.06 419.17 440.27 461.42 482.59 503.81 525.07 546.38 567.74 589.16 610.64 632.18 653.79 675.47 697.24 719.08 741.02 763.05 785.19 807.43 829.79 852.27 874.88 897.63 920.53 943.58 966.80 990.19 1013.77 1037.55 1061.55 1085.77 1110.23 1134.96 1159.96 1185.27 1210.90 1236.88 1263.25 1290.03 1317.27 1345.01 1402.22 1462.22 1525.87 1594.53 1670.89 1761.66 1890.69 2084.26
2500.92 2510.06 2519.21 2528.33 2537.43 2546.51 2555.55 2564.55 2573.51 2582.43 2591.29 2600.09 2608.83 2617.50 2626.10 2634.60 2643.02 2651.33 2659.53 2667.61 2675.53 2675.57 2683.39 2691.06 2698.58 2705.93 2713.10 2720.08 2726.87 2733.44 2739.80 2745.93 2751.81 2757.44 2762.81 2767.90 2772.71 2777.21 2781.41 2785.28 2788.82 2792.01 2794.83 2797.27 2799.32 2800.95 2802.15 2802.90 2803.17 2802.96 2802.22 2800.93 2799.07 2796.60 2793.49 2789.69 2785.17 2779.87 2773.73 2766.70 2758.70 2749.64 2727.95 2700.59 2666.03 2621.85 2563.64 2481.49 2334.52 2084.26
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid
Vapor
0.0000 0.0763 0.1511 0.2245 0.2965 0.3672 0.4368 0.5051 0.5724 0.6386 0.7038 0.7680 0.8313 0.8937 0.9551 1.0158 1.0756 1.1346 1.1929 1.2504 1.3069 1.3072 1.3633 1.4188 1.4737 1.5279 1.5816 1.6346 1.6872 1.7392 1.7907 1.8418 1.8924 1.9426 1.9923 2.0417 2.0906 2.1392 2.1875 2.2355 2.2832 2.3305 2.3777 2.4245 2.4712 2.5177 2.5640 2.6101 2.6561 2.7020 2.7478 2.7935 2.8392 2.8849 2.9307 2.9765 3.0224 3.0685 3.1147 3.1612 3.2080 3.2552 3.3510 3.4494 3.5518 3.6601 3.7784 3.9167 4.1112 4.4070
9.1555 4.220 9.0248 4.205 8.8998 4.196 8.7803 4.189 8.6660 4.184 8.5566 4.182 8.4520 4.180 8.3517 4.180 8.2555 4.180 8.1633 4.180 8.0748 4.182 7.9898 4.183 7.9081 4.185 7.8296 4.187 7.7540 4.190 7.6812 4.193 7.6111 4.197 7.5434 4.201 7.4781 4.205 7.4151 4.210 7.3544 4.216 7.3541 4.216 7.2952 4.222 7.2381 4.228 7.1828 4.236 7.1291 4.244 7.0770 4.252 7.0264 4.261 6.9772 4.272 6.9293 4.283 6.8826 4.294 6.8371 4.307 6.7926 4.321 6.7491 4.335 6.7066 4.351 6.6650 4.368 6.6241 4.386 6.5840 4.405 6.5447 4.425 6.5059 4.447 6.4678 4.471 6.4302 4.496 6.3930 4.523 6.3563 4.551 6.3200 4.582 6.2840 4.615 6.2483 4.650 6.2128 4.688 6.1775 4.728 6.1423 4.772 6.1072 4.819 6.0721 4.870 6.0369 4.925 6.0016 4.986 5.9661 5.051 5.9304 5.123 5.8944 5.202 5.8579 5.289 5.8209 5.385 5.7834 5.493 5.7451 5.614 5.7059 5.750 5.6244 6.085 5.5372 6.537 5.4422 7.186 5.3356 8.210 5.2110 10.120 5.0536 15.000 4.8012 45.160 4.4070 a Triple point
Liquid Vapor 1.884 1.889 1.895 1.900 1.906 1.912 1.918 1.925 1.931 1.939 1.947 1.955 1.965 1.975 1.986 1.999 2.012 2.027 2.043 2.061 2.080 2.080 2.101 2.124 2.150 2.177 2.207 2.239 2.274 2.311 2.351 2.394 2.440 2.488 2.540 2.594 2.652 2.713 2.777 2.844 2.915 2.990 3.068 3.150 3.237 3.329 3.426 3.528 3.638 3.754 3.878 4.011 4.153 4.308 4.475 4.656 4.855 5.073 5.314 5.582 5.882 6.220 7.045 8.159 9.753 12.24 16.69 27.36 96.60
Thermal Cond., Velocity of Viscosity, Surface Pa·s mW/(m·K) Sound, m/s cp /cv Tension, Temp.,* Vapor Liquid Vapor Liquid Vapor Liquid Vapor mN/m °C 1.329 1.328 1.328 1.328 1.327 1.327 1.327 1.327 1.327 1.327 1.328 1.328 1.328 1.329 1.330 1.331 1.332 1.333 1.334 1.335 1.337 1.337 1.339 1.341 1.343 1.346 1.349 1.352 1.355 1.359 1.363 1.368 1.373 1.379 1.385 1.392 1.399 1.407 1.416 1.425 1.436 1.447 1.459 1.472 1.486 1.501 1.518 1.536 1.556 1.578 1.601 1.627 1.655 1.686 1.720 1.757 1.798 1.845 1.896 1.954 2.019 2.094 2.277 2.528 2.889 3.45 4.46 6.83 21.15
1402 1426 1447 1466 1482 1497 1509 1520 1529 1536 1542 1547 1551 1553 1555 1555 1554 1553 1550 1547 1543 1543 1538 1533 1527 1520 1512 1504 1496 1486 1476 1466 1455 1443 1431 1419 1405 1392 1378 1363 1348 1332 1316 1299 1282 1264 1246 1228 1209 1189 1169 1148 1127 1105 1083 1060 1037 1013 988 962 936 909 853 793 729 658 578 480 360 0
409.0 1791.2 9.22 412.6 1518.3 9.34 416.2 1306.0 9.46 419.7 1137.6 9.59 423.2 1001.6 9.73 426.6 890.1 9.87 430.0 797.4 10.01 433.4 719.3 10.16 436.7 653.0 10.31 440.0 596.1 10.46 443.2 546.8 10.62 446.4 504.0 10.77 449.5 466.4 10.93 452.6 433.2 11.10 455.6 403.9 11.26 458.5 377.7 11.43 461.4 354.3 11.59 464.2 333.3 11.76 466.9 314.4 11.93 469.6 297.3 12.10 472.2 281.8 12.27 472.2 281.7 12.27 474.7 267.6 12.44 477.1 254.7 12.61 479.5 242.9 12.78 481.7 232.1 12.96 483.9 222.1 13.13 486.0 212.9 13.30 487.9 204.4 13.47 489.8 196.5 13.65 491.6 189.2 13.82 493.3 182.5 13.99 494.8 176.1 14.16 496.3 170.2 14.34 497.6 164.7 14.51 498.9 159.6 14.68 500.0 154.7 14.85 501.0 150.1 15.03 501.9 145.8 15.20 502.7 141.8 15.37 503.4 137.9 15.54 503.9 134.3 15.71 504.3 130.9 15.89 504.6 127.6 16.06 504.8 124.5 16.24 504.8 121.5 16.41 504.6 118.7 16.59 504.4 116.0 16.76 503.9 113.4 16.94 503.3 110.9 17.12 502.6 108.4 17.31 501.6 106.1 17.49 500.5 103.9 17.68 499.2 101.7 17.88 497.7 99.6 18.07 496.0 97.5 18.28 494.1 95.5 18.48 491.9 93.5 18.70 489.5 91.6 18.92 486.9 89.7 19.15 483.9 87.8 19.40 480.7 85.9 19.65 473.3 82.2 20.21 464.4 78.4 20.85 453.7 74.5 21.61 440.7 70.4 22.55 424.4 65.9 23.82 402.4 60.3 25.72 362.8 52.1 29.68 0.0 — — b Normal boiling point
561.0 570.5 580.0 589.3 598.4 607.2 615.5 623.3 630.6 637.3 643.6 649.2 654.3 659.0 663.1 666.8 670.0 672.8 675.3 677.3 679.1 679.1 680.5 681.7 682.6 683.2 683.6 683.7 683.6 683.3 682.8 682.0 681.1 680.0 678.6 677.0 675.3 673.3 671.1 668.8 666.1 663.3 660.3 657.0 653.4 649.7 645.6 641.3 636.7 631.8 626.7 621.2 615.4 609.2 602.8 595.9 588.7 581.1 573.2 565.0 556.3 547.4 528.7 509.2 489.1 468.5 447.4 425.7 425.0
17.07 17.34 17.62 17.92 18.23 18.55 18.89 19.24 19.60 19.97 20.36 20.77 21.19 21.62 22.07 22.53 23.01 23.51 24.02 24.55 25.09 25.10 25.66 26.24 26.85 27.47 28.11 28.76 29.44 30.14 30.86 31.60 32.35 33.13 33.93 34.75 35.59 36.45 37.33 38.24 39.16 40.11 41.09 42.09 43.11 44.17 45.26 46.38 47.53 48.73 49.97 51.26 52.61 54.03 55.53 57.11 58.80 60.61 62.57 64.71 67.05 69.65 75.84 83.91 94.94 110.91 135.95 181.51 323.84
75.65 0.01 74.94 5 74.22 10 73.49 15 72.74 20 71.97 25 71.19 30 70.40 35 69.60 40 68.78 45 67.94 50 67.10 55 66.24 60 65.37 65 64.48 70 63.58 75 62.67 80 61.75 85 60.82 90 59.87 95 58.92 99.97 58.91 100 57.94 105 56.96 110 55.97 115 54.97 120 53.96 125 52.93 130 51.90 135 50.86 140 49.80 145 48.74 150 47.67 155 46.59 160 45.50 165 44.41 170 43.30 175 42.19 180 41.07 185 39.95 190 38.81 195 37.67 200 36.53 205 35.38 210 34.23 215 33.07 220 31.90 225 30.74 230 29.57 235 28.39 240 27.22 245 26.04 250 24.87 255 23.69 260 22.51 265 21.34 270 20.16 275 18.99 280 17.83 285 16.66 290 15.51 295 14.36 300 12.09 310 9.86 320 7.70 330 5.63 340 3.67 350 1.88 360 0.39 370 0.00 373.95 c Critical point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 21 Pressure-Enthalpy Diagram for Refrigerant 744 (Carbon Dioxide)
30.44
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.45
Refrigerant 744 (Carbon Dioxide) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –56.56a –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –19 –18 –17 –16 –15 –14 –13 –12 –11 –10 –9 –8 –7 –6 –5 –4 –3 –2 –1 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 30.98c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.51796 0.68234 0.73949 0.80015 0.86445 0.93252 1.00450 1.08050 1.16070 1.24520 1.33420 1.42780 1.52610 1.62930 1.73750 1.85090 1.96960 2.03100 2.09380 2.15810 2.22370 2.29080 2.35930 2.42940 2.50100 2.57400 2.64870 2.72490 2.80270 2.88210 2.96320 3.04590 3.13030 3.21640 3.30420 3.39380 3.48510 3.57830 3.67330 3.77010 3.86880 3.96950 4.07200 4.17650 4.28310 4.39160 4.50220 4.61490 4.72970 4.84660 4.96580 5.08701 5.21080 5.33680 5.46510 5.59580 5.72910 5.86480 6.00310 6.14400 6.28770 6.43420 6.58370 6.73610 6.89180 7.05090 7.21370 7.37730
1178.5 1154.6 1147.1 1139.6 1132.0 1124.2 1116.4 1108.5 1100.5 1092.4 1084.1 1075.7 1067.2 1058.6 1049.8 1040.8 1031.7 1027.0 1022.3 1017.6 1012.8 1008.0 1003.1 998.1 993.1 988.1 982.9 977.7 972.5 967.1 961.7 956.2 950.6 945.0 939.2 933.4 927.4 921.4 915.2 909.0 902.6 896.0 889.4 882.6 875.6 868.4 861.1 853.6 845.9 837.9 829.7 821.2 812.4 803.3 793.8 783.8 773.4 762.4 750.8 738.4 725.0 710.5 694.5 676.4 655.3 629.4 593.3 467.6
0.07267 0.05579 0.05162 0.04782 0.04435 0.04118 0.03828 0.03562 0.03318 0.03093 0.02886 0.02696 0.02519 0.02356 0.02205 0.02065 0.01934 0.01873 0.01813 0.01756 0.01700 0.01647 0.01595 0.01545 0.01497 0.01450 0.01405 0.01361 0.01319 0.01278 0.01238 0.01200 0.01162 0.01126 0.01091 0.01057 0.01024 0.00992 0.00961 0.00931 0.00901 0.00872 0.00845 0.00817 0.00791 0.00765 0.00740 0.00715 0.00691 0.00668 0.00645 0.00622 0.00600 0.00578 0.00557 0.00536 0.00515 0.00494 0.00474 0.00453 0.00433 0.00412 0.00391 0.00369 0.00346 0.00320 0.00290 0.00214
*Temperatures on ITS-90 scale
80.04 92.94 96.90 100.88 104.87 108.88 112.90 116.95 121.01 125.10 129.20 133.34 137.50 141.69 145.91 150.16 154.45 156.61 158.77 160.95 163.14 165.34 167.55 169.78 172.01 174.26 176.52 178.80 181.09 183.39 185.71 188.05 190.40 192.77 195.16 197.57 200.00 202.45 204.93 207.43 209.95 212.50 215.08 217.69 220.34 223.01 225.73 228.49 231.29 234.13 237.03 239.99 243.01 246.10 249.26 252.52 255.87 259.33 262.93 266.68 270.61 274.78 279.26 284.14 289.62 296.07 304.55 332.25
430.42 432.68 433.29 433.86 434.39 434.88 435.32 435.72 436.07 436.37 436.62 436.82 436.96 437.04 437.06 437.01 436.89 436.81 436.70 436.58 436.44 436.27 436.09 435.89 435.66 435.41 435.14 434.84 434.51 434.17 433.79 433.38 432.95 432.48 431.99 431.46 430.89 430.29 429.65 428.97 428.25 427.48 426.67 425.81 424.89 423.92 422.88 421.79 420.62 419.37 418.05 416.64 415.12 413.50 411.76 409.89 407.87 405.67 403.26 400.63 397.70 394.43 390.71 386.39 381.20 374.61 365.13 332.25
Liquid
Vapor
Liquid Vapor
Velocity of Sound, m/s cp /cv Vapor Liquid Vapor
0.5213 0.5794 0.5968 0.6142 0.6314 0.6486 0.6656 0.6826 0.6995 0.7163 0.7331 0.7498 0.7665 0.7831 0.7997 0.8163 0.8328 0.8411 0.8494 0.8576 0.8659 0.8742 0.8825 0.8908 0.8991 0.9074 0.9157 0.9240 0.9324 0.9408 0.9491 0.9576 0.9660 0.9744 0.9829 0.9914 1.0000 1.0086 1.0172 1.0259 1.0346 1.0434 1.0523 1.0612 1.0702 1.0792 1.0884 1.0976 1.1070 1.1165 1.1261 1.1359 1.1458 1.1559 1.1663 1.1769 1.1877 1.1989 1.2105 1.2225 1.2352 1.2485 1.2627 1.2783 1.2958 1.3163 1.3435 1.4336
2.1390 2.1018 2.0909 2.0801 2.0694 2.0589 2.0485 2.0382 2.0281 2.0180 2.0079 1.9980 1.9880 1.9781 1.9683 1.9584 1.9485 1.9436 1.9386 1.9337 1.9287 1.9237 1.9187 1.9137 1.9086 1.9036 1.8985 1.8934 1.8882 1.8830 1.8778 1.8725 1.8672 1.8618 1.8563 1.8509 1.8453 1.8397 1.8340 1.8282 1.8223 1.8163 1.8102 1.8041 1.7977 1.7913 1.7847 1.7779 1.7710 1.7638 1.7565 1.7489 1.7411 1.7329 1.7244 1.7155 1.7062 1.6964 1.6860 1.6749 1.6629 1.6498 1.6353 1.6189 1.5999 1.5763 1.5433 1.4336
1.953 1.971 1.978 1.985 1.993 2.002 2.012 2.022 2.033 2.045 2.059 2.073 2.089 2.105 2.124 2.144 2.165 2.177 2.189 2.201 2.215 2.228 2.243 2.258 2.273 2.290 2.307 2.325 2.345 2.365 2.386 2.408 2.432 2.457 2.484 2.512 2.542 2.574 2.609 2.645 2.685 2.727 2.772 2.822 2.875 2.934 2.998 3.068 3.145 3.232 3.328 3.436 3.558 3.698 3.858 4.044 4.264 4.526 4.846 5.248 5.767 6.467 7.460 8.970 11.550 16.950 35.340
1.444 1.468 1.477 1.486 1.496 1.507 1.518 1.530 1.544 1.558 1.573 1.590 1.608 1.627 1.648 1.671 1.696 1.709 1.723 1.738 1.753 1.768 1.785 1.802 1.821 1.840 1.860 1.881 1.904 1.927 1.952 1.979 2.007 2.037 2.068 2.102 2.138 2.176 2.218 2.262 2.309 2.360 2.416 2.476 2.541 2.612 2.690 2.776 2.871 2.977 3.095 3.228 3.378 3.550 3.748 3.979 4.252 4.578 4.976 5.472 6.107 6.949 8.121 9.870 12.780 18.630 36.660
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
a Triple
0.909 0.952 0.967 0.982 0.998 1.015 1.033 1.052 1.072 1.094 1.116 1.141 1.166 1.194 1.223 1.255 1.289 1.307 1.326 1.346 1.366 1.388 1.410 1.433 1.457 1.483 1.509 1.537 1.566 1.597 1.629 1.663 1.699 1.737 1.777 1.819 1.865 1.913 1.965 2.020 2.080 2.144 2.213 2.289 2.370 2.460 2.558 2.666 2.786 2.919 3.068 3.237 3.429 3.649 3.905 4.204 4.560 4.990 5.519 6.185 7.049 8.212 9.862 12.38 16.69 25.74 55.82 point
976 928 914 900 885 871 856 842 827 813 798 783 768 753 738 723 708 700 692 684 676 668 660 651 643 635 626 617 609 600 591 582 573 564 555 546 536 527 518 508 499 489 480 470 460 451 441 431 421 411 401 391 381 370 360 349 338 326 314 302 288 274 259 243 225 205 177 0
222.8 223.4 223.5 223.6 223.6 223.6 223.5 223.4 223.2 223.1 222.8 222.5 222.2 221.8 221.4 220.9 220.4 220.1 219.8 219.5 219.2 218.8 218.5 218.1 217.7 217.4 216.9 216.5 216.1 215.6 215.2 214.7 214.2 213.7 213.1 212.6 212.0 211.5 210.9 210.3 209.6 209.0 208.3 207.6 206.9 206.2 205.4 204.6 203.8 203.0 202.1 201.2 200.3 199.3 198.3 197.2 196.1 194.9 193.6 192.3 190.8 189.1 187.2 185.0 182.1 178.2 171.3 0.0
Viscosity, Pa·s Liquid Vapor 256.7 229.3 221.6 214.3 207.2 200.3 193.8 187.4 181.3 175.4 169.7 164.2 158.9 153.8 148.8 144.0 139.3 137.1 134.8 132.6 130.4 128.3 126.2 124.1 122.0 120.0 118.0 116.1 114.1 112.2 110.3 108.4 106.6 104.8 102.9 101.2 99.4 97.6 95.9 94.2 92.5 90.8 89.1 87.5 85.8 84.2 82.6 80.9 79.3 77.7 76.1 74.4 72.8 71.2 69.5 67.8 66.1 64.4 62.7 60.9 59.0 57.0 55.0 52.8 50.3 47.5 43.8 —
10.95 11.31 11.42 11.53 11.64 11.75 11.87 11.98 12.10 12.22 12.34 12.46 12.59 12.72 12.85 12.98 13.12 13.18 13.26 13.33 13.40 13.47 13.55 13.63 13.70 13.78 13.86 13.95 14.03 14.12 14.20 14.30 14.39 14.48 14.58 14.68 14.79 14.89 15.00 15.12 15.24 15.36 15.49 15.62 15.76 15.91 16.06 16.22 16.39 16.56 16.75 16.95 17.16 17.39 17.64 17.90 18.19 18.50 18.85 19.23 19.66 20.16 20.73 21.42 22.27 23.41 25.17 —
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 180.6 172.1 169.5 166.9 164.4 161.8 159.3 156.8 154.3 151.8 149.3 146.9 144.4 141.9 139.5 137.1 134.6 133.4 132.2 131.0 129.8 128.6 127.4 126.2 125.0 123.8 122.5 121.3 120.1 118.9 117.7 116.5 115.3 114.1 112.9 111.6 110.4 109.2 108.0 106.8 105.5 104.3 103.1 101.8 100.6 99.4 98.1 96.9 95.6 94.4 93.1 91.9 90.6 89.4 88.1 86.9 85.7 84.5 83.4 82.4 81.5 80.8 80.5 80.7 81.9 85.2 95.4
11.01 11.58 11.76 11.95 12.14 12.34 12.54 12.75 12.97 13.20 13.43 13.68 13.94 14.20 14.49 14.78 15.09 15.25 15.42 15.59 15.77 15.95 16.14 16.34 16.54 16.74 16.96 17.18 17.42 17.66 17.91 18.17 18.44 18.73 19.03 19.34 19.67 20.02 20.38 20.76 21.17 21.60 22.06 22.54 23.06 23.61 24.21 24.84 25.53 26.27 27.08 27.96 28.93 29.99 31.16 32.47 33.94 35.61 37.52 39.74 42.35 45.51 49.44 54.56 61.73 73.19 98.02
17.16 15.53 15.04 14.56 14.07 13.60 13.12 12.65 12.18 11.72 11.26 10.80 10.35 9.90 9.46 9.02 8.59 8.37 8.16 7.95 7.74 7.53 7.32 7.11 6.90 6.70 6.50 6.29 6.09 5.89 5.70 5.50 5.30 5.11 4.92 4.73 4.54 4.35 4.17 3.99 3.80 3.62 3.45 3.27 3.10 2.93 2.76 2.59 2.42 2.26 2.10 1.95 1.79 1.64 1.49 1.35 1.20 1.06 0.93 0.80 0.67 0.55 0.44 0.33 0.23 0.13 0.05 0.00
–56.56 –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –19 –18 –17 –16 –15 –14 –13 –12 –11 –10 –9 –8 –7 –6 –5 –4 –3 –2 –1 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 30.98
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
PressurePressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 22 Pressure-Enthalpy Diagram for Refrigerant 50 (Methane)
30.46
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.47
Refrigerant 50 (Methane) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –182.46a –180 –175 –170 –165 –161.48b –160 –155 –150 –145 –140 –135 –130 –125 –120 –118 –116 –114 –112 –110 –108 –106 –104 –102 –100 –95 –90 –85 –82.59c
0.01170 0.01590 0.02823 0.04723 0.07509 0.10133 0.11429 0.16757 0.23784 0.32817 0.44177 0.58192 0.75201 0.95550 1.1959 1.3033 1.4174 1.5384 1.6668 1.8026 1.9462 2.0978 2.2578 2.4264 2.6040 3.0895 3.6399 4.2648 4.5992
451.5 448.2 441.4 434.5 427.4 422.4 420.2 412.7 405.0 397.1 388.8 380.2 371.1 361.6 351.4 347.2 342.8 338.3 333.6 328.8 323.7 318.4 312.9 307.1 301.0 283.6 261.7 227.5 162.7
3.9881 3.0071 1.7755 1.1081 0.72466 0.55054 0.49291 0.34665 0.25077 0.18583 0.14054 0.10813 0.08440 0.06666 0.05315 0.04865 0.04457 0.04087 0.03750 0.03442 0.03160 0.02900 0.02662 0.02441 0.02236 0.01781 0.01384 0.00993 0.00615
–71.82 –63.53 –46.58 –29.49 –12.24 0.00 5.19 22.82 40.69 58.84 77.30 96.13 115.39 135.17 155.58 163.94 172.44 181.08 189.88 198.85 208.03 217.42 227.06 236.99 247.25 274.80 306.71 349.75 415.59
472.44 477.22 486.76 495.99 504.85 510.83 513.28 521.22 528.60 535.34 541.38 546.63 550.97 554.29 556.43 556.91 557.16 557.15 556.87 556.28 555.37 554.09 552.39 550.24 547.56 537.86 521.52 488.81 415.59
Entropy, kJ/(kg·K)
Liquid Vapor Liquid Vapor –0.7099 –0.6199 –0.4429 –0.2735 –0.1108 0.0000 0.0459 0.1973 0.3440 0.4866 0.6257 0.7618 0.8956 1.0276 1.1585 1.2108 1.2631 1.3155 1.3681 1.4209 1.4741 1.5278 1.5821 1.6372 1.6934 1.8408 2.0062 2.2241 2.5624 a Triple
*Temperatures on ITS-90 scale
Specific Heat cp , kJ/(kg·K)
5.2911 3.368 2.110 5.1853 3.377 2.118 4.9910 3.399 2.137 4.8208 3.426 2.162 4.6704 3.457 2.192 4.5746 3.481 2.218 4.5363 3.492 2.229 4.4156 3.533 2.274 4.3058 3.580 2.328 4.2049 3.635 2.393 4.1111 3.701 2.472 4.0228 3.780 2.569 3.9384 3.876 2.690 3.8566 3.996 2.842 3.7759 4.148 3.038 3.7437 4.220 3.133 3.7112 4.302 3.240 3.6785 4.393 3.362 3.6454 4.497 3.501 3.6117 4.615 3.662 3.5773 4.751 3.848 3.5419 4.910 4.068 3.5054 5.097 4.331 3.4675 5.322 4.650 3.4278 5.596 5.044 3.3174 6.654 6.608 3.1791 9.10 10.37 2.9632 21.88 30.57 2.5624
cp /cv Vapor 1.341 1.343 1.348 1.355 1.365 1.373 1.377 1.392 1.412 1.436 1.466 1.504 1.552 1.613 1.694 1.733 1.777 1.827 1.885 1.951 2.028 2.119 2.227 2.358 2.520 3.153 4.64 12.16
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
1539 1516 1470 1422 1373 1338 1323 1273 1221 1168 1114 1059 1002 943 882 858 832 807 780 754 726 698 670 641 610 529 437 320 0
249.1 252.2 258.0 263.4 268.3 271.5 272.7 276.5 279.7 282.3 284.3 285.6 286.2 286.2 285.4 284.8 284.2 283.4 282.5 281.4 280.2 278.9 277.4 275.7 273.9 268.4 261.3 250.2 0.0 b Normal
point
204.5 189.5 163.9 143.4 126.8 116.8 112.9 101.3 91.4 82.8 75.2 68.6 62.7 57.3 52.4 50.5 48.7 47.0 45.3 43.6 42.0 40.4 38.8 37.2 35.6 31.7 27.5 22.3 —
3.64 3.73 3.93 4.12 4.32 4.46 4.52 4.73 4.95 5.17 5.40 5.63 5.88 6.15 6.43 6.55 6.68 6.81 6.95 7.09 7.24 7.40 7.57 7.76 7.96 8.55 9.39 10.95 —
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 211.2 208.3 202.1 195.5 188.7 183.9 181.8 174.8 167.7 160.5 153.4 146.3 139.2 132.1 125.0 122.1 119.3 116.4 113.5 110.7 107.8 104.9 102.0 99.0 96.1 88.6 81.2 77.5
8.85 9.15 9.78 10.42 11.09 11.58 11.79 12.53 13.30 14.13 15.01 15.96 16.99 18.13 19.39 19.95 20.53 21.15 21.81 22.52 23.28 24.12 25.05 26.09 27.28 31.27 38.68 62.99
18.76 18.09 16.74 15.43 14.17 13.30 12.94 11.75 10.60 9.49 8.43 7.42 6.44 5.52 4.64 4.31 3.98 3.66 3.35 3.04 2.75 2.46 2.19 1.92 1.66 1.07 0.55 0.13 0.00
–182.46 –180 –175 –170 –165 –161.48 –160 –155 –150 –145 –140 –135 –130 –125 –120 –118 –116 –114 –112 –110 –108 –106 –104 –102 –100 –95 –90 –85 –82.59
c Critical
boiling point
point
Refrigerant 50 (Methane) Properties of Gas at 0.101 325 MPa (one standard atmosphere) Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg·K) (kg·K)
cp /cv
–161.5a –160 –155 –150 –145 –140 –135 –130 –120 –110 –100 –90 –80 –70 –60 –50 –40 –30 –20 –10 0 10 20 30 40
1.373 1.371 1.366 1.362 1.359 1.356 1.354 1.351 1.348 1.345 1.343 1.341 1.339 1.337 1.335 1.332 1.330 1.327 1.324 1.320 1.316 1.312 1.308 1.304 1.299
1.8164 1.7899 1.7064 1.6310 1.5626 1.5000 1.4425 1.3894 1.2947 1.2124 1.1403 1.0764 1.0194 0.9683 0.9221 0.8802 0.8419 0.8069 0.7747 0.7450 0.7175 0.6919 0.6682 0.6460 0.6252
a Saturated
510.83 514.11 525.08 535.95 546.74 557.47 568.15 578.79 599.99 621.11 642.18 663.23 684.27 705.33 726.42 747.57 768.78 790.10 811.53 833.09 854.82 876.72 898.82 921.14 943.69
4.5746 4.6037 4.6986 4.7887 4.8746 4.9567 5.0355 5.1112 5.2543 5.3879 5.5133 5.6315 5.7433 5.8496 5.9509 6.0479 6.1409 6.2304 6.3168 6.4003 6.4814 6.5601 6.6368 6.7117 6.7849
vapor at normal boiling point
2.218 2.208 2.183 2.165 2.151 2.141 2.132 2.125 2.116 2.109 2.106 2.104 2.105 2.107 2.111 2.118 2.126 2.137 2.149 2.164 2.181 2.200 2.221 2.243 2.268
Thermal Vel. of Cond., Sound, Visc., m/s Pa·s mW/(m·K) 271.5 273.5 280.3 286.9 293.2 299.4 305.4 311.2 322.5 333.3 343.8 353.8 363.5 372.9 382.0 390.7 399.2 407.5 415.4 423.2 430.7 437.9 445.0 451.9 458.6
4.46 4.52 4.71 4.90 5.10 5.29 5.48 5.67 6.06 6.44 6.81 7.19 7.56 7.92 8.29 8.64 9.00 9.35 9.69 10.03 10.37 10.70 11.02 11.35 11.66
11.58 11.74 12.30 12.87 13.45 14.03 14.62 15.21 16.40 17.58 18.77 19.96 21.15 22.31 23.49 24.68 25.89 27.12 28.36 29.63 30.93 32.25 33.61 34.99 36.41
Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg·K) (kg·K) 50 60 70 80 90 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300
0.6058 0.5875 0.5703 0.5541 0.5388 0.5243 0.5106 0.4975 0.4852 0.4734 0.4622 0.4515 0.4413 0.4315 0.4222 0.4132 0.4047 0.3965 0.3886 0.3810 0.3737 0.3667 0.3599 0.3534 0.3471 0.3411
966.50 989.57 1012.92 1036.57 1060.52 1084.79 1109.39 1134.32 1159.60 1185.22 1211.21 1237.55 1264.25 1291.32 1318.76 1346.57 1374.76 1403.32 1432.26 1461.57 1491.26 1521.32 1551.76 1582.57 1613.75 1645.31
6.8565 6.9269 6.9959 7.0638 7.1307 7.1967 7.2617 7.3259 7.3894 7.4522 7.5143 7.5759 7.6368 7.6972 7.7571 7.8165 7.8755 7.9340 7.9921 8.0498 8.1071 8.1640 8.2205 8.2768 8.3326 8.3882
2.293 2.321 2.350 2.380 2.411 2.443 2.476 2.510 2.545 2.580 2.616 2.652 2.689 2.726 2.763 2.800 2.837 2.875 2.912 2.950 2.988 3.025 3.062 3.100 3.137 3.174
cp /cv
Vel. of Sound, m/s
1.295 1.290 1.285 1.281 1.276 1.271 1.266 1.262 1.257 1.253 1.248 1.244 1.240 1.236 1.232 1.228 1.224 1.221 1.217 1.214 1.211 1.207 1.204 1.201 1.198 1.196
465.1 471.4 477.6 483.7 489.6 495.5 501.2 506.8 512.3 517.7 523.1 528.3 533.5 538.6 543.7 548.7 553.6 558.5 563.4 568.2 572.9 577.6 582.3 586.9 591.5 596.0
Thermal Cond., Visc., Pa·s mW/(m·K) 11.98 12.29 12.60 12.90 13.20 13.49 13.79 14.08 14.36 14.64 14.92 15.20 15.47 15.74 16.01 16.28 16.54 16.80 17.06 17.31 17.56 17.81 18.06 18.31 18.55 18.79
37.86 39.35 40.87 42.42 44.01 45.63 47.28 48.96 50.67 52.41 54.18 55.97 57.79 59.63 61.50 63.39 65.30 67.23 69.18 71.15 73.14 75.14 77.16 79.20 81.25 83.31
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 23 Pressure-Enthalpy Diagram for Refrigerant 170 (Ethane)
30.48
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.49
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 170 (Ethane) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg Temp.,* sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor °C
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
–175 0.00001 –170 0.00002 –165 0.00005 –160 0.00013 –155 0.00027 –150 0.00054 –145 0.00103 –140 0.00185 –135 0.00316 –130 0.00519 –125 0.00821 –120 0.01255 –115 0.01863 –110 0.02691 –105 0.03793 –100 0.05230 –95 0.07068 –90 0.09380 –88.58b 0.10132 –85 0.12243 –80 0.15741 –78 0.17338 –76 0.19055 –74 0.20899 –72 0.22877 –70 0.24993 –68 0.27255 –66 0.29668 –64 0.32238 –62 0.34972 –60 0.37877 –58 0.40958 –56 0.44223 –54 0.47677 –52 0.51328 –50 0.55183 –48 0.59247 –46 0.63529 –44 0.68034 –42 0.72770 –40 0.77744 –38 0.82963 –36 0.88433 –34 0.94163 –32 1.0016 –30 1.0643 –28 1.1298 –26 1.1982 –24 1.2696 –22 1.3440 –20 1.4215 –18 1.5023 –16 1.5863 –14 1.6737 –12 1.7645 –10 1.8588 –8 1.9568 –6 2.0585 –4 2.1640 –2 2.2734 0 2.3867 5 2.6883 10 3.0172 15 3.3755 20 3.7655 25 4.1903 30 4.6551 32.17c 4.8722
–1.4643 –1.3510 –1.2433 –1.1405 –1.0421 –0.9476 –0.8566 –0.7688 –0.6839 –0.6017 –0.5219 –0.4445 –0.3691 –0.2957 –0.2241 –0.1542 –0.0858 –0.0188 0.0000 0.0469 0.1115 0.1370 0.1623 0.1875 0.2125 0.2373 0.2621 0.2867 0.3111 0.3355 0.3597 0.3838 0.4078 0.4317 0.4556 0.4793 0.5029 0.5265 0.5500 0.5735 0.5969 0.6202 0.6435 0.6668 0.6900 0.7133 0.7365 0.7597 0.7829 0.8062 0.8295 0.8528 0.8761 0.8996 0.9231 0.9467 0.9704 0.9942 1.0182 1.0423 1.0667 1.1287 1.1928 1.2602 1.3328 1.4152 1.5263 1.6901
2.288 1.184 2.277 1.194 2.273 1.204 2.274 1.215 2.278 1.226 2.283 1.238 2.290 1.251 2.299 1.265 2.308 1.279 2.318 1.293 2.328 1.307 2.340 1.320 2.353 1.333 2.366 1.347 2.381 1.362 2.397 1.380 2.415 1.401 2.434 1.425 2.439 1.433 2.455 1.454 2.478 1.487 2.487 1.501 2.497 1.515 2.508 1.531 2.519 1.546 2.530 1.562 2.542 1.579 2.554 1.596 2.567 1.614 2.580 1.633 2.594 1.651 2.608 1.671 2.623 1.691 2.638 1.711 2.654 1.732 2.671 1.754 2.689 1.777 2.707 1.800 2.726 1.824 2.746 1.850 2.767 1.876 2.789 1.904 2.812 1.932 2.836 1.963 2.862 1.995 2.889 2.028 2.918 2.064 2.948 2.102 2.980 2.142 3.014 2.185 3.050 2.231 3.089 2.281 3.130 2.334 3.175 2.392 3.223 2.455 3.275 2.523 3.331 2.598 3.393 2.681 3.460 2.773 3.534 2.875 3.617 2.990 3.871 3.352 4.227 3.878 4.773 4.713 5.746 6.249 8.05 10.02 21.69 32.79
1.305 1.302 1.298 1.295 1.292 1.289 1.287 1.284 1.282 1.280 1.278 1.277 1.277 1.277 1.278 1.280 1.282 1.286 1.287 1.290 1.295 1.297 1.300 1.302 1.305 1.308 1.312 1.315 1.319 1.323 1.327 1.331 1.336 1.342 1.347 1.353 1.359 1.366 1.373 1.381 1.390 1.399 1.408 1.419 1.430 1.442 1.455 1.469 1.485 1.501 1.519 1.539 1.561 1.584 1.610 1.639 1.671 1.706 1.746 1.791 1.841 2.001 2.236 2.611 3.299 4.96 14.35
643.0 637.5 632.0 626.5 621.0 615.5 609.9 604.3 598.7 593.0 587.3 581.5 575.7 569.8 563.9 557.9 551.8 545.6 543.8 539.3 532.9 530.3 527.7 525.1 522.5 519.8 517.1 514.4 511.6 508.9 506.1 503.2 500.4 497.5 494.6 491.6 488.7 485.7 482.6 479.5 476.4 473.2 470.0 466.7 463.4 460.0 456.6 453.1 449.6 446.0 442.3 438.5 434.7 430.8 426.8 422.6 418.4 414.1 409.7 405.1 400.3 387.7 373.7 357.9 339.0 315.0 276.3 206.2
3649.3 1353.1 555.95 249.56 120.95 62.674 34.435 19.920 12.060 7.6020 4.9668 3.3504 2.3254 1.6556 1.2060 0.89662 0.67900 0.52278 0.48676 0.40856 0.32361 0.29583 0.27094 0.24859 0.22847 0.21033 0.19394 0.17909 0.16563 0.15339 0.14224 0.13208 0.12279 0.11429 0.10650 0.09935 0.09277 0.08671 0.08113 0.07597 0.07120 0.06679 0.06269 0.05889 0.05536 0.05207 0.04900 0.04614 0.04347 0.04097 0.03863 0.03644 0.03439 0.03245 0.03064 0.02892 0.02731 0.02579 0.02435 0.02298 0.02169 0.01874 0.01612 0.01377 0.01160 0.00952 0.00723 0.00485
*Temperatures on ITS-90 scale
–201.26 –189.85 –178.48 –167.11 –155.73 –144.33 –132.89 –121.42 –109.90 –98.34 –86.72 –75.04 –63.31 –51.50 –39.62 –27.66 –15.61 –3.46 0.00 8.79 21.15 26.13 31.13 36.16 41.20 46.27 51.36 56.48 61.62 66.79 71.99 77.21 82.47 87.75 93.07 98.42 103.80 109.22 114.68 120.18 125.71 131.29 136.91 142.57 148.28 154.04 159.86 165.72 171.64 177.62 183.67 189.78 195.95 202.20 208.53 214.94 221.43 228.02 234.71 241.50 248.41 266.27 285.14 305.36 327.57 353.26 388.24 438.99
384.73 390.67 396.66 402.69 408.76 414.87 421.00 427.16 433.34 439.53 445.71 451.89 458.04 464.15 470.20 476.16 482.04 487.79 489.40 493.42 498.91 501.07 503.20 505.30 507.38 509.43 511.45 513.44 515.40 517.34 519.23 521.10 522.93 524.73 526.48 528.20 529.88 531.52 533.11 534.66 536.16 537.60 539.00 540.34 541.62 542.85 544.00 545.10 546.12 547.06 547.93 548.72 549.41 550.02 550.52 550.91 551.19 551.35 551.37 551.24 550.96 549.46 546.55 541.71 533.99 521.20 494.95 438.99
4.5060 4.2770 4.0747 3.8953 3.7356 3.5932 3.4657 3.3513 3.2484 3.1557 3.0719 2.9962 2.9274 2.8649 2.8078 2.7556 2.7076 2.6635 2.6516 2.6227 2.5850 2.5707 2.5567 2.5432 2.5300 2.5172 2.5048 2.4926 2.4808 2.4692 2.4580 2.4470 2.4362 2.4257 2.4154 2.4053 2.3954 2.3856 2.3760 2.3666 2.3573 2.3481 2.3390 2.3301 2.3211 2.3123 2.3035 2.2947 2.2859 2.2772 2.2684 2.2596 2.2507 2.2417 2.2326 2.2234 2.2140 2.2045 2.1947 2.1847 2.1743 2.1468 2.1161 2.0804 2.0369 1.9785 1.8783 1.6901
bNormal
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
1952 1916 1880 1844 1808 1772 1735 1699 1663 1626 1589 1552 1515 1478 1440 1403 1365 1326 1316 1288 1249 1234 1218 1203 1187 1171 1156 1140 1124 1108 1092 1076 1060 1044 1028 1012 996 979 963 947 930 914 897 880 864 847 830 813 796 778 761 744 726 708 690 672 654 635 617 598 579 531 480 427 368 302 220 0
boiling point
188.2 192.7 197.0 201.3 205.4 209.4 213.3 217.0 220.7 224.2 227.6 230.9 234.0 237.0 239.7 242.3 244.7 246.9 247.4 248.8 250.4 251.0 251.6 252.1 252.5 252.9 253.3 253.6 253.9 254.2 254.3 254.5 254.6 254.6 254.6 254.6 254.5 254.4 254.2 253.9 253.6 253.2 252.8 252.4 251.8 251.2 250.6 249.9 249.1 248.3 247.4 246.4 245.4 244.3 243.1 241.9 240.6 239.2 237.7 236.1 234.5 229.9 224.8 218.9 212.1 204.1 192.0 0.0
933.9 783.6 669.0 579.6 508.5 450.9 403.6 364.1 330.8 302.4 277.8 256.4 237.6 220.9 206.0 192.5 180.4 169.3 166.4 159.2 149.9 146.4 143.0 139.6 136.4 133.3 130.3 127.3 124.4 121.6 118.9 116.2 113.6 111.1 108.6 106.2 103.9 101.5 99.3 97.1 94.9 92.8 90.7 88.7 86.7 84.7 82.7 80.8 79.0 77.1 75.3 73.5 71.8 70.0 68.3 66.6 64.9 63.3 61.6 60.0 58.4 54.3 50.3 46.2 41.8 37.0 30.6 —
3.26 3.41 3.55 3.70 3.84 3.99 4.14 4.29 4.44 4.59 4.74 4.90 5.05 5.21 5.36 5.52 5.68 5.84 5.89 6.00 6.16 6.23 6.30 6.36 6.43 6.50 6.56 6.63 6.70 6.77 6.84 6.91 6.99 7.06 7.13 7.21 7.28 7.36 7.44 7.51 7.59 7.68 7.76 7.84 7.93 8.02 8.11 8.20 8.30 8.40 8.50 8.61 8.71 8.83 8.94 9.07 9.19 9.33 9.46 9.61 9.77 10.20 10.71 11.34 12.17 13.38 15.75 —
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 249.4 245.2 240.8 236.3 231.6 226.9 222.1 217.3 212.4 207.5 202.5 197.6 192.7 187.7 182.8 178.0 173.1 168.3 167.0 163.6 158.9 157.0 155.2 153.3 151.5 149.7 147.9 146.0 144.2 142.4 140.6 138.9 137.1 135.3 133.6 131.8 130.1 128.4 126.7 125.0 123.3 121.6 119.9 118.2 116.6 114.9 113.3 111.6 110.0 108.4 106.7 105.1 103.5 101.9 100.3 98.7 97.1 95.6 94.0 92.4 90.8 86.9 83.0 79.0 75.2 72.0 74.9
3.35 3.64 3.93 4.23 4.54 4.86 5.18 5.50 5.84 6.18 6.54 6.90 7.28 7.67 8.07 8.48 8.92 9.36 9.49 9.83 10.31 10.51 10.72 10.92 11.13 11.35 11.56 11.79 12.01 12.24 12.48 12.72 12.96 13.21 13.46 13.73 13.99 14.26 14.54 14.83 15.12 15.42 15.73 16.04 16.37 16.70 17.05 17.40 17.77 18.15 18.55 18.96 19.38 19.83 20.29 20.78 21.30 21.84 22.41 23.02 23.68 25.53 27.85 30.90 35.30 42.79 64.07
30.35 29.51 28.67 27.83 26.99 26.15 25.31 24.48 23.64 22.81 21.98 21.15 20.33 19.51 18.69 17.88 17.07 16.26 16.03 15.46 14.67 14.35 14.04 13.72 13.41 13.09 12.78 12.47 12.16 11.86 11.55 11.24 10.94 10.63 10.33 10.03 9.73 9.44 9.14 8.85 8.55 8.26 7.97 7.68 7.40 7.11 6.83 6.55 6.27 6.00 5.72 5.45 5.18 4.92 4.65 4.39 4.13 3.88 3.63 3.38 3.13 2.54 1.97 1.43 0.93 0.48 0.11 0.00
–175 –170 –165 –160 –155 –150 –145 –140 –135 –130 –125 –120 –115 –110 –105 –100 –95 –90 –88.58 –85 –80 –78 –76 –74 –72 –70 –68 –66 –64 –62 –60 –58 –56 –54 –52 –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 5 10 15 20 25 30 32.17
cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 24 Pressure-Enthalpy Diagram for Refrigerant 290 (Propane)
30.50
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.51
Refrigerant 290 (Propane) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp., °C
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor
–150 0.00001 –140 0.00003 –130 0.00012 –120 0.00041 –110 0.00116 –100 0.00290 –90 0.00645 –80 0.01305 –70 0.02440 –60 0.04269 –50 0.07057 –42.11b 0.10133 –40 0.11112 –38 0.12105 –36 0.13166 –34 0.14297 –32 0.15502 –30 0.16783 –28 0.18144 –26 0.19589 –24 0.21119 –22 0.22739 –20 0.24452 –18 0.26261 –16 0.28170 –14 0.30181 –12 0.32300 –10 0.34528 –8 0.36870 –6 0.39329 –4 0.41909 –2 0.44613 0 0.47446 2 0.50410 4 0.53510 6 0.56749 8 0.60131 10 0.63660 12 0.67340 14 0.71175 16 0.75168 18 0.79324 20 0.83646 22 0.88139 24 0.92807 26 0.97653 28 1.0268 30 1.0790 32 1.1331 34 1.1891 36 1.2472 38 1.3072 40 1.3694 42 1.4337 44 1.5002 46 1.5690 48 1.6400 50 1.7133 55 1.9072 60 2.1168 65 2.3430 70 2.5868 75 2.8493 80 3.1319 85 3.4361 90 3.7641 95 4.1195 96.74c 4.2512 bNormal
694.6 684.5 674.4 664.3 654.0 643.7 633.3 622.8 612.0 601.1 589.9 580.9 578.4 576.1 573.8 571.4 569.0 566.6 564.2 561.8 559.4 556.9 554.5 552.0 549.5 546.9 544.4 541.8 539.2 536.6 533.9 531.3 528.6 525.9 523.1 520.4 517.6 514.7 511.9 509.0 506.0 503.1 500.1 497.0 493.9 490.8 487.6 484.4 481.1 477.8 474.4 471.0 467.5 463.9 460.3 456.5 452.7 448.9 438.8 428.0 416.3 403.6 389.5 373.3 354.0 328.8 286.5 220.5
boiling point
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
1.962 1.977 1.994 2.012 2.032 2.054 2.078 2.106 2.137 2.172 2.212 2.246 2.256 2.265 2.275 2.285 2.295 2.305 2.316 2.327 2.338 2.349 2.361 2.373 2.385 2.397 2.410 2.423 2.436 2.450 2.464 2.478 2.493 2.508 2.524 2.540 2.556 2.573 2.591 2.609 2.627 2.646 2.666 2.687 2.708 2.730 2.753 2.777 2.802 2.827 2.855 2.883 2.913 2.944 2.977 3.012 3.050 3.089 3.201 3.337 3.509 3.735 4.053 4.545 5.433 7.623 23.59
1.227 1.218 1.210 1.203 1.197 1.192 1.188 1.184 1.182 1.181 1.182 1.183 1.184 1.185 1.185 1.186 1.187 1.188 1.189 1.191 1.192 1.193 1.195 1.197 1.198 1.200 1.202 1.205 1.207 1.209 1.212 1.215 1.218 1.221 1.225 1.229 1.232 1.237 1.241 1.246 1.251 1.256 1.262 1.268 1.275 1.282 1.290 1.298 1.307 1.316 1.326 1.337 1.349 1.362 1.375 1.391 1.407 1.425 1.478 1.548 1.641 1.773 1.970 2.288 2.883 4.374 14.62
4316.4 –123.78 402.06 –0.6903 3.5796 864.49 –104.09 412.43 –0.5366 3.3426 223.53 –84.23 423.12 –0.3929 3.1514 70.785 –64.21 434.11 –0.2576 2.9962 26.386 –43.99 445.38 –0.1298 2.8697 11.231 –23.56 456.88 –0.0083 2.7664 5.3300 –2.90 468.58 0.1077 2.6820 2.7676 18.03 480.44 0.2189 2.6130 1.5487 39.25 492.41 0.3259 2.5566 0.92250 60.81 504.44 0.4294 2.5107 0.57905 82.75 516.48 0.5298 2.4734 0.41388 100.36 525.95 0.6070 2.4491 0.37985 105.12 528.48 0.6275 2.4433 0.35076 109.65 530.87 0.6468 2.4380 0.32437 114.20 533.26 0.6660 2.4330 0.30037 118.77 535.64 0.6851 2.4282 0.27853 123.36 538.01 0.7041 2.4236 0.25861 127.97 540.38 0.7231 2.4192 0.24041 132.61 542.75 0.7419 2.4150 0.22376 137.26 545.11 0.7607 2.4109 0.20851 141.94 547.46 0.7795 2.4071 0.19452 146.64 549.80 0.7982 2.4034 0.18167 151.36 552.13 0.8168 2.3999 0.16984 156.11 554.46 0.8353 2.3965 0.15894 160.88 556.77 0.8538 2.3933 0.14889 165.68 559.08 0.8722 2.3903 0.13961 170.50 561.37 0.8906 2.3874 0.13103 175.35 563.65 0.9090 2.3846 0.12308 180.22 565.92 0.9273 2.3819 0.11571 185.12 568.18 0.9455 2.3794 0.10887 190.05 570.42 0.9637 2.3769 0.10252 195.01 572.65 0.9819 2.3746 0.09661 200.00 574.87 1.0000 2.3724 0.09111 205.02 577.06 1.0181 2.3703 0.08598 210.06 579.24 1.0362 2.3682 0.08120 215.14 581.41 1.0542 2.3663 0.07673 220.25 583.55 1.0722 2.3644 0.07255 225.40 585.67 1.0902 2.3626 0.06865 230.57 587.77 1.1082 2.3608 0.06498 235.79 589.85 1.1261 2.3592 0.06155 241.03 591.91 1.1440 2.3575 0.05833 246.32 593.94 1.1620 2.3560 0.05530 251.64 595.95 1.1799 2.3544 0.05246 256.99 597.93 1.1978 2.3529 0.04978 262.39 599.88 1.2157 2.3514 0.04726 267.83 601.80 1.2336 2.3500 0.04488 273.31 603.68 1.2515 2.3486 0.04264 278.83 605.54 1.2695 2.3471 0.04053 284.40 607.35 1.2874 2.3457 0.03853 290.01 609.13 1.3053 2.3443 0.03664 295.68 610.87 1.3233 2.3429 0.03485 301.39 612.57 1.3413 2.3414 0.03315 307.15 614.21 1.3594 2.3399 0.03154 312.96 615.81 1.3774 2.3384 0.03002 318.83 617.36 1.3955 2.3368 0.02857 324.76 618.86 1.4137 2.3352 0.02720 330.75 620.29 1.4319 2.3335 0.02589 336.80 621.66 1.4502 2.3317 0.02288 352.23 624.77 1.4962 2.3268 0.02020 368.14 627.36 1.5429 2.3210 0.01781 384.60 629.29 1.5903 2.3139 0.01565 401.75 630.37 1.6389 2.3052 0.01367 419.76 630.33 1.6891 2.2939 0.01185 438.93 628.73 1.7417 2.2791 0.01012 459.81 624.75 1.7980 2.2586 0.00840 483.71 616.47 1.8616 2.2272 0.00640 516.33 595.81 1.9476 2.1635 0.00454 555.24 555.24 2.0516 2.0516
1.020 1.054 1.087 1.119 1.151 1.184 1.220 1.258 1.300 1.346 1.397 1.440 1.453 1.464 1.476 1.488 1.501 1.513 1.526 1.539 1.553 1.566 1.580 1.595 1.609 1.624 1.639 1.655 1.671 1.687 1.704 1.721 1.739 1.757 1.776 1.795 1.815 1.835 1.856 1.878 1.901 1.925 1.949 1.975 2.001 2.029 2.058 2.088 2.119 2.152 2.187 2.224 2.263 2.304 2.348 2.395 2.445 2.499 2.652 2.841 3.086 3.421 3.914 4.707 6.182 9.888 36.07
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
1880 1813 1745 1679 1612 1545 1478 1411 1345 1278 1213 1161 1147 1134 1121 1108 1095 1082 1069 1056 1043 1030 1016 1003 990 977 964 951 938 925 912 899 885 872 859 846 833 819 806 793 780 766 753 739 726 713 699 685 672 658 645 631 617 603 589 575 561 547 511 474 437 398 358 315 269 218 158 0
168.8 174.9 180.7 186.3 191.7 196.8 201.5 205.9 209.9 213.5 216.5 218.4 218.9 219.3 219.6 220.0 220.3 220.6 220.9 221.1 221.3 221.5 221.6 221.8 221.8 221.9 221.9 221.9 221.8 221.8 221.6 221.5 221.3 221.1 220.8 220.5 220.2 219.8 219.3 218.9 218.4 217.8 217.2 216.6 215.9 215.2 214.4 213.5 212.6 211.7 210.7 209.7 208.6 207.4 206.2 204.9 203.6 202.2 198.3 194.1 189.3 184.0 178.2 171.6 164.1 155.5 144.1 0.0
1343.0 985.4 761.7 611.6 505.0 425.7 364.5 315.9 276.4 243.6 216.0 197.2 192.6 188.3 184.1 180.1 176.1 172.3 168.6 165.0 161.5 158.1 154.7 151.5 148.3 145.2 142.2 139.3 136.4 133.6 130.9 128.2 125.6 123.0 120.5 118.1 115.7 113.3 111.0 108.8 106.6 104.4 102.3 100.2 98.1 96.1 94.1 92.2 90.3 88.4 86.5 84.7 82.8 81.0 79.3 77.5 75.8 74.1 69.8 65.7 61.5 57.4 53.2 48.8 44.1 38.8 31.4 —
3.55 3.80 4.05 4.31 4.56 4.82 5.08 5.34 5.60 5.85 6.11 6.31 6.36 6.41 6.47 6.52 6.57 6.62 6.67 6.73 6.78 6.83 6.89 6.94 6.99 7.05 7.10 7.16 7.22 7.27 7.33 7.39 7.45 7.51 7.57 7.63 7.69 7.75 7.82 7.88 7.95 8.02 8.09 8.16 8.23 8.31 8.38 8.46 8.54 8.63 8.71 8.80 8.89 8.99 9.08 9.19 9.29 9.40 9.70 10.03 10.42 10.86 11.40 12.07 12.96 14.28 17.00 —
Thermal Cond., Surface mW/(m·K) Tension, Temp., Liquid Vapor mN/m °C 192.9 187.7 182.2 176.4 170.4 164.4 158.2 152.1 145.9 139.8 133.8 129.2 128.0 126.8 125.6 124.5 123.3 122.2 121.1 120.0 118.8 117.7 116.6 115.5 114.4 113.4 112.3 111.2 110.1 109.1 108.0 107.0 106.0 104.9 103.9 102.9 101.9 100.9 99.9 99.0 98.0 97.0 96.1 95.1 94.2 93.3 92.3 91.4 90.5 89.6 88.7 87.8 86.9 86.0 85.2 84.3 83.5 82.6 80.5 78.4 76.3 74.3 72.2 70.2 68.3 67.1 73.6
3.68 4.28 4.90 5.55 6.23 6.94 7.67 8.43 9.22 10.04 10.88 11.57 11.76 11.94 12.12 12.30 12.48 12.67 12.86 13.05 13.24 13.43 13.63 13.83 14.03 14.23 14.44 14.65 14.86 15.08 15.29 15.52 15.74 15.97 16.20 16.44 16.68 16.93 17.18 17.44 17.70 17.97 18.24 18.53 18.81 19.11 19.41 19.72 20.05 20.38 20.72 21.07 21.43 21.81 22.20 22.60 23.03 23.47 24.65 26.00 27.56 29.41 31.71 34.75 39.13 46.66 69.48
31.84 30.29 28.76 27.24 25.73 24.23 22.74 21.27 19.81 18.37 16.94 15.83 15.54 15.26 14.98 14.70 14.42 14.15 13.87 13.60 13.33 13.06 12.79 12.52 12.25 11.98 11.71 11.45 11.18 10.92 10.65 10.39 10.13 9.87 9.62 9.36 9.10 8.85 8.60 8.34 8.09 7.85 7.60 7.35 7.11 6.87 6.62 6.38 6.15 5.91 5.68 5.44 5.21 4.98 4.76 4.53 4.31 4.09 3.55 3.02 2.52 2.03 1.56 1.12 0.72 0.36 0.06 0.00
–150 –140 –130 –120 –110 –100 –90 –80 –70 –60 –50 –42.11 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 55 60 65 70 75 80 85 90 95 96.74
cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 25 Pressure-Enthalpy Diagram for Refrigerant 600 (n-Butane)
Licensed for single user. © 2017 ASHRAE, Inc.
30.52
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.53
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 600 (n-Butane) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg Temp.,* sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor °C
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
–100 0.00016 –95 0.00028 –90 0.00045 –85 0.00072 –80 0.00111 –75 0.00168 –70 0.00247 –65 0.00355 –60 0.00501 –55 0.00695 –50 0.00947 –45 0.01270 –40 0.01679 –35 0.02190 –30 0.02821 –25 0.03591 –20 0.04521 –15 0.05635 –10 0.06955 –5 0.08509 –0.49b 0.10132 0 0.10323 2 0.11127 4 0.11980 6 0.12882 8 0.13837 10 0.14845 12 0.15909 14 0.17031 16 0.18213 18 0.19457 20 0.20765 22 0.22139 24 0.23582 26 0.25095 28 0.26680 30 0.28341 32 0.30079 34 0.31897 36 0.33796 38 0.35779 40 0.37849 42 0.40007 44 0.42256 46 0.44599 48 0.47038 50 0.49575 55 0.56365 60 0.63824 65 0.71991 70 0.80908 75 0.90616 80 1.0116 85 1.1258 90 1.2493 95 1.3825 100 1.5259 105 1.6801 110 1.8456 115 2.0230 120 2.2131 125 2.4166 130 2.6344 135 2.8675 140 3.1172 145 3.3853 150 3.6746 151.98c 3.7960
0.0318 0.0892 0.1453 0.2000 0.2536 0.3061 0.3575 0.4080 0.4575 0.5062 0.5541 0.6013 0.6478 0.6937 0.7389 0.7836 0.8278 0.8715 0.9147 0.9576 0.9959 1.0000 1.0169 1.0337 1.0505 1.0672 1.0838 1.1005 1.1170 1.1335 1.1500 1.1665 1.1829 1.1992 1.2155 1.2318 1.2481 1.2643 1.2805 1.2966 1.3127 1.3288 1.3449 1.3609 1.3769 1.3929 1.4089 1.4488 1.4885 1.5282 1.5679 1.6075 1.6471 1.6868 1.7266 1.7665 1.8066 1.8469 1.8876 1.9287 1.9704 2.0129 2.0566 2.1020 2.1502 2.2041 2.2755 2.3631
2.013 1.231 2.021 1.247 2.029 1.263 2.038 1.279 2.048 1.295 2.058 1.312 2.069 1.330 2.081 1.347 2.094 1.366 2.108 1.385 2.122 1.404 2.137 1.425 2.153 1.446 2.170 1.468 2.188 1.490 2.206 1.514 2.226 1.538 2.246 1.563 2.267 1.589 2.289 1.616 2.310 1.641 2.312 1.644 2.321 1.655 2.331 1.667 2.341 1.678 2.350 1.690 2.360 1.702 2.371 1.715 2.381 1.727 2.391 1.740 2.402 1.752 2.413 1.765 2.424 1.778 2.435 1.792 2.446 1.805 2.458 1.819 2.470 1.833 2.481 1.847 2.494 1.862 2.506 1.876 2.518 1.891 2.531 1.906 2.544 1.922 2.557 1.937 2.571 1.953 2.585 1.970 2.598 1.986 2.635 2.029 2.673 2.075 2.713 2.123 2.756 2.174 2.802 2.229 2.851 2.288 2.905 2.353 2.964 2.425 3.029 2.506 3.102 2.599 3.186 2.708 3.285 2.841 3.403 3.004 3.552 3.213 3.748 3.493 4.023 3.891 4.450 4.511 5.227 5.631 7.147 8.349 19.80 25.55
1.132 1.130 1.128 1.127 1.125 1.124 1.122 1.121 1.120 1.119 1.118 1.118 1.117 1.117 1.116 1.116 1.116 1.117 1.117 1.118 1.119 1.119 1.120 1.120 1.121 1.122 1.122 1.123 1.124 1.125 1.126 1.127 1.128 1.129 1.130 1.131 1.133 1.134 1.136 1.137 1.139 1.141 1.143 1.145 1.147 1.149 1.151 1.158 1.165 1.173 1.183 1.194 1.207 1.223 1.241 1.263 1.290 1.324 1.366 1.420 1.492 1.592 1.739 1.973 2.405 3.462 10.14
699.3 694.6 689.9 685.2 680.5 675.8 671.0 666.2 661.4 656.6 651.7 646.8 641.9 636.9 631.9 626.8 621.7 616.6 611.4 606.1 601.3 600.7 598.6 596.4 594.2 592.0 589.8 587.6 585.4 583.1 580.9 578.6 576.3 574.0 571.7 569.3 567.0 564.6 562.2 559.8 557.4 554.9 552.4 550.0 547.4 544.9 542.3 535.8 529.1 522.3 515.2 507.9 500.4 492.6 484.5 476.0 467.1 457.8 447.9 437.3 425.9 413.4 399.6 383.7 364.7 339.9 297.7 228.0
150.44 91.606 57.588 37.271 24.773 16.873 11.752 8.3565 6.0558 4.4659 3.3470 2.5462 1.9638 1.5341 1.2127 0.96911 0.78237 0.63759 0.52415 0.43441 0.36910 0.36275 0.33818 0.31562 0.29488 0.27578 0.25817 0.24192 0.22691 0.21302 0.20016 0.18823 0.17717 0.16688 0.15732 0.14842 0.14012 0.13238 0.12516 0.11841 0.11209 0.10618 0.10065 0.09545 0.09058 0.08600 0.08170 0.07201 0.06366 0.05642 0.05012 0.04462 0.03978 0.03552 0.03175 0.02840 0.02541 0.02273 0.02032 0.01813 0.01615 0.01432 0.01264 0.01105 0.00953 0.00801 0.00621 0.00439
*Temperatures on ITS-90 scale
–13.65 –3.57 6.56 16.73 26.94 37.21 47.53 57.90 68.34 78.85 89.42 100.07 110.80 121.62 132.52 143.51 154.60 165.79 177.08 188.48 198.87 200.00 204.64 209.30 213.98 218.68 223.40 228.13 232.89 237.68 242.48 247.30 252.15 257.02 261.91 266.82 271.76 276.72 281.71 286.72 291.76 296.82 301.90 307.02 312.15 317.32 322.51 335.62 348.91 362.39 376.06 389.95 404.06 418.40 433.00 447.87 463.03 478.51 494.36 510.61 527.34 544.65 562.68 581.69 602.17 625.32 656.27 693.91
450.85 457.02 463.27 469.58 475.97 482.41 488.92 495.50 502.13 508.82 515.56 522.35 529.19 536.08 543.01 549.98 556.98 564.02 571.08 578.17 584.58 585.27 588.12 590.97 593.82 596.67 599.53 602.38 605.24 608.09 610.95 613.80 616.66 619.51 622.36 625.21 628.06 630.91 633.75 636.59 639.42 642.25 645.08 647.90 650.71 653.52 656.32 663.28 670.19 677.02 683.77 690.41 696.94 703.32 709.53 715.53 721.29 726.75 731.87 736.55 740.69 744.15 746.70 747.97 747.29 743.11 729.01 693.91
2.7144 2.6746 2.6389 2.6069 2.5784 2.5529 2.5303 2.5103 2.4926 2.4772 2.4638 2.4522 2.4423 2.4340 2.4271 2.4216 2.4173 2.4141 2.4120 2.4108 2.4105 2.4105 2.4106 2.4108 2.4112 2.4116 2.4122 2.4129 2.4137 2.4146 2.4156 2.4167 2.4179 2.4191 2.4205 2.4219 2.4234 2.4250 2.4266 2.4283 2.4301 2.4319 2.4338 2.4358 2.4378 2.4398 2.4419 2.4473 2.4529 2.4587 2.4646 2.4705 2.4765 2.4824 2.4881 2.4936 2.4987 2.5034 2.5075 2.5108 2.5131 2.5140 2.5130 2.5094 2.5015 2.4858 2.4474 2.3631
bNormal
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
1592 1563 1534 1505 1477 1449 1420 1392 1364 1336 1309 1281 1253 1226 1198 1171 1144 1116 1089 1062 1038 1035 1024 1014 1003 992 981 971 960 949 938 928 917 906 895 885 874 863 853 842 831 820 810 799 788 777 767 740 713 685 658 631 603 575 546 518 488 458 428 396 364 331 296 260 222 182 136 0
boiling point
167.4 169.6 171.8 174.0 176.1 178.1 180.1 182.1 184.0 185.8 187.5 189.2 190.8 192.3 193.8 195.1 196.3 197.5 198.5 199.4 200.1 200.2 200.5 200.7 201.0 201.2 201.4 201.6 201.7 201.9 202.0 202.0 202.1 202.1 202.2 202.1 202.1 202.0 201.9 201.8 201.7 201.5 201.3 201.1 200.8 200.5 200.2 199.2 198.1 196.7 195.1 193.3 191.2 188.8 186.1 183.1 179.8 176.1 172.0 167.5 162.5 157.0 150.8 144.1 136.5 128.0 117.8 0.0
792.2 719.0 655.9 601.2 553.2 510.9 473.4 439.9 409.8 382.8 358.3 336.1 315.8 297.3 280.3 264.7 250.3 237.0 224.7 213.2 203.5 202.5 198.4 194.4 190.6 186.8 183.2 179.6 176.1 172.7 169.4 166.2 163.0 159.9 156.9 153.9 151.1 148.2 145.5 142.8 140.2 137.6 135.0 132.5 130.1 127.7 125.4 119.7 114.3 109.1 104.1 99.3 94.7 90.2 85.8 81.6 77.4 73.3 69.3 65.4 61.4 57.4 53.4 49.2 44.7 39.6 32.5 —
4.30 4.43 4.55 4.68 4.81 4.93 5.06 5.18 5.31 5.43 5.55 5.68 5.80 5.92 6.04 6.16 6.28 6.41 6.53 6.65 6.76 6.77 6.82 6.87 6.91 6.96 7.01 7.06 7.11 7.16 7.21 7.26 7.31 7.36 7.41 7.47 7.52 7.57 7.62 7.68 7.73 7.79 7.84 7.90 7.95 8.01 8.07 8.22 8.38 8.54 8.71 8.89 9.08 9.29 9.51 9.75 10.01 10.29 10.61 10.96 11.37 11.83 12.39 13.07 13.97 15.27 17.87 —
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 161.6 159.4 157.2 154.9 152.6 150.2 147.9 145.5 143.1 140.8 138.4 136.0 133.6 131.3 128.9 126.6 124.3 122.0 119.8 117.5 115.5 115.3 114.4 113.6 112.7 111.8 111.0 110.1 109.3 108.4 107.6 106.7 105.9 105.1 104.3 103.5 102.7 101.9 101.1 100.3 99.5 98.7 97.9 97.2 96.4 95.7 94.9 93.1 91.3 89.5 87.8 86.1 84.5 82.9 81.3 79.7 78.2 76.8 75.3 73.9 72.5 71.1 69.8 68.4 67.3 66.7 73.0
6.83 7.12 7.43 7.73 8.05 8.38 8.71 9.05 9.40 9.75 10.11 10.48 10.86 11.25 11.64 12.05 12.46 12.88 13.30 13.74 14.14 14.19 14.37 14.55 14.74 14.93 15.11 15.31 15.50 15.69 15.89 16.09 16.29 16.49 16.70 16.90 17.11 17.33 17.54 17.76 17.98 18.21 18.43 18.66 18.90 19.14 19.38 20.00 20.64 21.32 22.03 22.77 23.56 24.39 25.28 26.23 27.26 28.37 29.60 30.96 32.50 34.29 36.44 39.16 42.94 49.21 67.32
28.03 –100 27.33 –95 26.64 –90 25.94 –85 25.26 –80 24.57 –75 23.89 –70 23.22 –65 22.54 –60 21.88 –55 21.21 –50 20.55 –45 19.90 –40 19.25 –35 18.60 –30 17.96 –25 17.32 –20 16.69 –15 16.06 –10 15.44 –5 14.88 –0.49 14.82 0 14.58 2 14.33 4 14.09 6 13.85 8 13.60 10 13.36 12 13.12 14 12.88 16 12.65 18 12.41 20 12.17 22 11.94 24 11.70 26 11.47 28 11.24 30 11.00 32 10.77 34 10.54 36 10.32 38 10.09 40 9.86 42 9.64 44 9.41 46 9.19 48 8.97 50 8.42 55 7.87 60 7.34 65 6.81 70 6.29 75 5.78 80 5.28 85 4.79 90 4.31 95 3.84 100 3.38 105 2.93 110 2.50 115 2.08 120 1.68 125 1.30 130 0.94 135 0.60 140 0.31 145 0.06 150 0.00 151.98 cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Pressure-Enthalpy Diagram for Refrigerant 600a (Isobutane)
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 26
Licensed for single user. © 2017 ASHRAE, Inc.
30.54
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.55
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 600a (Isobutane) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Pres- Density, Volume, kJ/kg Temp.,* sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor °C
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
Liquid Vapor
Liquid Vapor
cp /cv Vapor
–100 0.00038 –95 0.00062 –90 0.00098 –85 0.00151 –80 0.00227 –75 0.00333 –70 0.00478 –65 0.00671 –60 0.00927 –55 0.01258 –50 0.01680 –45 0.02211 –40 0.02870 –35 0.03680 –30 0.04662 –25 0.05843 –20 0.07248 –15 0.08905 –11.75b 0.10133 –10 0.10845 –5 0.13098 0 0.15696 2 0.16839 4 0.18045 6 0.19316 8 0.20654 10 0.22061 12 0.23541 14 0.25094 16 0.26724 18 0.28432 20 0.30222 22 0.32095 24 0.34054 26 0.36102 28 0.38240 30 0.40472 32 0.42800 34 0.45226 36 0.47753 38 0.50384 40 0.53121 42 0.55966 44 0.58923 46 0.61995 48 0.65182 50 0.68490 55 0.77299 60 0.86916 65 0.97386 70 1.0875 75 1.2107 80 1.3438 85 1.4874 90 1.6420 95 1.8081 100 1.9865 105 2.1778 110 2.3826 115 2.6019 120 2.8366 125 3.0880 130 3.3578 134.66c 3.6290
0.0671 0.1208 0.1734 0.2251 0.2759 0.3258 0.3749 0.4232 0.4708 0.5177 0.5640 0.6098 0.6549 0.6995 0.7437 0.7874 0.8306 0.8735 0.9012 0.9160 0.9582 1.0000 1.0167 1.0333 1.0498 1.0663 1.0828 1.0993 1.1157 1.1320 1.1484 1.1647 1.1810 1.1972 1.2134 1.2296 1.2458 1.2619 1.2780 1.2941 1.3102 1.3263 1.3423 1.3583 1.3744 1.3904 1.4064 1.4464 1.4863 1.5263 1.5664 1.6065 1.6469 1.6874 1.7283 1.7696 1.8114 1.8539 1.8974 1.9423 1.9893 2.0400 2.0985 2.2259
1.878 1.894 1.911 1.927 1.944 1.961 1.979 1.997 2.015 2.034 2.054 2.074 2.094 2.115 2.137 2.159 2.182 2.206 2.222 2.231 2.256 2.283 2.293 2.304 2.316 2.327 2.338 2.350 2.362 2.374 2.386 2.398 2.411 2.424 2.437 2.450 2.463 2.477 2.491 2.505 2.520 2.535 2.550 2.566 2.582 2.598 2.615 2.659 2.706 2.757 2.812 2.874 2.942 3.020 3.110 3.217 3.347 3.513 3.736 4.059 4.585 5.629 8.91
1.145 1.143 1.140 1.138 1.136 1.134 1.132 1.130 1.129 1.128 1.126 1.126 1.125 1.125 1.125 1.125 1.125 1.126 1.126 1.127 1.128 1.130 1.131 1.132 1.132 1.133 1.135 1.136 1.137 1.138 1.140 1.141 1.143 1.144 1.146 1.148 1.150 1.152 1.154 1.157 1.159 1.162 1.165 1.168 1.171 1.174 1.178 1.188 1.199 1.213 1.229 1.248 1.272 1.301 1.338 1.385 1.447 1.531 1.650 1.831 2.139 2.770 4.77
683.9 679.1 674.2 669.4 664.5 659.5 654.6 649.6 644.6 639.5 634.4 629.3 624.1 618.9 613.6 608.3 602.9 597.4 593.8 591.9 586.3 580.6 578.3 576.0 573.6 571.3 568.9 566.5 564.2 561.7 559.3 556.9 554.4 551.9 549.4 546.9 544.3 541.7 539.1 536.5 533.9 531.2 528.5 525.8 523.0 520.2 517.4 510.2 502.7 495.0 487.0 478.6 469.9 460.7 451.1 440.7 429.6 417.6 404.3 389.4 372.0 350.6 321.0 225.5
65.234 41.078 26.648 17.764 12.140 8.4874 6.0592 4.4096 3.2662 2.4590 1.8792 1.4561 1.1427 0.90737 0.72839 0.59062 0.48339 0.39904 0.35378 0.33204 0.27833 0.23491 0.21989 0.20604 0.19324 0.18140 0.17044 0.16028 0.15086 0.14210 0.13395 0.12637 0.11930 0.11271 0.10655 0.10080 0.09542 0.09038 0.08566 0.08124 0.07708 0.07317 0.06950 0.06605 0.06279 0.05973 0.05683 0.05029 0.04459 0.03962 0.03525 0.03140 0.02799 0.02496 0.02226 0.01983 0.01764 0.01565 0.01383 0.01214 0.01056 0.00902 0.00742 0.00443
*Temperatures on ITS-90 scale
–6.40 3.04 12.55 22.14 31.82 41.59 51.44 61.38 71.41 81.54 91.76 102.09 112.51 123.04 133.68 144.43 155.30 166.29 173.49 177.40 188.63 200.00 204.59 209.19 213.82 218.47 223.15 227.85 232.57 237.32 242.09 246.88 251.70 256.55 261.42 266.32 271.24 276.19 281.17 286.18 291.22 296.28 301.37 306.50 311.65 316.84 322.06 335.25 348.66 362.29 376.17 390.31 404.73 419.46 434.54 450.00 465.90 482.33 499.39 517.26 536.26 557.01 581.26 633.94
428.19 433.86 439.62 445.46 451.39 457.40 463.48 469.63 475.86 482.14 488.49 494.89 501.35 507.85 514.40 520.99 527.61 534.26 538.60 540.93 547.63 554.34 557.02 559.71 562.40 565.09 567.78 570.47 573.15 575.84 578.52 581.21 583.89 586.56 589.24 591.91 594.57 597.23 599.88 602.53 605.17 607.80 610.43 613.04 615.65 618.24 620.82 627.22 633.53 639.72 645.77 651.64 657.31 662.73 667.86 672.62 676.94 680.70 683.74 685.81 686.46 684.81 678.44 633.94
2.5770 2.5391 2.5052 2.4750 2.4481 2.4242 2.4031 2.3845 2.3683 2.3541 2.3419 2.3315 2.3227 2.3154 2.3095 2.3048 2.3013 2.2989 2.2979 2.2975 2.2969 2.2972 2.2975 2.2980 2.2985 2.2992 2.3000 2.3008 2.3018 2.3028 2.3039 2.3051 2.3064 2.3078 2.3093 2.3108 2.3123 2.3140 2.3157 2.3174 2.3192 2.3211 2.3230 2.3249 2.3269 2.3289 2.3309 2.3361 2.3414 2.3467 2.3520 2.3572 2.3621 2.3667 2.3708 2.3743 2.3769 2.3785 2.3785 2.3765 2.3714 2.3610 2.3396 2.2259
1.131 1.151 1.171 1.191 1.212 1.233 1.254 1.276 1.298 1.321 1.344 1.368 1.393 1.418 1.444 1.471 1.499 1.527 1.547 1.557 1.587 1.619 1.632 1.645 1.658 1.672 1.686 1.699 1.714 1.728 1.743 1.757 1.772 1.788 1.803 1.819 1.835 1.852 1.869 1.886 1.903 1.921 1.939 1.958 1.977 1.997 2.017 2.069 2.125 2.186 2.252 2.326 2.409 2.507 2.625 2.769 2.951 3.189 3.517 4.002 4.806 6.428 11.56 bNormal
Velocity of Sound, m/s
Viscosity, Pa·s
Liquid Vapor
Liquid Vapor
1558 1526 1494 1462 1431 1400 1370 1340 1309 1280 1250 1220 1191 1162 1133 1104 1075 1047 1028 1018 990 961 950 939 928 916 905 894 883 871 860 849 838 826 815 804 793 781 770 759 747 736 725 714 702 691 680 651 622 593 564 535 505 475 444 413 381 348 313 278 240 200 156 0
boiling point
168.3 170.5 172.6 174.7 176.7 178.7 180.6 182.4 184.2 185.9 187.5 189.0 190.4 191.8 193.0 194.1 195.1 196.0 196.5 196.8 197.4 197.9 198.0 198.2 198.3 198.3 198.4 198.4 198.4 198.4 198.3 198.2 198.1 198.0 197.8 197.6 197.4 197.1 196.8 196.5 196.2 195.8 195.3 194.9 194.4 193.9 193.3 191.7 189.9 187.8 185.4 182.7 179.7 176.3 172.6 168.4 163.8 158.7 153.1 146.9 139.9 132.1 123.0 0.0
936.5 837.2 753.2 681.3 619.2 565.3 518.1 476.6 439.8 407.0 377.6 351.2 327.4 305.8 286.2 268.3 251.9 236.9 227.8 223.1 210.3 198.6 194.1 189.8 185.6 181.5 177.5 173.7 170.0 166.3 162.8 159.3 156.0 152.7 149.5 146.5 143.4 140.5 137.6 134.8 132.1 129.4 126.8 124.2 121.7 119.3 116.9 111.1 105.6 100.3 95.2 90.4 85.6 81.0 76.6 72.1 67.8 63.4 59.0 54.5 49.8 44.6 38.5 —
4.40 4.53 4.65 4.78 4.91 5.03 5.16 5.28 5.41 5.53 5.65 5.78 5.90 6.02 6.14 6.26 6.38 6.50 6.58 6.62 6.74 6.86 6.91 6.96 7.01 7.06 7.11 7.16 7.21 7.26 7.31 7.37 7.42 7.47 7.52 7.58 7.63 7.69 7.74 7.80 7.85 7.91 7.97 8.03 8.09 8.15 8.22 8.38 8.56 8.74 8.94 9.16 9.39 9.65 9.95 10.27 10.65 11.09 11.62 12.27 13.09 14.22 15.99 —
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 140.0 138.0 136.0 134.0 131.9 129.8 127.7 125.6 123.5 121.3 119.2 117.1 115.0 112.9 110.8 108.7 106.6 104.6 103.3 102.6 100.6 98.6 97.9 97.1 96.3 95.5 94.8 94.0 93.3 92.5 91.8 91.1 90.3 89.6 88.9 88.2 87.5 86.8 86.1 85.4 84.7 84.1 83.4 82.7 82.1 81.4 80.8 79.2 77.6 76.1 74.6 73.2 71.8 70.4 69.1 67.8 66.6 65.4 64.2 63.1 62.2 61.8 63.3
6.03 6.39 6.76 7.13 7.50 7.88 8.27 8.67 9.06 9.47 9.88 10.29 10.71 11.14 11.57 12.01 12.45 12.90 13.20 13.36 13.83 14.30 14.49 14.68 14.88 15.07 15.27 15.47 15.67 15.88 16.08 16.29 16.50 16.72 16.93 17.15 17.37 17.59 17.82 18.05 18.29 18.52 18.77 19.01 19.26 19.52 19.78 20.45 21.16 21.92 22.72 23.59 24.53 25.56 26.70 27.98 29.44 31.14 33.18 35.72 39.05 43.91 52.88
25.52 –100 24.88 –95 24.24 –90 23.59 –85 22.96 –80 22.32 –75 21.68 –70 21.05 –65 20.41 –60 19.78 –55 19.16 –50 18.53 –45 17.91 –40 17.29 –35 16.67 –30 16.05 –25 15.44 –20 14.83 –15 14.44 –11.75 14.23 –10 13.63 –5 13.03 0 12.79 2 12.55 4 12.32 6 12.08 8 11.84 10 11.61 12 11.38 14 11.14 16 10.91 18 10.68 20 10.45 22 10.22 24 9.99 26 9.76 28 9.53 30 9.31 32 9.08 34 8.86 36 8.63 38 8.41 40 8.19 42 7.97 44 7.75 46 7.53 48 7.31 50 6.78 55 6.24 60 5.72 65 5.21 70 4.71 75 4.21 80 3.73 85 3.26 90 2.80 95 2.36 100 1.93 105 1.53 110 1.14 115 0.78 120 0.46 125 0.18 130 0.00 134.66 cCritical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 27
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 1150 (Ethylene)
30.56
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.57
Refrigerant 1150 (Ethylene) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –169.16a –165 –160 –155 –150 –145 –140 –135 –130 –125 –120 –115 –110 –105 –103.77b –100 –98 –96 –94 –92 –90 –88 –86 –84 –82 –80 –78 –76 –74 –72 –70 –68 –66 –64 –62 –60 –58 –56 –54 –52 –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 5 9.20c
Enthalpy, Pres- Density, Volume, kJ/kg sure, kg/m3 m3/kg MPa Liquid Vapor Liquid Vapor 0.00012 0.00025 0.00053 0.00107 0.00203 0.00362 0.00614 0.01000 0.01565 0.02368 0.03474 0.04961 0.06911 0.09420 0.10133 0.12585 0.14059 0.15662 0.17402 0.19285 0.21320 0.23514 0.25874 0.28409 0.31127 0.34034 0.37141 0.40454 0.43982 0.47733 0.51716 0.55939 0.60411 0.65141 0.70136 0.75406 0.80960 0.86807 0.92955 0.99414 1.0619 1.1330 1.2075 1.2854 1.3669 1.4521 1.5410 1.6339 1.7307 1.8315 1.9366 2.0459 2.1596 2.2779 2.4008 2.5284 2.6610 2.7985 2.9412 3.0893 3.2428 3.4019 3.5669 3.7379 3.9152 4.0990 4.5896 5.0418
654.6 649.3 643.0 636.6 630.1 623.6 617.1 610.5 603.8 597.1 590.3 583.4 576.5 569.4 567.7 562.2 559.3 556.4 553.5 550.5 547.5 544.5 541.4 538.4 535.3 532.2 529.0 525.8 522.6 519.3 516.1 512.7 509.4 506.0 502.5 499.0 495.5 491.9 488.2 484.5 480.8 476.9 473.0 469.1 465.0 460.9 456.7 452.4 448.0 443.5 438.9 434.1 429.2 424.2 419.0 413.6 408.0 402.2 396.1 389.8 383.0 375.9 368.3 360.1 351.1 341.2 308.8 214.2
252.640 129.690 62.7190 32.5570 17.9740 10.4720 6.39510 4.07100 2.68800 1.83310 1.28650 0.92612 0.68196 0.51239 0.47899 0.39198 0.35377 0.32007 0.29026 0.26382 0.24030 0.21933 0.20058 0.18378 0.16869 0.15510 0.14284 0.13176 0.12172 0.11260 0.10431 0.09675 0.08985 0.08354 0.07776 0.07246 0.06758 0.06310 0.05896 0.05514 0.05161 0.04834 0.04530 0.04249 0.03987 0.03743 0.03515 0.03303 0.03105 0.02919 0.02745 0.02581 0.02427 0.02283 0.02146 0.02017 0.01895 0.01780 0.01670 0.01565 0.01466 0.01370 0.01278 0.01189 0.01103 0.01018 0.00802 0.00467
*Temperatures on ITS-90 scale
–158.09 –147.97 –135.81 –123.66 –111.52 –99.41 –87.33 –75.27 –63.22 –51.19 –39.16 –27.12 –15.06 –2.97 0.00 9.16 14.02 18.90 23.79 28.69 33.60 38.53 43.48 48.44 53.41 58.41 63.43 68.47 73.53 78.61 83.72 88.86 94.03 99.23 104.46 109.72 115.02 120.36 125.74 131.16 136.62 142.14 147.70 153.32 158.99 164.73 170.52 176.39 182.33 188.35 194.45 200.65 206.94 213.34 219.85 226.49 233.27 240.21 247.32 254.63 262.17 269.96 278.07 286.56 295.51 305.06 333.52 399.43
409.42 414.35 420.24 426.12 431.97 437.78 443.54 449.24 454.85 460.37 465.79 471.08 476.22 481.21 482.41 486.03 487.90 489.74 491.55 493.32 495.06 496.76 498.43 500.05 501.64 503.18 504.68 506.14 507.55 508.91 510.23 511.49 512.70 513.85 514.95 515.99 516.97 517.88 518.73 519.51 520.21 520.84 521.39 521.86 522.24 522.53 522.72 522.81 522.79 522.65 522.38 521.98 521.44 520.74 519.86 518.80 517.53 516.04 514.28 512.24 509.87 507.10 503.89 500.12 495.66 490.32 470.25 399.43
Liquid
Vapor
Liquid Vapor
Velocity of Sound, m/s cp /cv Vapor Liquid Vapor
–1.1789 –1.0835 –0.9736 –0.8685 –0.7679 –0.6715 –0.5790 –0.4902 –0.4046 –0.3220 –0.2423 –0.1651 –0.0903 –0.0176 0.0000 0.0532 0.0810 0.1085 0.1358 0.1628 0.1896 0.2161 0.2425 0.2686 0.2945 0.3202 0.3457 0.3711 0.3963 0.4214 0.4463 0.4710 0.4957 0.5202 0.5446 0.5689 0.5932 0.6173 0.6414 0.6654 0.6894 0.7133 0.7372 0.7611 0.7850 0.8089 0.8328 0.8568 0.8809 0.9050 0.9292 0.9535 0.9780 1.0027 1.0276 1.0527 1.0781 1.1039 1.1300 1.1567 1.1839 1.2118 1.2406 1.2705 1.3018 1.3350 1.4327 1.6614
4.2787 4.1160 3.9408 3.7848 3.6454 3.5204 3.4080 3.3065 3.2145 3.1310 3.0548 2.9850 2.9210 2.8619 2.8481 2.8073 2.7866 2.7664 2.7468 2.7277 2.7092 2.6911 2.6734 2.6562 2.6394 2.6229 2.6069 2.5911 2.5757 2.5606 2.5457 2.5311 2.5168 2.5026 2.4887 2.4749 2.4614 2.4479 2.4346 2.4214 2.4083 2.3953 2.3824 2.3694 2.3565 2.3436 2.3306 2.3176 2.3045 2.2913 2.2779 2.2643 2.2505 2.2365 2.2221 2.2074 2.1922 2.1765 2.1602 2.1431 2.1252 2.1062 2.0859 2.0640 2.0400 2.0132 1.9243 1.6614
2.429 1.187 2.432 1.188 2.432 1.189 2.429 1.191 2.424 1.194 2.419 1.198 2.414 1.203 2.409 1.210 2.406 1.218 2.404 1.228 2.404 1.240 2.406 1.254 2.409 1.271 2.416 1.290 2.418 1.295 2.424 1.312 2.429 1.321 2.434 1.331 2.439 1.342 2.445 1.353 2.451 1.365 2.458 1.377 2.465 1.391 2.473 1.404 2.482 1.419 2.491 1.434 2.501 1.450 2.512 1.467 2.524 1.484 2.536 1.503 2.549 1.522 2.563 1.543 2.578 1.565 2.594 1.588 2.611 1.612 2.629 1.638 2.648 1.665 2.668 1.694 2.690 1.725 2.714 1.757 2.739 1.792 2.766 1.829 2.795 1.869 2.826 1.912 2.859 1.958 2.895 2.007 2.934 2.061 2.976 2.119 3.022 2.182 3.072 2.252 3.127 2.328 3.188 2.413 3.254 2.507 3.329 2.612 3.412 2.731 3.506 2.866 3.612 3.021 3.735 3.200 3.878 3.411 4.046 3.661 4.247 3.965 4.493 4.339 4.802 4.815 5.201 5.438 5.739 6.291 6.507 7.526 11.680 16.100
1.333 1.333 1.333 1.334 1.334 1.335 1.336 1.337 1.339 1.341 1.344 1.348 1.353 1.358 1.360 1.365 1.368 1.371 1.375 1.379 1.383 1.387 1.392 1.397 1.402 1.408 1.414 1.420 1.427 1.435 1.443 1.451 1.460 1.470 1.480 1.491 1.503 1.516 1.529 1.544 1.560 1.577 1.596 1.616 1.638 1.662 1.688 1.717 1.749 1.784 1.823 1.866 1.914 1.969 2.031 2.101 2.182 2.277 2.387 2.520 2.680 2.877 3.127 3.452 3.894 4.526 8.720
Entropy, kJ/(kg·K)
a Triple
Specific Heat cp , kJ/(kg·K)
point
1767 1740 1707 1673 1639 1604 1569 1534 1498 1463 1427 1390 1354 1316 1307 1279 1264 1249 1233 1218 1203 1187 1172 1156 1140 1125 1109 1093 1077 1061 1044 1028 1012 995 978 962 945 928 911 894 876 859 841 823 806 787 769 751 732 714 695 676 656 637 617 596 576 555 533 511 489 465 441 416 390 362 283 0
202.7 206.7 211.3 215.9 220.3 224.5 228.6 232.5 236.3 239.8 243.2 246.3 249.2 251.8 252.4 254.2 255.1 255.9 256.7 257.4 258.1 258.7 259.3 259.9 260.4 260.8 261.2 261.5 261.8 262.1 262.3 262.4 262.5 262.5 262.4 262.3 262.2 262.0 261.7 261.3 260.9 260.5 259.9 259.3 258.7 257.9 257.1 256.3 255.3 254.3 253.2 252.0 250.7 249.4 248.0 246.4 244.8 243.1 241.3 239.3 237.3 235.1 232.8 230.3 227.6 224.6 214.9 0.0
Viscosity, Pa·s Liquid Vapor 685.7 598.1 514.5 448.6 395.8 352.7 317.2 287.5 262.3 240.7 222.0 205.7 191.3 178.5 175.6 167.0 162.7 158.6 154.7 150.9 147.2 143.6 140.2 136.9 133.7 130.6 127.5 124.6 121.7 118.9 116.2 113.6 111.0 108.5 106.0 103.6 101.2 98.9 96.6 94.4 92.2 90.1 87.9 85.8 83.8 81.8 79.7 77.8 75.8 73.8 71.9 70.0 68.1 66.2 64.3 62.4 60.5 58.7 56.8 54.9 52.9 51.0 49.0 46.9 44.8 42.6 36.2 —
b Normal
0.77 1.64 2.47 3.13 3.66 4.09 4.44 4.74 5.00 5.23 5.44 5.63 5.81 5.99 6.03 6.16 6.23 6.30 6.37 6.44 6.51 6.58 6.65 6.72 6.80 6.87 6.95 7.02 7.10 7.18 7.26 7.34 7.42 7.51 7.60 7.68 7.78 7.87 7.96 8.06 8.16 8.27 8.38 8.49 8.60 8.72 8.84 8.97 9.10 9.24 9.39 9.54 9.69 9.86 10.04 10.22 10.42 10.63 10.86 11.10 11.36 11.65 11.98 12.34 12.75 13.23 15.01 —
boiling point
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor mN/m °C 270.6 264.9 258.1 251.3 244.5 237.9 231.2 224.7 218.3 212.0 205.8 199.7 193.7 187.9 186.5 182.2 179.9 177.7 175.5 173.3 171.2 169.0 166.9 164.8 162.7 160.7 158.6 156.6 154.6 152.6 150.6 148.6 146.7 144.7 142.8 140.9 139.0 137.1 135.2 133.4 131.5 129.7 127.8 126.0 124.2 122.3 120.5 118.7 116.8 115.0 113.2 111.3 109.5 107.6 105.8 103.9 102.0 100.0 98.1 96.1 94.1 92.0 89.9 87.8 85.6 83.5 79.8
6.80 6.62 6.59 6.71 6.92 7.18 7.47 7.78 8.09 8.39 8.70 9.00 9.30 9.61 9.68 9.92 10.04 10.17 10.30 10.44 10.57 10.71 10.85 11.00 11.15 11.30 11.46 11.62 11.79 11.96 12.14 12.33 12.52 12.72 12.92 13.14 13.36 13.59 13.84 14.09 14.35 14.63 14.91 15.22 15.53 15.86 16.21 16.58 16.97 17.38 17.82 18.29 18.78 19.32 19.90 20.53 21.22 21.98 22.83 23.78 24.88 26.15 27.66 29.52 31.88 35.04 52.44
28.14 27.31 26.33 25.35 24.38 23.42 22.47 21.53 20.59 19.66 18.74 17.83 16.93 16.04 15.83 15.16 14.82 14.47 14.12 13.78 13.44 13.10 12.76 12.42 12.09 11.75 11.42 11.09 10.77 10.44 10.12 9.80 9.48 9.16 8.85 8.53 8.23 7.92 7.61 7.31 7.01 6.71 6.42 6.13 5.84 5.55 5.27 4.99 4.71 4.44 4.17 3.90 3.64 3.38 3.13 2.88 2.63 2.39 2.15 1.92 1.70 1.48 1.26 1.06 0.86 0.67 0.25 0.00
–169.16 –165 –160 –155 –150 –145 –140 –135 –130 –125 –120 –115 –110 –105 –103.77 –100 –98 –96 –94 –92 –90 –88 –86 –84 –82 –80 –78 –76 –74 –72 –70 –68 –66 –64 –62 –60 –58 –56 –54 –52 –50 –48 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 5 9.20
c Critical
point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 28 Pressure-Enthalpy Diagram for Refrigerant 1270 (Propylene)
30.58
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.59
Refrigerant 1270 (Propylene) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Temp.,* °C –150 –140 –130 –120 –110 –100 –90 –80 –70 –60 –50 –48 –47.69b –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 55 60 65 70 75 80 85 90 92.42c
Pres- Density, sure, kg/m3 MPa Liquid 0.00001 0.00005 0.00018 0.00059 0.00166 0.00404 0.00881 0.01754 0.03232 0.05578 0.09111 0.09991 0.10133 0.10935 0.11949 0.13033 0.14192 0.15429 0.16748 0.18152 0.19644 0.21228 0.22908 0.24687 0.26569 0.28558 0.30657 0.32871 0.35203 0.37657 0.40237 0.42947 0.45791 0.48773 0.51897 0.55167 0.58588 0.62163 0.65896 0.69793 0.73856 0.78091 0.82502 0.87093 0.91868 0.96832 1.01990 1.07350 1.12900 1.18670 1.24650 1.30840 1.37250 1.43900 1.50770 1.57880 1.65230 1.72820 1.80670 1.88780 1.97150 2.05790 2.28610 2.53250 2.79810 3.08420 3.39220 3.72380 4.08100 4.46680 4.66460
Volume, m3/kg Vapor
Enthalpy, kJ/kg Liquid Vapor
727.9 2789.30000 –117.02 419.51 716.3 583.08000 –98.37 429.53 704.9 156.04000 –79.20 439.78 693.6 50.84700 –59.62 450.25 682.2 19.42300 –39.73 460.91 670.8 8.44420 –19.59 471.73 659.4 4.08250 0.80 482.68 647.8 2.15470 21.43 493.72 636.1 1.22310 42.34 504.79 624.1 0.73772 63.55 515.84 611.9 0.46816 85.11 526.79 609.4 0.42975 89.47 528.96 609.1 0.42416 90.14 529.30 606.9 0.39515 93.84 531.13 604.4 0.36390 98.23 533.29 601.9 0.33564 102.64 535.45 599.4 0.31004 107.06 537.59 596.8 0.28680 111.51 539.73 594.3 0.26566 115.97 541.85 591.7 0.24642 120.45 543.97 589.1 0.22886 124.95 546.08 586.5 0.21282 129.47 548.17 583.8 0.19814 134.01 550.26 581.2 0.18469 138.57 552.33 578.5 0.17234 143.15 554.38 575.8 0.16100 147.76 556.43 573.1 0.15056 152.38 558.46 570.4 0.14093 157.03 560.48 567.6 0.13206 161.71 562.48 564.8 0.12385 166.40 564.46 562.0 0.11626 171.12 566.43 559.2 0.10924 175.87 568.37 556.3 0.10272 180.64 570.30 553.5 0.09667 185.44 572.22 550.6 0.09105 190.27 574.11 547.6 0.08582 195.12 575.98 544.6 0.08094 200.00 577.82 541.6 0.07640 204.91 579.65 538.6 0.07216 209.85 581.45 535.5 0.06820 214.82 583.22 532.4 0.06449 219.83 584.97 529.3 0.06102 224.86 586.69 526.1 0.05777 229.93 588.39 522.9 0.05472 235.03 590.05 519.7 0.05186 240.17 591.68 516.4 0.04917 245.34 593.28 513.0 0.04664 250.55 594.84 509.6 0.04426 255.8 596.37 506.2 0.04202 261.09 597.86 502.7 0.03990 266.42 599.31 499.2 0.03791 271.79 600.72 495.6 0.03602 277.21 602.08 491.9 0.03424 282.67 603.40 488.2 0.03255 288.18 604.67 484.4 0.03095 293.74 605.88 480.6 0.02944 299.35 607.05 476.7 0.02801 305.01 608.15 472.7 0.02664 310.72 609.19 468.6 0.02535 316.50 610.17 464.4 0.02412 322.34 611.07 460.2 0.02295 328.24 611.90 455.8 0.02183 334.21 612.65 444.4 0.01926 349.47 614.13 432.1 0.01696 365.26 614.95 418.8 0.01490 381.71 614.94 404.1 0.01302 399.01 613.87 387.4 0.01130 417.46 611.33 367.5 0.00967 437.66 606.61 341.7 0.00808 461.03 598.06 298.9 0.00630 493.45 579.55 223.4 0.00448 540.41 540.41
*Temperatures on IPTS-68 scale
Entropy, kJ/(kg·K)
Specific Heat cp , kJ/(kg·K)
cp /cv Liquid Vapor Liquid Vapor Vapor –0.6521 –0.5066 –0.3677 –0.2356 –0.1098 0.0100 0.1244 0.2341 0.3395 0.4412 0.5398 0.5592 0.5621 0.5784 0.5976 0.6167 0.6356 0.6545 0.6733 0.6921 0.7107 0.7292 0.7477 0.7661 0.7845 0.8027 0.8209 0.8391 0.8571 0.8752 0.8931 0.9111 0.9289 0.9468 0.9645 0.9823 1.0000 1.0177 1.0353 1.0529 1.0705 1.0881 1.1056 1.1231 1.1407 1.1582 1.1756 1.1931 1.2106 1.2281 1.2456 1.2631 1.2807 1.2982 1.3158 1.3334 1.3510 1.3687 1.3865 1.4043 1.4221 1.4401 1.4854 1.5315 1.5786 1.6273 1.6785 1.7335 1.7964 1.8829 2.0097
3.7046 1.834 0.991 1.249 3.4582 1.893 1.015 1.242 3.2577 1.939 1.039 1.235 3.0936 1.974 1.064 1.229 2.9588 2.002 1.090 1.223 2.8475 2.026 1.118 1.218 2.7555 2.050 1.149 1.214 2.6793 2.075 1.183 1.211 2.6159 2.103 1.220 1.209 2.5631 2.135 1.263 1.209 2.5191 2.171 1.309 1.211 2.5112 2.179 1.319 1.211 2.5100 2.180 1.321 1.211 2.5035 2.187 1.330 1.212 2.4962 2.195 1.340 1.213 2.4891 2.203 1.351 1.214 2.4822 2.212 1.362 1.215 2.4756 2.221 1.373 1.216 2.4692 2.230 1.385 1.217 2.4630 2.239 1.396 1.218 2.4570 2.249 1.408 1.220 2.4512 2.259 1.421 1.221 2.4456 2.269 1.433 1.223 2.4402 2.279 1.446 1.225 2.4350 2.290 1.459 1.227 2.4299 2.300 1.473 1.229 2.4250 2.311 1.487 1.231 2.4203 2.323 1.501 1.234 2.4157 2.335 1.515 1.236 2.4112 2.347 1.530 1.239 2.4068 2.359 1.546 1.242 2.4026 2.372 1.562 1.245 2.3985 2.385 1.578 1.248 2.3945 2.398 1.594 1.252 2.3907 2.412 1.612 1.256 2.3869 2.426 1.629 1.260 2.3832 2.441 1.648 1.264 2.3796 2.456 1.666 1.268 2.3761 2.471 1.686 1.273 2.3726 2.487 1.706 1.278 2.3693 2.504 1.726 1.284 2.3660 2.521 1.748 1.289 2.3627 2.538 1.770 1.296 2.3595 2.557 1.793 1.302 2.3563 2.575 1.817 1.309 2.3532 2.595 1.842 1.316 2.3501 2.615 1.867 1.324 2.3470 2.637 1.894 1.333 2.3440 2.659 1.923 1.342 2.3409 2.682 1.952 1.351 2.3378 2.706 1.983 1.362 2.3348 2.731 2.015 1.373 2.3317 2.757 2.049 1.385 2.3286 2.785 2.085 1.398 2.3255 2.814 2.123 1.412 2.3223 2.845 2.163 1.427 2.3191 2.878 2.206 1.443 2.3158 2.913 2.251 1.460 2.3124 2.950 2.300 1.480 2.3090 2.990 2.352 1.501 2.3054 3.032 2.408 1.523 2.3017 3.078 2.469 1.549 2.2919 3.211 2.644 1.624 2.2809 3.380 2.868 1.723 2.2683 3.604 3.163 1.859 2.2535 3.921 3.578 2.055 2.2353 4.416 4.213 2.363 2.2120 5.315 5.332 2.918 2.1790 7.507 7.921 4.221 2.1200 20.390 21.470 11.110 2.0097 bNormal
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor Liquid Vapor Liquid Vapor mN/m °C Velocity of Sound, m/s
1834 1754 1683 1617 1554 1493 1432 1370 1309 1247 1184 1171 1169 1158 1145 1133 1120 1107 1094 1081 1068 1056 1043 1030 1017 1004 991 978 965 952 939 925 912 899 886 873 860 847 834 820 807 794 781 767 754 741 727 714 701 687 674 661 647 634 620 606 593 579 565 552 538 524 488 452 415 377 337 294 248 195 0
boiling point
174.3 180.7 186.9 192.7 198.3 203.5 208.4 212.9 216.9 220.4 223.4 223.9 224.0 224.4 224.8 225.3 225.7 226.1 226.4 226.7 227.0 227.3 227.5 227.7 227.9 228.1 228.2 228.3 228.3 228.3 228.3 228.2 228.1 228.0 227.8 227.6 227.4 227.1 226.7 226.4 226.0 225.5 225.0 224.5 223.9 223.3 222.6 221.9 221.1 220.3 219.4 218.5 217.5 216.5 215.4 214.3 213.1 211.8 210.5 209.1 207.7 206.2 202.2 197.7 192.8 187.4 181.3 174.7 167.2 159.2 0.0
Viscosity, Pa·s
1324.0 945.7 718.7 571.2 468.9 394.1 337.1 292.2 255.9 226.1 201.1 196.6 195.9 192.2 188.0 183.9 179.9 176.0 172.3 168.7 165.1 161.7 158.4 155.1 152.0 148.9 145.9 143.0 140.2 137.4 134.8 132.1 129.6 127.1 124.6 122.2 119.9 117.6 115.4 113.2 111.1 109.0 106.9 104.9 102.9 101.0 99.1 97.2 95.4 93.6 91.8 90.0 88.3 86.6 84.9 83.3 81.6 80.0 78.4 76.8 75.2 73.6 69.7 65.8 61.9 57.9 53.7 49.1 43.8 36.3 —
3.51 3.79 4.06 4.34 4.62 4.90 5.18 5.46 5.73 6.01 6.29 6.35 6.36 6.40 6.46 6.52 6.57 6.63 6.69 6.75 6.80 6.86 6.92 6.98 7.04 7.10 7.17 7.23 7.29 7.36 7.42 7.49 7.56 7.63 7.69 7.77 7.84 7.91 7.99 8.07 8.15 8.23 8.31 8.4 8.48 8.57 8.67 8.76 8.86 8.97 9.06 9.17 9.28 9.40 9.52 9.64 9.77 9.91 10.05 10.20 10.36 10.52 10.97 11.49 12.10 12.83 13.73 14.90 16.55 19.59 —
179.9 177.6 175.0 172.2 169.1 165.8 162.3 158.6 154.7 150.6 146.4 145.6 145.4 144.7 143.8 143.0 142.1 141.2 140.3 139.4 138.5 137.6 136.7 135.7 134.8 133.9 132.9 132.0 131.0 130.1 129.1 128.2 127.2 126.2 125.3 124.3 123.3 122.3 121.3 120.3 119.3 118.3 117.3 116.3 115.3 114.3 113.3 112.2 111.2 110.2 109.2 108.1 107.1 106.0 105.0 104.0 102.9 101.9 100.8 99.7 98.7 97.6 94.9 92.2 89.4 86.6 83.8 81.1 79.2 81.9
4.60 5.07 5.55 6.06 6.60 7.16 7.75 8.36 9.01 9.69 10.40 10.55 10.58 10.70 10.86 11.01 11.17 11.33 11.49 11.65 11.82 11.99 12.16 12.34 12.52 12.70 12.89 13.08 13.28 13.48 13.68 13.89 14.11 14.33 14.55 14.78 15.02 15.27 15.52 15.79 16.06 16.33 16.62 16.92 17.23 17.55 17.88 18.22 18.58 18.95 19.34 19.74 20.15 20.58 21.03 21.50 22.00 22.52 23.06 23.63 24.24 24.88 26.63 28.67 31.07 33.94 37.50 42.14 48.98 64.05
33.49 31.76 30.05 28.36 26.69 25.04 23.41 21.80 20.22 18.66 17.14 16.83 16.79 16.53 16.23 15.93 15.63 15.34 15.04 14.75 14.45 14.16 13.87 13.58 13.29 13.00 12.72 12.43 12.15 11.87 11.59 11.31 11.03 10.76 10.48 10.21 9.94 9.67 9.40 9.13 8.87 8.60 8.34 8.08 7.82 7.56 7.31 7.05 6.80 6.55 6.31 6.06 5.82 5.58 5.34 5.10 4.86 4.63 4.40 4.17 3.95 3.73 3.18 2.65 2.15 1.67 1.21 0.79 0.41 0.10 0.00
–150 –140 –130 –120 –110 –100 –90 –80 –70 –60 –50 –48 –47.69 –46 –44 –42 –40 –38 –36 –34 –32 –30 –28 –26 –24 –22 –20 –18 –16 –14 –12 –10 –8 –6 –4 –2 0 2 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 44 46 48 50 55 60 65 70 75 80 85 90 92.42
c Critical
point
Licensed for single user. © 2017 ASHRAE, Inc. Pressure
30.60 This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 29 Pressure-Enthalpy Diagram for Refrigerant 704 (Helium) Note: The reference states for enthalpy and entropy differ from those in the table.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.61
Refrigerant 704 (Helium) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Enthalpy, Absolute Density, Volume, kJ/kg Temp.,* Pressure, kg/m3 m3/kg Liquid Vapor Liquid Vapor MPa K 2.18a 2.20 2.30 2.40 2.50 2.60 2.70 2.80 2.90 3.00 3.10 3.20 3.30 3.40 3.50 3.60 3.70 3.80 3.90 4.00 4.10 4.20 4.23b 4.30 4.40 4.50 4.60 4.70 4.80 4.90 5.00 5.10 5.20c
0.00486 0.00515 0.00653 0.00814 0.01000 0.01213 0.01454 0.01727 0.02032 0.02373 0.02750 0.03166 0.03622 0.04121 0.04664 0.05252 0.05888 0.06573 0.07310 0.08100 0.08945 0.09847 0.10132 0.10809 0.11832 0.12920 0.14075 0.15301 0.16602 0.17983 0.19453 0.21023 0.22746
146.24 146.19 145.87 145.46 144.96 144.38 143.72 143.00 142.21 141.34 140.42 139.43 138.38 137.25 136.06 134.80 133.45 132.03 130.51 128.90 127.17 125.32 124.73 123.33 121.17 118.81 116.20 113.27 109.90 105.89 100.83 93.53 69.64
0.87297 0.83068 0.67831 0.56257 0.47252 0.40113 0.34364 0.29674 0.25805 0.22582 0.19871 0.17574 0.15612 0.13925 0.12466 0.11195 0.10081 0.09101 0.08232 0.07459 0.06767 0.06144 0.05967 0.05581 0.05067 0.04596 0.04161 0.03753 0.03367 0.02993 0.02617 0.02206 0.01436
2.34 2.48 2.98 3.35 3.66 3.93 4.18 4.44 4.70 4.97 5.26 5.56 5.88 6.22 6.58 6.96 7.35 7.77 8.21 8.67 9.16 9.68 9.84 10.22 10.80 11.42 12.09 12.83 13.64 14.57 15.69 17.20 21.71
25.56 25.66 26.05 26.44 26.81 27.17 27.52 27.86 28.19 28.50 28.79 29.07 29.33 29.57 29.79 29.98 30.16 30.31 30.43 30.52 30.57 30.59 30.59 30.57 30.50 30.36 30.16 29.87 29.45 28.87 28.02 26.63 21.71
Velocity of Specific Heat cp , Sound, m/s kJ/(kg·K) cp /cv Vapor Liquid Vapor Vapor Liquid Vapor
Liquid 1.4040 1.4682 1.6865 1.8414 1.9607 2.0604 2.1502 2.2356 2.3196 2.4039 2.4894 2.5765 2.6653 2.7559 2.8482 2.9422 3.0380 3.1355 3.2351 3.3367 3.4407 3.5475 3.5806 3.6577 3.7719 3.8912 4.0171 4.1517 4.2986 4.4641 4.6615 4.9283 5.7639
12.0746 12.0035 11.7174 11.4586 11.2211 11.0011 10.7956 10.6024 10.4198 10.2464 10.0808 9.9222 9.7694 9.6216 9.4781 9.3381 9.2007 9.0653 8.9312 8.7975 8.6633 8.5277 8.4861 8.3896 8.2476 8.0999 7.9440 7.7767 7.5924 7.3821 7.1273 6.7774 5.7639
aLower
*Temperatures on EPT-76 scale
Viscosity, Pa·s
Thermal Cond., mW/(m·K)
Liquid Vapor
Liquid Vapor
Entropy, kJ/(kg·K)
6.318 6.061 5.800 6.076 4.164 6.139 3.217 6.199 2.700 6.258 2.453 6.318 2.372 6.380 2.394 6.446 2.477 6.516 2.598 6.592 2.740 6.676 2.897 6.768 3.062 6.872 3.234 6.989 3.414 7.122 3.603 7.274 3.803 7.449 4.020 7.654 4.257 7.894 4.523 8.179 4.826 8.521 5.179 8.938 5.299 9.083 5.600 9.455 6.118 10.1100 6.776 10.9640 7.646 12.117 8.863 13.7540 10.7000 16.2440 13.813 20.464 20.2400 29.0940 40.7700 55.8660
1.747 1.750 1.763 1.778 1.795 1.813 1.834 1.857 1.882 1.910 1.941 1.976 2.015 2.059 2.108 2.165 2.229 2.303 2.389 2.491 2.611 2.756 2.806 2.934 3.158 3.448 3.838 4.388 5.224 6.637 9.531 18.545
217 217 216 216 216 217 217 217 216 214 213 210 208 205 202 199 196 193 189 186 182 178 176 173 169 164 159 153 147 140 133 124 0
83.2 83.6 85.0 86.3 87.6 88.8 90.0 91.1 92.1 93.0 93.9 94.8 95.5 96.3 96.9 97.5 98.1 98.6 99.1 99.5 99.8 100.1 100.2 100.4 100.6 100.8 100.9 101.0 101.1 101.3 101.6 102.5 0.0 b Normal
lambda point
— — — — — — — — — — — — — — — 3.5 3.4 3.4 3.3 3.3 3.2 3.2 3.2 3.1 3.1 3.0 3.0 2.9 2.8 2.7 2.6 2.5 —
— — — — — — — — — — — — — — — 1.00 1.04 1.07 1.11 1.15 1.19 1.24 1.25 1.28 1.33 1.38 1.43 1.48 1.55 1.62 1.70 1.80 —
— — — — — — — — — — — — — — — 17.9 18.1 18.2 18.4 18.5 18.6 18.6 18.7 18.7 18.8 18.8 18.8 18.9 19.0 19.1 19.3 19.9
— — — — — — — — — — — — — — — 7.31 7.56 7.82 8.09 8.36 8.66 8.97 9.06 9.30 9.66 10.07 10.54 11.08 11.76 12.63 13.86 15.77
Surface Tension, Temp.,* mN/m K 0.388 0.385 0.371 0.356 0.342 0.328 0.314 0.300 0.286 0.272 0.258 0.244 0.231 0.217 0.203 0.190 0.177 0.164 0.150 0.138 0.125 0.112 0.108 0.099 0.087 0.075 0.063 0.051 0.040 0.029 0.018 0.008 0.000
2.18 2.20 2.30 2.40 2.50 2.60 2.70 2.80 2.90 3.00 3.10 3.20 3.30 3.40 3.50 3.60 3.70 3.80 3.90 4.00 4.10 4.20 4.23 4.30 4.40 4.50 4.60 4.70 4.80 4.90 5.00 5.10 5.20
c Critical
boiling point
point
Refrigerant 704 (Helium) Properties of Gas at 0.101 325 MPa (one standard atmosphere) Thermal Cond., mW/(m·K)
Vel. of Sound, m/s
Thermal Cond., mW/(m·K)
cp /cv
Vel. of Sound, m/s
Visc., Pa·s
2.806
100.2
1.25
9.05
100
0.1307
1953.16
32.7129
5.193
1.667 1137.0
23.15
181.41
5.327
1.708
213.6
2.72
20.17
120
0.1240
2057.02
32.9840
5.193
1.667 1167.0
24.00
188.10
18.2572
5.236
1.678
284.4
3.93
28.69
140
0.1180
2160.88
33.2417
5.193
1.667 1196.3
24.84
194.69
20.1325
5.213
1.671
340.1
4.93
35.90
160
0.1126
2264.74
33.4872
5.193
1.667 1224.9
25.67
201.19 207.60
Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) –268.9b 16.758
30.59
8.4861
9.083
–260
3.7508
82.19
15.2748
–250
2.1049
134.88
–240
1.4677
187.10
Temp., Density, Enthalpy, Entropy, cp , kJ/ kJ/kg kJ/(kg· K) (kg· K) °C kg/m3
cp /cv
Visc., Pa·s
–220
0.9155
291.20
22.5899
5.200
1.668
430.1
6.60
48.56
180
0.1076
2368.60
33.7216
5.193
1.667 1252.9
26.49
–200
0.6655
395.16
24.2500
5.196
1.667
504.3
8.04
59.86
200
0.1031
2472.46
33.9459
5.193
1.667 1280.2
27.29
213.93
–180
0.5228
499.06
25.5057
5.195
1.667
568.8
9.35
70.30
220
0.0989
2576.32
34.1609
5.193
1.667 1307.0
28.09
220.18 226.35
–160
0.4305
602.95
26.5160
5.194
1.667
626.7
10.39
80.09
240
0.0950
2680.18
34.3673
5.193
1.667 1333.2
28.88
–140
0.3659
706.82
27.3614
5.194
1.667
679.7
11.56
89.41
260
0.0915
2784.04
34.5659
5.193
1.667 1358.9
29.66
232.46
–120
0.3182
810.69
28.0882
5.193
1.666
728.8
12.68
98.32
280
0.0882
2887.90
34.7571
5.193
1.667 1384.1
30.43
238.50
–100
0.2815
914.56
28.7256
5.193
1.666
774.9
13.75
106.90
300
0.0851
2991.76
34.9416
5.193
1.667 1408.9
31.20
244.47
–80
0.2524
1018.42
29.2932
5.193
1.667
818.3
14.79
115.20
320
0.0822
3095.62
35.1197
5.193
1.667 1433.3
31.96
250.39
–60
0.2287
1122.28
29.8049
5.193
1.667
859.6
15.80
123.25
340
0.0795
3199.48
35.2919
5.193
1.667 1457.2
32.71
256.24
–40
0.2091
1226.14
30.2707
5.193
1.667
898.9
16.79
131.09
360
0.0770
3303.34
35.4586
5.193
1.667 1480.8
33.45
262.04
–20
0.1926
1330.00
30.6980
5.193
1.667
936.7
17.75
138.73
380
0.0747
3407.20
35.6201
5.193
1.667 1504.0
34.19
267.79
0
0.1785
1433.87
31.0929
5.193
1.667
972.9
18.70
146.20
400
0.0725
3511.06
35.7767
5.193
1.667 1526.8
34.92
273.48
20
0.1663
1537.73
31.4599
5.193
1.667 1007.9
19.62
153.50
420
0.0704
3614.92
35.9288
5.193
1.667 1549.3
35.65
279.13
40
0.1557
1641.58
31.8026
5.193
1.667 1041.6
20.52
160.67
440
0.0684
3718.78
36.0765
5.193
1.667 1571.5
36.37
284.73
60
0.1464
1745.44
32.1241
5.193
1.667 1074.4
21.41
167.70
460
0.0665
3822.64
36.2201
5.193
1.667 1593.4
37.08
290.28
80
0.1381
1849.30
32.4269
5.193
1.667 1106.1
22.29
174.61
480
0.0648
3926.50
36.3599
5.193
1.667 1615.0
37.79
295.78
500
0.0631
4030.36
36.4960
5.193
1.667 1636.3
38.49
301.25
b Saturated
vapor at normal boiling point
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 30 Pressure-Enthalpy Diagram for Refrigerant 728 (Nitrogen)
30.62
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.63
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 728 (Nitrogen) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Absolute Density, Volume, kJ/kg Temp.,* Pressure, kg/m3 m3/kg Liquid Vapor Liquid Vapor MPa K
Liquid
63.15a 64 66 68 70 72 74 76 77.35b 78 80 82 84 86 88 90 92 94 96 98 100 105 110 115 120 125 126.19c
2.4257 2.4524 2.5140 2.5739 2.6321 2.6889 2.7442 2.7983 2.8342 2.8511 2.9028 2.9534 3.0031 3.0520 3.1000 3.1473 3.1940 3.2401 3.2858 3.3311 3.3761 3.4882 3.6015 3.7198 3.8514 4.0373 4.2149
0.01252 0.01460 0.02062 0.02848 0.03854 0.05121 0.06691 0.08610 0.10132 0.10926 0.13687 0.16947 0.20757 0.25174 0.30251 0.36046 0.42616 0.50020 0.58316 0.67565 0.77827 1.08330 1.46580 1.93700 2.51060 3.20690 3.39580
867.2 863.7 855.4 847.0 838.5 829.9 821.1 812.2 806.1 803.1 793.9 784.6 775.0 765.2 755.2 745.0 734.5 723.8 712.7 701.2 689.4 657.5 621.5 578.7 523.4 426.1 313.3
1.48310 1.28720 0.93698 0.69642 0.52743 0.40625 0.31772 0.25192 0.21682 0.20226 0.16422 0.13470 0.11152 0.09310 0.07831 0.06632 0.05651 0.04842 0.04169 0.03605 0.03129 0.02224 0.01598 0.01146 0.00799 0.00487 0.00319
–150.73 –149.03 –145.02 –141.00 –136.97 –132.93 –128.87 –124.79 –122.02 –120.70 –116.58 –112.43 –108.26 –104.05 –99.81 –95.52 –91.18 –86.79 –82.34 –77.81 –73.21 –61.27 –48.49 –34.39 –17.87 6.40 29.23
64.78 65.59 67.47 69.31 71.10 72.83 74.50 76.11 77.16 77.64 79.10 80.47 81.75 82.93 84.01 84.97 85.81 86.52 87.10 87.51 87.77 87.56 85.84 81.91 74.17 55.03 29.23
Velocity of Specific Heat cp , Sound, m/s kJ/(kg·K) cp /cv Vapor Liquid Vapor Vapor Liquid Vapor
Entropy, kJ/(kg·K)
5.8383 5.8059 5.7336 5.6667 5.6045 5.5466 5.4925 5.4417 5.4090 5.3939 5.3487 5.3059 5.2651 5.2262 5.1888 5.1527 5.1178 5.0839 5.0507 5.0181 4.9858 4.9055 4.8226 4.7311 4.6185 4.4263 4.2149 a Triple
*Temperatures on ITS-90 scale
2.000 1.058 1.411 2.002 1.061 1.413 2.005 1.066 1.417 2.010 1.073 1.421 2.014 1.082 1.427 2.020 1.091 1.433 2.027 1.102 1.441 2.035 1.114 1.450 2.041 1.124 1.457 2.045 1.129 1.461 2.056 1.145 1.473 2.068 1.163 1.487 2.083 1.184 1.503 2.099 1.208 1.521 2.119 1.235 1.542 2.141 1.266 1.567 2.166 1.300 1.594 2.196 1.341 1.626 2.231 1.387 1.664 2.271 1.440 1.707 2.318 1.503 1.758 2.479 1.714 1.931 2.743 2.062 2.221 3.240 2.749 2.778 4.508 4.631 4.216 16.720 23.740 16.930
995 987 966 946 926 906 885 865 851 845 824 804 783 762 741 719 697 675 652 629 605 543 476 403 317 195 0 b Normal
point
161.1 162.1 164.3 166.4 168.4 170.3 172.1 173.8 174.8 175.3 176.7 178.0 179.2 180.2 181.0 181.8 182.4 182.8 183.1 183.3 183.3 182.5 180.8 177.7 172.6 160.3 0.0
Viscosity, Pa·s
Thermal Cond., Surface mW/(m·K) Tension, Temp.,* Liquid Vapor Liquid Vapor mN/m K 311.6 4.38 297.5 4.44 267.8 4.59 242.3 4.73 220.2 4.88 201.1 5.03 184.3 5.19 169.6 5.34 160.7 5.44 156.6 5.49 145.1 5.65 134.8 5.81 125.6 5.97 117.2 6.14 109.7 6.31 102.8 6.48 96.5 6.66 90.7 6.84 85.4 7.03 80.4 7.23 75.8 7.43 65.3 7.98 56.0 8.63 47.3 9.44 38.4 10.62 26.9 13.33 — —
173.2 171.5 167.5 163.5 159.5 155.5 151.5 147.5 144.8 143.5 139.5 135.6 131.6 127.7 123.7 119.8 115.9 111.9 108.0 104.0 100.1 90.3 80.4 70.6 61.0 56.4
5.62 5.71 5.92 6.14 6.35 6.58 6.80 7.03 7.19 7.26 7.51 7.76 8.01 8.29 8.57 8.87 9.19 9.53 9.89 10.29 10.73 12.04 13.83 16.58 21.72 41.54
12.24 12.03 11.55 11.07 10.59 10.12 9.65 9.19 8.87 8.73 8.27 7.83 7.38 6.94 6.51 6.09 5.67 5.25 4.84 4.44 4.05 3.10 2.21 1.39 0.66 0.08 0.00
63.15 64 66 68 70 72 74 76 77.35 78 80 82 84 86 88 90 92 94 96 98 100 105 110 115 120 125 126.19
c Critical
boiling point
point
Refrigerant 728 (Nitrogen) Properties of Gas at 0.101 325 MPa (one standard atmosphere) Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) –195.8b –180 –160 –140 –120 –100 –90 –80 –70 –60 –50 –40 –30 –20 –10 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 b Saturated
4.6121 3.7571 3.0593 2.5858 2.2414 1.9789 1.8696 1.7719 1.6839 1.6043 1.5320 1.4659 1.4053 1.3496 1.2981 1.2504 1.2061 1.1648 1.1263 1.0903 1.0565 1.0247 0.9948 0.9666 0.9399 0.9147 0.8908 0.8681 0.8466 0.8261
77.16 94.51 115.90 137.02 158.02 178.95 189.40 199.85 210.28 220.71 231.14 241.56 251.98 262.40 272.82 283.23 293.65 304.06 314.47 324.89 335.30 345.72 356.14 366.56 376.99 387.42 397.85 408.30 418.74 429.20
5.4090 5.6133 5.8213 5.9932 6.1402 6.2687 6.3273 6.3829 6.4355 6.4857 6.5335 6.5791 6.6229 6.6649 6.7053 6.7441 6.7815 6.8177 6.8526 6.8864 6.9191 6.9509 6.9817 7.0117 7.0408 7.0691 7.0967 7.1236 7.1498 7.1755
vapor at normal boiling point
1.124 1.081 1.061 1.052 1.048 1.045 1.045 1.044 1.043 1.043 1.042 1.042 1.042 1.042 1.042 1.041 1.041 1.041 1.041 1.041 1.042 1.042 1.042 1.042 1.043 1.043 1.044 1.044 1.045 1.046
cp /cv
Vel. of Sound, m/s
Thermal Cond., Visc., Pa·s mW/(m· K)
1.457 1.432 1.419 1.412 1.409 1.406 1.405 1.405 1.404 1.404 1.403 1.403 1.403 1.402 1.402 1.402 1.402 1.401 1.401 1.401 1.401 1.400 1.400 1.400 1.400 1.399 1.399 1.398 1.398 1.397
174.8 194.0 215.2 234.2 251.6 267.9 275.6 283.1 290.4 297.5 304.4 311.2 317.9 324.4 330.7 337.0 343.1 349.1 355.0 360.8 366.5 372.1 377.7 383.1 388.5 393.7 398.9 404.0 409.1 414.1
5.44 6.51 7.81 9.06 10.27 11.42 11.99 12.54 13.08 13.61 14.14 14.65 15.16 15.66 16.15 16.63 17.10 17.57 18.03 18.49 18.94 19.38 19.82 20.25 20.68 21.10 21.52 21.93 22.34 22.74
7.19 8.72 10.63 12.49 14.28 16.03 16.88 17.71 18.54 19.35 20.15 20.94 21.72 22.49 23.25 24.00 24.74 25.47 26.20 26.91 27.62 28.32 29.01 29.69 30.37 31.04 31.70 32.36 33.01 33.66
Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300 310 320 330 340 350 360 370 380 390 400 420 440 460 480 500
0.8065 0.7879 0.7701 0.7531 0.7368 0.7212 0.7063 0.6920 0.6782 0.6650 0.6523 0.6401 0.6283 0.6169 0.6060 0.5954 0.5852 0.5753 0.5658 0.5565 0.5476 0.5390 0.5306 0.5225 0.5146 0.5069 0.4923 0.4785 0.4655 0.4531 0.4414
439.66 450.14 460.62 471.12 481.62 492.14 502.67 513.22 523.78 534.36 544.95 555.56 566.19 576.84 587.50 598.19 608.90 619.62 630.37 641.14 651.93 662.74 673.58 684.44 695.32 706.22 728.10 750.08 772.15 794.32 816.59
7.2005 7.2249 7.2489 7.2723 7.2952 7.3177 7.3397 7.3613 7.3825 7.4033 7.4238 7.4439 7.4636 7.4831 7.5022 7.5210 7.5395 7.5577 7.5757 7.5934 7.6109 7.6281 7.6451 7.6618 7.6783 7.6947 7.7267 7.7580 7.7885 7.8183 7.8475
1.047 1.048 1.049 1.050 1.051 1.053 1.054 1.055 1.057 1.059 1.060 1.062 1.064 1.066 1.068 1.070 1.072 1.074 1.076 1.078 1.080 1.082 1.085 1.087 1.089 1.092 1.096 1.101 1.106 1.111 1.116
cp /cv
Vel. of Sound, m/s
Thermal Cond., Visc., Pa·s mW/(m· K)
1.397 1.396 1.396 1.395 1.394 1.394 1.393 1.392 1.391 1.390 1.389 1.389 1.388 1.387 1.386 1.385 1.384 1.383 1.381 1.380 1.379 1.378 1.377 1.376 1.375 1.374 1.371 1.369 1.367 1.365 1.363
419.0 423.8 428.6 433.3 438.0 442.6 447.1 451.6 456.0 460.4 464.7 468.9 473.2 477.3 481.5 485.5 489.6 493.6 497.5 501.4 505.3 509.1 512.9 516.7 520.4 524.1 531.4 538.6 545.6 552.6 559.4
23.14 23.53 23.92 24.31 24.69 25.07 25.44 25.81 26.18 26.54 26.90 27.26 27.62 27.97 28.32 28.66 29.00 29.35 29.68 30.02 30.35 30.68 31.01 31.34 31.66 31.98 32.62 33.24 33.87 34.48 35.08
34.30 34.93 35.56 36.18 36.80 37.42 38.03 38.63 39.23 39.83 40.42 41.01 41.59 42.17 42.75 43.32 43.89 44.46 45.02 45.58 46.13 46.69 47.23 47.78 48.33 48.87 49.94 51.00 52.06 53.10 54.14
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 31 Pressure-Enthalpy Diagram for Refrigerant 729 (Air)
30.64
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.65
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 729 (Air) Properties of Liquid on the Bubble Line and Vapor on the Dew Line Enthalpy, Density, Volume, Absolute Temperature,* K kJ/kg kg/m3 m3/kg Pressure, MPa Bubble Dew Liquid Vapor Liquid Vapor 0.00526 0.008 0.01 0.015 0.02 0.03 0.04 0.05 0.06 0.08 0.1 0.10132b 0.15 0.2 0.3 0.4 0.5 0.6 0.8 1 1.5 2 2.5 3 3.78781c
59.75 61.83 63.00 65.27 66.99 69.58 71.57 73.19 74.58 76.89 78.79 78.90 82.51 85.39 89.82 93.26 96.12 98.59 102.76 106.23 113.15 118.58 123.11 127.03 132.53
63.09 65.12 66.27 68.47 70.15 72.68 74.61 76.19 77.53 79.78 81.62 81.73 85.23 88.02 92.32 95.65 98.42 100.80 104.83 108.18 114.83 120.02 124.32 127.99 132.53
957.8 949.2 944.4 934.9 927.6 916.5 907.9 900.8 894.7 884.3 875.7 875.1 858.4 844.6 822.7 805.0 789.7 776.0 751.9 730.4 682.6 638.2 592.9 541.8 342.6
3.42745 2.32401 1.88990 1.29870 0.99558 0.68481 0.52523 0.42756 0.36139 0.27713 0.22550 0.22277 0.15488 0.11849 0.08100 0.06165 0.04977 0.04169 0.03136 0.02500 0.01623 0.01162 0.00869 0.00658 0.00292
–162.40 –158.45 –156.22 –151.92 –148.64 –143.68 –139.88 –136.77 –134.10 –129.64 –125.95 –125.73 –118.68 –113.01 –104.17 –97.17 –91.26 –86.07 –77.12 –69.40 –53.12 –39.06 –25.82 –12.26 29.38
62.73 64.67 65.75 67.82 69.35 71.62 73.30 74.63 75.75 77.55 78.96 79.05 81.55 83.38 85.84 87.42 88.49 89.22 89.99 90.12 88.65 85.27 80.04 72.34 29.38
Viscosity, Pa·s
Thermal Cond., mW/(m·K)
Liquid
Velocity of Specific Heat cp , Sound, m/s kJ/(kg·K) cp /cv Vapor Liquid Vapor Vapor Liquid Vapor
Liquid Vapor
Liquid Vapor
Surface Tension, mN/m
2.4517 2.5165 2.5522 2.6193 2.6687 2.7412 2.7948 2.8378 2.8738 2.9323 2.9794 2.9822 3.0688 3.1357 3.2353 3.3104 3.3715 3.4235 3.5098 3.5811 3.7230 3.8379 3.9407 4.0419 4.3479
6.1507 6.0614 6.0142 5.9293 5.8697 5.7866 5.7282 5.6832 5.6467 5.5893 5.5450 5.5424 5.4648 5.4079 5.3272 5.2691 5.2231 5.1848 5.1221 5.0708 4.9678 4.8805 4.7966 4.7056 4.3479
424.8 4.47 377.9 4.62 354.6 4.70 315.3 4.87 289.6 4.99 256.4 5.18 234.8 5.33 219.0 5.45 206.8 5.55 188.7 5.72 175.6 5.86 174.8 5.87 153.5 6.15 139.3 6.37 120.9 6.72 108.9 7.00 100.1 7.25 93.3 7.47 83.0 7.86 75.3 8.21 61.9 9.00 52.5 9.79 44.8 10.68 37.7 11.82 — —
185.2 181.2 179.0 174.7 171.4 166.5 162.7 159.7 157.0 152.7 149.1 148.9 142.1 136.8 128.5 122.1 116.8 112.2 104.4 98.0 85.2 75.2 66.9 60.0 —
14.95 14.41 14.11 13.54 13.11 12.46 11.97 11.57 11.24 10.68 10.22 10.20 9.35 8.68 7.67 6.90 6.28 5.75 4.88 4.18 2.85 1.89 1.16 0.59 0.00
Entropy, kJ/(kg·K)
1.895 1.897 1.899 1.902 1.904 1.909 1.913 1.917 1.921 1.928 1.936 1.936 1.953 1.969 2.002 2.034 2.068 2.102 2.176 2.258 2.515 2.899 3.558 5.008 — bBubble
*Temperatures on ITS-90 scale
1.014 1.017 1.019 1.023 1.026 1.033 1.039 1.044 1.050 1.060 1.070 1.071 1.094 1.117 1.162 1.207 1.253 1.299 1.398 1.506 1.844 2.359 3.27 5.303 —
1.405 1.407 1.408 1.411 1.414 1.420 1.425 1.430 1.434 1.443 1.452 1.452 1.471 1.409 1.527 1.564 1.601 1.639 1.721 1.813 2.101 2.538 3.286 4.895 —
1056 1035 1023 1001 984 959 940 925 911 889 871 869 834 806 762 727 697 670 625 585 500 427 359 290 —
158.9 161.3 162.6 165.0 166.8 169.4 171.3 172.8 174.0 175.9 177.3 177.4 179.9 181.6 183.7 184.9 185.6 186.0 186.3 186.0 184.2 181.4 177.8 173.4 —
5.83 6.04 6.16 6.38 6.55 6.80 7.00 7.16 7.29 7.52 7.71 7.72 8.07 8.37 8.88 9.37 9.82 10.24 11.03 11.81 13.87 16.31 19.55 24.56 —
c Critical
and dew points at one standard atmosphere
point
Refrigerant 729 (Air) Properties of Gas at 0.101 325 MPa (one standard atmosphere) Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) –191.4d –180 –160 –140 –120 –100 –90 –80 –70 –60 –50 –40 –30 –20 –10 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 dSaturated
4.4889 3.8887 3.1648 2.6744 2.3178 2.0462 1.9332 1.8321 1.7411 1.6588 1.5840 1.5156 1.4530 1.3953 1.3421 1.2927 1.2469 1.2043 1.1644 1.1272 1.0922 1.0594 1.0284 0.9993 0.9717 0.9456 0.9209 0.8975 0.8752 0.8540
79.05 91.09 111.75 132.15 152.43 172.63 182.71 192.79 202.86 212.93 222.99 233.06 243.11 253.17 263.23 273.29 283.35 293.41 303.48 313.55 323.62 333.70 343.79 353.88 363.98 374.09 384.21 394.34 404.49 414.64
5.5424 5.6804 5.8814 6.0474 6.1893 6.3133 6.3699 6.4235 6.4743 6.5227 6.5688 6.6130 6.6552 6.6957 6.7347 6.7722 6.8084 6.8433 6.8771 6.9098 6.9414 6.9721 7.0020 7.0310 7.0592 7.0866 7.1134 7.1395 7.1650 7.1899
1.071 1.044 1.025 1.016 1.012 1.009 1.008 1.007 1.007 1.007 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.006 1.007 1.007 1.008 1.008 1.009 1.010 1.011 1.011 1.012 1.014 1.015 1.016
vapor at dew-point temperature
cp /cv
Vel. of Sound, m/s
1.452 1.435 1.421 1.414 1.410 1.408 1.407 1.406 1.406 1.405 1.405 1.404 1.404 1.404 1.403 1.403 1.402 1.402 1.402 1.401 1.401 1.400 1.400 1.399 1.399 1.398 1.397 1.397 1.396 1.395
177.4 190.8 211.7 230.4 247.6 263.5 271.1 278.5 285.7 292.7 299.5 306.2 312.7 319.1 325.4 331.5 337.5 343.4 349.2 354.9 360.5 366.0 371.4 376.7 381.9 387.1 392.1 397.1 402.1 406.9
Visc., Pa·s 5.87 6.68 8.04 9.35 10.61 11.82 12.41 12.99 13.55 14.11 14.66 15.20 15.73 16.25 16.76 17.27 17.77 18.26 18.75 19.22 19.70 20.16 20.62 21.08 21.52 21.97 22.41 22.84 23.27 23.69
Thermal Cond., mW/(m·K) 7.72 8.84 10.76 12.63 14.45 16.20 17.06 17.91 18.74 19.56 20.37 21.17 21.96 22.74 23.51 24.27 25.02 25.77 26.50 27.22 27.94 28.65 29.36 30.05 30.74 31.42 32.10 32.77 33.43 34.09
Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300 310 320 330 340 350 360 370 380 390 400 420 440 460 480 500
0.8338 0.8145 0.7961 0.7785 0.7617 0.7456 0.7302 0.7154 0.7011 0.6875 0.6743 0.6617 0.6495 0.6378 0.6264 0.6155 0.6049 0.5947 0.5849 0.5753 0.5661 0.5572 0.5485 0.5401 0.5320 0.5241 0.5089 0.4947 0.4812 0.4684 0.4563
424.81 434.99 445.18 455.39 465.62 475.86 486.12 496.40 506.70 517.02 527.35 537.71 548.09 558.49 568.91 579.35 589.82 600.30 610.81 621.35 631.90 642.48 653.09 663.71 674.37 685.04 706.47 727.98 749.60 771.31 793.12
7.2142 7.2379 7.2612 7.2840 7.3063 7.3282 7.3497 7.3707 7.3914 7.4117 7.4317 7.4513 7.4705 7.4895 7.5082 7.5266 7.5447 7.5625 7.5801 7.5974 7.6145 7.6313 7.6479 7.6643 7.6805 7.6965 7.7279 7.7585 7.7884 7.8176 7.8461
1.017 1.019 1.020 1.022 1.023 1.025 1.027 1.029 1.031 1.033 1.035 1.037 1.039 1.041 1.043 1.045 1.048 1.050 1.052 1.055 1.057 1.059 1.062 1.064 1.066 1.069 1.074 1.078 1.083 1.088 1.093
cp /cv 1.394 1.394 1.393 1.392 1.391 1.390 1.389 1.388 1.387 1.386 1.385 1.384 1.383 1.381 1.380 1.379 1.378 1.377 1.376 1.375 1.373 1.372 1.371 1.370 1.369 1.368 1.365 1.363 1.361 1.359 1.357
Vel. of Sound, m/s 411.7 416.4 421.1 425.7 430.2 434.7 439.1 443.5 447.8 452.0 456.2 460.4 464.5 468.6 472.6 476.6 480.5 484.4 488.3 492.1 495.9 499.7 503.4 507.1 510.7 514.3 521.5 528.5 535.4 542.3 549.0
Thermal Cond., Visc., Pa·s mW/(m·K) 24.11 24.52 24.93 25.34 25.74 26.13 26.53 26.92 27.30 27.69 28.06 28.44 28.81 29.18 29.55 29.91 30.27 30.63 30.98 31.34 31.69 32.03 32.38 32.72 33.06 33.40 34.06 34.72 35.37 36.02 36.65
34.75 35.39 36.04 36.67 37.31 37.93 38.56 39.17 39.79 40.40 41.00 41.60 42.20 42.80 43.39 43.97 44.56 45.14 45.71 46.29 46.86 47.42 47.99 48.55 49.11 49.66 50.76 51.86 52.94 54.01 55.08
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Fig. 32
Licensed for single user. © 2017 ASHRAE, Inc.
Pressure-Enthalpy Diagram for Refrigerant 732 (Oxygen)
30.66
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.67
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant 732 (Oxygen) Properties of Saturated Liquid and Saturated Vapor Enthalpy, Absolute Density, Volume, kJ/kg Temp.,* Pressure, kg/m3 m3/kg MPa K Liquid Vapor Liquid Vapor
Liquid
54.36a 55 60 65 70 75 80 85 90 90.19b 95 100 105 110 115 120 125 130 135 140 145 150 154.58c
2.0921 2.1117 2.2571 2.3912 2.5156 2.6313 2.7397 2.8417 2.9383 2.9419 3.0303 3.1185 3.2033 3.2855 3.3657 3.4444 3.5222 3.6001 3.6791 3.7612 3.8498 3.9546 4.2008
0.00015 0.00018 0.00073 0.00233 0.00626 0.01455 0.03012 0.05683 0.09935 0.10133 0.16308 0.25400 0.37853 0.54340 0.75559 1.02230 1.35090 1.74910 2.22500 2.78780 3.44770 4.21860 5.04300
1306.1 96.54300 –193.61 1303.5 80.00900 –192.55 1282.0 21.46200 –184.19 1259.7 7.21880 –175.81 1237.0 2.89250 –167.42 1213.9 1.32930 –159.02 1190.5 0.68099 –150.61 1166.6 0.38047 –142.18 1142.1 0.22794 –133.69 1141.2 0.22386 –133.37 1116.9 0.14450 –125.12 1090.9 0.09592 –116.45 1063.8 0.06612 –107.64 1035.5 0.04699 –98.64 1005.6 0.03424 –89.42 973.9 0.02544 –79.90 939.7 0.01919 –70.02 902.5 0.01463 –59.66 861.0 0.01120 –48.65 813.2 0.00856 –36.70 755.1 0.00646 –23.22 675.5 0.00465 –6.67 436.1 0.00229 32.42
49.11 49.68 54.19 58.66 63.09 67.45 71.69 75.75 79.55 79.69 83.04 86.16 88.85 91.05 92.72 93.75 94.06 93.47 91.74 88.47 82.83 72.56 32.42
Velocity of Viscosity, Thermal Cond., Specific Heat cp , Surface Sound, m/s Pa·s mW/(m·K) kJ/(kg·K) cp /cv Tension, Temp.,* Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor mN/m K
Entropy, kJ/(kg·K)
6.5571 6.5159 6.2301 5.9985 5.8086 5.6510 5.5185 5.4055 5.3076 5.3042 5.2215 5.1445 5.0746 5.0100 4.9495 4.8915 4.8349 4.7780 4.7191 4.6552 4.5812 4.4828 4.2008 a Triple
*Temperatures on ITS-90 scale
1.673 1.672 1.673 1.677 1.678 1.679 1.682 1.688 1.699 1.699 1.715 1.738 1.767 1.807 1.858 1.927 2.021 2.153 2.354 2.691 3.368 5.464
0.926 0.928 0.948 0.967 0.978 0.979 0.974 0.969 0.970 0.971 0.982 1.006 1.046 1.101 1.177 1.276 1.411 1.600 1.886 2.370 3.369 6.625
1.395 1.394 1.390 1.387 1.387 1.392 1.402 1.417 1.436 1.437 1.460 1.491 1.528 1.576 1.638 1.721 1.835 2.000 2.252 2.682 3.561 6.315
1123 1127 1127 1102 1066 1027 987 947 906 904 864 822 779 735 689 642 592 540 484 423 355 274 0 b Normal
point
140.3 141.1 147.0 152.6 158.1 163.3 168.4 173.1 177.3 177.5 181.0 184.1 186.4 188.1 189.1 189.4 189.0 187.8 185.7 182.8 178.8 172.8 0.0
773.6 747.5 578.1 457.9 371.8 308.7 261.2 224.6 195.6 194.7 172.1 152.6 135.9 121.5 108.8 97.4 87.1 77.6 68.7 60.2 51.9 42.9 —
4.10 4.15 4.55 4.96 5.36 5.75 6.15 6.54 6.94 6.95 7.33 7.73 8.13 8.55 8.98 9.43 9.91 10.45 11.06 11.82 12.88 14.72 —
201.9 201.0 193.9 186.8 179.7 172.6 165.4 158.3 151.0 150.8 143.8 136.6 129.2 121.9 114.6 107.2 99.9 92.6 85.4 78.2 71.1 64.2
4.42 4.48 4.98 5.49 5.99 6.51 7.03 7.57 8.12 8.15 8.71 9.34 10.01 10.75 11.57 12.51 13.61 14.94 16.64 18.98 22.58 29.67
22.68 22.50 21.12 19.76 18.41 17.09 15.78 14.49 13.22 13.17 11.98 10.75 9.56 8.39 7.25 6.14 5.07 4.04 3.05 2.13 1.27 0.51 0.00
54.36 55 60 65 70 75 80 85 90 90.19 95 100 105 110 115 120 125 130 135 140 145 150 154.58
c Critical
boiling point
point
Refrigerant 732 (Oxygen) Properties of Gas at 0.101 325 MPa (one standard atmosphere) Thermal Cond., Temp.,* Density, Enthalpy, Entropy, cp , kJ/ Visc., °C kg/m3 Pa·s mW/(m·K) kJ/kg kJ/(kg· K) (kg· K)
Visc., Pa·s
1.379 1.377
389.3 393.7
26.96 27.43
35.69 36.41
0.954 0.957 0.961
1.375 1.373 1.372
398.0 402.2 406.4
27.90 28.37 28.83
37.11 37.81 38.51
6.8400 6.8602 6.8801
0.964 0.967 0.970
1.370 1.368 1.367
410.5 414.5 418.5
29.28 29.73 30.18
39.21 39.89 40.58
464.40 474.14 527.35
6.8996 6.9188 7.4345
0.973 0.976 1.035
1.365 1.363 1.385
422.5 426.4 456.2
30.63 31.06 27.56
41.26 41.94 40.55
0.7313 0.7178 0.7048
493.73 503.58 513.45
6.9562 6.9745 6.9925
0.983 0.986 0.989
1.360 1.359 1.357
434.2 438.0 441.7
31.93 32.36 32.78
43.28 43.94 44.61
290 300 310
0.6923 0.6802 0.6686
523.36 533.29 543.26
7.0103 7.0278 7.0450
0.992 0.995 0.998
1.355 1.354 1.352
445.4 449.1 452.8
33.20 33.62 34.04
45.27 45.92 46.57
25.95 26.73 27.51
320 330 340
0.6573 0.6464 0.6358
553.26 563.28 573.34
7.0620 7.0788 7.0953
1.001 1.004 1.007
1.351 1.350 1.348
456.4 460.0 463.6
34.45 34.85 35.26
47.22 47.87 48.51
21.91 22.44
28.29 29.05
350 360
0.6256 0.6157
583.43 593.54
7.1116 7.1277
1.010 1.013
1.347 1.345
467.1 470.6
35.66 36.06
49.15 49.78
352.1 357.0 361.9 366.6 371.3 375.9
22.97 23.49 24.00 24.51 25.01 25.50
29.82 30.57 31.32 32.06 32.80 33.53
370 380 390 400 420 440
0.6062 0.5969 0.5879 0.5792 0.5624 0.5467
603.68 613.86 624.05 634.28 654.81 675.45
7.1436 7.1593 7.1748 7.1901 7.2202 7.2495
1.016 1.019 1.021 1.024 1.029 1.034
1.344 1.343 1.342 1.340 1.338 1.336
474.1 477.5 480.9 484.3 491.0 497.6
36.46 36.85 37.24 37.63 38.39 39.15
50.42 51.05 51.68 52.30 53.54 54.78
380.5 384.9
25.99 26.48
34.26 34.98
460 480 500
0.5318 0.5176 0.5042
696.19 717.02 737.95
7.2782 7.3062 7.3336
1.039 1.044 1.049
1.334 1.332 1.330
504.2 510.6 517.0
39.90 40.64 41.36
56.00 57.21 58.42
cp /cv
Vel. of Sound, m/s
–183.0b –180
4.4671 4.3120
79.69 82.52
5.3042 5.3352
0.971 0.948
1.437 1.439
177.5 181.1
6.95 7.18
8.15 8.43
150 160
0.9215 0.9002
387.55 397.05
6.7333 6.7555
0.948 0.951
–160 –140 –120
3.5050 2.9596 2.564
101.23 119.78 138.20
5.5171 5.6681 5.7970
0.931 0.924 0.919
1.422 1.414 1.409
201.0 218.8 235.2
8.71 10.18 11.59
10.34 12.22 14.07
170 180 190
0.8799 0.8605 0.8419
406.58 416.14 425.72
6.7772 6.7986 6.8195
–100 –90 –80
2.2629 2.1378 2.0258
156.56 165.72 174.87
5.9096 5.9611 6.0097
0.916 0.916 0.915
1.407 1.406 1.405
250.4 257.6 264.6
12.96 13.62 14.28
15.88 16.76 17.64
200 210 220
0.8241 0.8070 0.7906
435.35 445.00 454.68
–70 –60 –50
1.9251 1.8340 1.7512
184.02 193.16 202.30
6.0559 6.0998 6.1418
0.915 0.914 0.914
1.404 1.404 1.403
271.5 278.1 284.6
14.92 15.55 16.17
18.51 19.37 20.22
230 240 250
0.7749 0.7598 0.6743
–40 –30 –20
1.6756 1.6063 1.5425
211.45 220.59 229.74
6.1818 6.2202 6.2571
0.914 0.915 0.915
1.402 1.401 1.401
290.9 297.1 303.1
16.78 17.39 17.98
21.06 21.90 22.72
260 270 280
–10 0 10
1.4836 1.4290 1.3784
238.90 248.06 257.23
6.2926 6.3268 6.3597
0.916 0.917 0.918
1.400 1.399 1.398
309.0 314.8 320.5
18.57 19.14 19.71
23.54 24.35 25.15
20 30 40
1.3312 1.2871 1.2459
266.41 275.61 284.82
6.3916 6.4225 6.4524
0.919 0.920 0.922
1.397 1.396 1.395
326.0 331.4 336.7
20.27 20.83 21.37
50 60
1.2073 1.1709
294.05 303.29
6.4814 6.5095
0.924 0.926
1.394 1.392
342.0 347.1
70 80 90 100 110 120
1.1367 1.1045 1.0740 1.0452 1.0179 0.9919
312.56 321.85 331.16 340.49 349.85 359.23
6.5369 6.5636 6.5896 6.6150 6.6397 6.6639
0.928 0.930 0.932 0.935 0.937 0.940
1.391 1.390 1.388 1.387 1.385 1.384
130 140
0.9673 0.9439
368.64 378.08
6.6875 6.7107
0.943 0.945
1.382 1.38
*Temperatures on IPTS-68 scale
Thermal Cond., mW/(m·K)
Vel. of Sound, m/s
Temp.,* Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K)
cp /cv
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Pressure
Licensed for single user. © 2017 ASHRAE, Inc.
Fig. 33 Pressure-Enthalpy Diagram for Refrigerant 740 (Argon)
30.68
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.69
Refrigerant 740 (Argon) Properties of Saturated Liquid and Saturated Vapor
Licensed for single user. © 2017 ASHRAE, Inc.
Absolute Density, Volume, Temp.,* Pressure, kg/m3 m3/kg MPa K Liquid Vapor 83.81a 84 86 87.30b 88 90 92 94 96 98 100 102 104 106 108 110 112 114 116 118 120 125 130 135 140 145 150 150.69c
0.06889 0.07045 0.08811 0.10132 0.10901 0.13351 0.16199 0.19485 0.23249 0.27532 0.32377 0.37825 0.43920 0.50706 0.58226 0.66526 0.75650 0.85644 0.96553 1.08420 1.21300 1.58230 2.02550 2.55090 3.16820 3.88960 4.73460 4.86300
1416.8 1415.6 1403.4 1395.4 1391.1 1378.6 1366.0 1353.2 1340.3 1327.1 1313.7 1300.1 1286.2 1272.0 1257.6 1242.8 1227.6 1212.1 1196.2 1179.7 1162.8 1117.9 1068.1 1011.5 943.7 854.3 680.4 535.6
0.24663 0.24164 0.19687 0.17320 0.16198 0.13448 0.11258 0.09496 0.08067 0.06897 0.05932 0.05129 0.04458 0.03892 0.03413 0.03004 0.02654 0.02351 0.02090 0.01862 0.01663 0.01263 0.00966 0.00739 0.00559 0.00409 0.00253 0.00187
Enthalpy, kJ/kg
Entropy, kJ/(kg·K)
Velocity of Sound, m/s cp /cv Liquid Vapor Vapor Liquid Vapor
Liquid
Vapor
Liquid
Vapor
–121.44 –121.22 –118.98 –117.52 –116.74 –114.49 –112.23 –109.96 –107.68 –105.38 –103.06 –100.72 –98.37 –95.98 –93.57 –91.13 –88.65 –86.14 –83.59 –81.00 –78.35 –71.49 –64.16 –56.18 –47.16 –36.19 –17.88 –4.33
42.28 42.36 43.13 43.62 43.87 44.57 45.23 45.85 46.41 46.93 47.40 47.81 48.17 48.46 48.68 48.84 48.92 48.93 48.85 48.68 48.41 47.27 45.30 42.21 37.47 29.76 11.52 –4.33
1.3295 1.3321 1.3583 1.3750 1.3839 1.4090 1.4335 1.4577 1.4814 1.5048 1.5278 1.5505 1.5730 1.5952 1.6172 1.6390 1.6606 1.6821 1.7035 1.7248 1.7461 1.7995 1.8538 1.9102 1.9712 2.0425 2.1589 2.2476
3.2830 1.116 0.555 1.709 3.2794 1.116 0.556 1.710 3.2433 1.116 0.562 1.719 3.2208 1.117 0.566 1.725 3.2090 1.118 0.568 1.728 3.1763 1.121 0.576 1.740 3.1451 1.125 0.584 1.752 3.1152 1.131 0.593 1.766 3.0865 1.137 0.603 1.782 3.0590 1.145 0.614 1.800 3.0324 1.154 0.627 1.820 3.0068 1.164 0.641 1.842 2.9819 1.175 0.656 1.867 2.9578 1.188 0.673 1.894 2.9343 1.202 0.691 1.926 2.9114 1.218 0.712 1.960 2.8890 1.235 0.736 2.000 2.8669 1.255 0.762 2.044 2.8452 1.278 0.791 2.094 2.8238 1.303 0.825 2.151 2.8025 1.332 0.863 2.216 2.7495 1.425 0.986 2.426 2.6957 1.564 1.172 2.741 2.6390 1.790 1.482 3.262 2.5757 2.225 2.104 4.258 2.4973 3.399 3.896 6.880 2.3550 23.580 35.470 43.160 2.2476 a Triple
*Temperatures on ITS-90 scale
Viscosity, Pa·s
Thermal Cond., mW/(m·K)
Liquid Vapor
Liquid Vapor
Surface Tension, Temp.,* mN/m K
290.2 288.4 270.9 260.3 254.8 240.0 226.4 213.8 202.2 191.3 181.3 172.0 163.3 155.2 147.6 140.4 133.7 127.4 121.4 115.7 110.2 97.6 85.9 74.7 63.6 52.1 36.8 —
133.6 133.3 130.4 128.5 127.4 124.5 121.6 118.7 115.8 113.0 110.2 107.4 104.6 101.9 99.1 96.4 93.7 91.0 88.4 85.7 83.1 76.7 70.4 64.2 58.1 52.0 57.6
13.42 13.37 12.86 12.53 12.35 11.85 11.36 10.86 10.38 9.89 9.42 8.94 8.48 8.02 7.56 7.11 6.67 6.23 5.80 5.38 4.96 3.95 3.00 2.10 1.29 0.58 0.04 0.00
Specific Heat cp , kJ/(kg·K)
862 861 847 838 833 819 805 791 776 762 747 732 717 701 685 669 653 636 619 602 584 538 488 433 372 297 175 0
168.1 168.3 169.9 170.9 171.4 172.8 174.2 175.5 176.7 177.8 178.9 179.9 180.8 181.6 182.4 183.0 183.6 184.1 184.5 184.8 185.1 185.3 184.8 183.7 181.5 176.6 157.0 0.0
b Normal
point
6.86 6.87 7.05 7.17 7.23 7.41 7.60 7.78 7.97 8.16 8.35 8.54 8.74 8.95 9.15 9.37 9.59 9.81 10.04 10.29 10.54 11.23 12.03 13.02 14.32 16.27 21.18 —
5.36 5.37 5.52 5.62 5.68 5.83 6.00 6.16 6.33 6.51 6.69 6.88 7.08 7.28 7.50 7.73 7.98 8.24 8.52 8.82 9.15 10.14 11.45 13.34 16.39 22.53 55.88
83.81 84 86 87.3 88 90 92 94 96 98 100 102 104 106 108 110 112 114 116 118 120 125 130 135 140 145 150 150.69
c Critical
boiling point
point
Refrigerant 740 (Argon) Properties of Gas at 0.101 325 MPa (one standard atmosphere) Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) –185.8b –180 –160 –140 –120 –100 –90 –80 –70 –60 –50 –40 –30 –20 –10 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140
5.7736 5.3774 4.3721 3.6930 3.2000 2.8246 2.6684 2.5287 2.4031 2.2894 2.1861 2.0917 2.0052 1.9256 1.8520 1.7840 1.7207 1.6618 1.6068 1.5554 1.5071 1.4618 1.4191 1.3788 1.3408 1.3048 1.2707 1.2383 1.2076 1.1783
b Saturated
vapor at normal boiling point
43.62 46.90 57.80 68.48 79.06 89.58 94.83 100.08 105.31 110.55 115.78 121.00 126.23 131.45 136.67 141.89 147.11 152.33 157.54 162.76 167.97 173.18 178.39 183.61 188.82 194.03 199.24 204.45 209.66 214.87
3.2208 3.2571 3.3632 3.4501 3.5242 3.5887 3.6182 3.6461 3.6725 3.6977 3.7217 3.7446 3.7665 3.7876 3.8078 3.8273 3.8460 3.8641 3.8816 3.8985 3.9149 3.9308 3.9462 3.9612 3.9758 3.9899 4.0037 4.0171 4.0302 4.0430
0.566 0.556 0.538 0.531 0.527 0.525 0.525 0.524 0.524 0.523 0.523 0.523 0.522 0.522 0.522 0.522 0.522 0.522 0.522 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521
cp /cv 1.725 1.715 1.696 1.686 1.681 1.677 1.676 1.675 1.674 1.673 1.672 1.672 1.671 1.671 1.671 1.670 1.670 1.670 1.669 1.669 1.669 1.669 1.669 1.669 1.669 1.668 1.668 1.668 1.668 1.668
Vel. of Sound, m/s
Thermal Cond., Visc., Pa·s mW/(m·K)
170.9 177.2 196.6 214.0 229.9 244.7 251.8 258.6 265.3 271.8 278.1 284.3 290.4 296.3 302.2 307.9 313.5 319.0 324.4 329.7 334.9 340.1 345.1 350.1 355.1 359.9 364.7 369.4 374.1 378.7
7.17 7.66 9.32 10.93 12.50 14.02 14.77 15.50 16.22 16.94 17.64 18.33 19.02 19.69 20.36 21.02 21.67 22.31 22.94 23.56 24.18 24.79 25.40 25.99 26.58 27.17 27.74 28.32 28.88 29.44
5.62 6.01 7.30 8.56 9.79 10.99 11.57 12.15 12.72 13.28 13.83 14.38 14.91 15.44 15.97 16.48 16.99 17.50 17.99 18.48 18.97 19.45 19.92 20.39 20.85 21.31 21.76 22.21 22.66 23.09
Temp., Density, Enthalpy, Entropy, cp , kJ/ °C kg/m3 kJ/kg kJ/(kg· K) (kg· K) 150 160 170 180 190 200 210 220 230 240 250 260 270 280 290 300 310 320 330 340 350 360 370 380 390 400 420 440 460 480 500
1.1504 1.1239 1.0985 1.0742 1.0510 1.0288 1.0075 0.9870 0.9674 0.9485 0.9304 0.9129 0.8961 0.8799 0.8643 0.8492 0.8346 0.8206 0.8069 0.7938 0.7810 0.7687 0.7567 0.7452 0.7339 0.7230 0.7022 0.6825 0.6638 0.6462 0.6295
220.08 225.28 230.49 235.70 240.91 246.12 251.32 256.53 261.74 266.94 272.15 277.36 282.56 287.77 292.98 298.18 303.39 308.59 313.80 319.01 324.21 329.42 334.62 339.83 345.03 350.24 360.65 371.06 381.47 391.88 402.29
4.0554 4.0676 4.0795 4.0911 4.1025 4.1136 4.1245 4.1351 4.1456 4.1558 4.1659 4.1758 4.1854 4.1949 4.2043 4.2134 4.2224 4.2313 4.2400 4.2485 4.2570 4.2652 4.2734 4.2814 4.2893 4.2971 4.3124 4.3272 4.3416 4.3556 4.3692
0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.521 0.520 0.520 0.520 0.520
cp /cv
Vel. of Sound, m/s
1.668 1.668 1.668 1.668 1.668 1.668 1.668 1.668 1.668 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667 1.667
383.3 387.8 392.2 396.6 401.0 405.3 409.6 413.8 418.0 422.1 426.2 430.2 434.2 438.2 442.2 446.1 450.0 453.8 457.6 461.4 465.1 468.8 472.5 476.2 479.8 483.4 490.5 497.6 504.5 511.3 518.1
Thermal Cond., Visc., Pa·s mW/(m·K) 29.99 30.54 31.09 31.62 32.16 32.68 33.21 33.73 34.24 34.75 35.25 35.75 36.25 36.74 37.23 37.71 38.20 38.67 39.15 39.61 40.08 40.54 41.00 41.46 41.91 42.36 43.25 44.13 45.00 45.86 46.70
23.53 23.96 24.38 24.80 25.22 25.64 26.04 26.45 26.85 27.25 27.65 28.04 28.43 28.81 29.19 29.57 29.95 30.32 30.69 31.06 31.42 31.78 32.14 32.50 32.85 33.20 33.90 34.58 35.26 35.93 36.59
Licensed for single user. © 2017 ASHRAE, Inc. Enthalpy
30.70 This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 34 Enthalpy-Concentration Diagram for Ammonia/Water Solutions Prepared by Kwang Kim and Keith Herold, Center for Environmental Energy Engineering, University of Maryland at College Park
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
30.71
Specific Volume of Saturated Ammonia-Water Solutions, m3/kg Concentration, Ammonia (Mass basis)
Temp., °C
0
10
20
30
40
50
60
70
80
90
100
Temp., °C
–10
0.00100
0.00103
0.00106
0.00109
0.00114
0.00118
0.00122
0.00128
0.00135
0.00142
0.00151
–10
0
0.00100
0.00103
0.00107
0.00110
0.00114
0.00119
0.00124
0.00130
0.00137
0.00146
0.00156
0
10
0.00100
0.00104
0.00107
0.00111
0.00115
0.00120
0.00125
0.00132
0.00139
0.00149
0.00160
10
20
0.00100
0.00104
0.00108
0.00112
0.00116
0.00121
0.00127
0.00133
0.00142
0.00152
0.00164
20
30
0.00100
0.00105
0.00108
0.00113
0.00117
0.00123
0.00128
0.00135
0.00145
0.00156
0.00168
30
40
0.00101
0.00105
0.00109
0.00114
0.00119
0.00124
0.00130
0.00138
0.00148
0.00159
0.00173
40
50
0.00101
0.00106
0.00110
0.00115
0.00120
0.00125
0.00132
0.00140
0.00151
0.00163
0.00177
50
60
0.00102
0.00106
0.00111
0.00116
0.00121
0.00127
0.00134
0.00143
0.00154
0.00167
0.00183
60
70
0.00102
0.00107
0.00112
0.00117
0.00122
0.00129
0.00136
0.00146
0.00158
0.00172
0.00190
70
80
0.00103
0.00108
0.00113
0.00118
0.00124
0.00130
0.00139
0.00149
0.00162
0.00178
0.00198
80
90
0.00104
0.00109
0.00114
0.00119
0.00125
0.00132
0.00141
0.00153
0.00167
0.00184
0.00208
90
100
0.00104
0.00110
0.00115
0.00121
0.00127
0.00135
0.00145
0.00157
0.00172
0.00191
0.00219
100
Prepared under ASHRAE research project RP-271, sponsored by TC 8.3. Data reference: B.H. Jennings, Ammonia water properties (paper presented at ASHRAE meeting, January 1965).
Licensed for single user. © 2017 ASHRAE, Inc.
Refrigerant Temperature (t = °C) and Enthalpy (h = kJ/kg) of Lithium Bromide Solutions Percent LiBr
Temp., (t = °C) 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180
0
10
20
30
40
45
50
55
60
65
70
t
20.0
19.1
17.7
15.0
9.8
5.8
–0.4
–7.7
–15.8
–23.4#
–29.3#
h
84.0
67.4
52.6
40.4
33.5
33.5
38.9
53.2
78.0
111.0#
145.0#
t
30.0
29.0
27.5
24.6
19.2
15.0
8.6
1.0
–7.3
–15.2#
–21.6#
h
125.8
103.3
84.0
68.6
58.3
56.8
60.5
73.5
96.8
128.4#
161.7#
t
40.0
38.9
37.3
34.3
28.5
24.1
17.5
9.8
1.3
–7.0#
–14.0#
146.0#
178.3#
h
167.6
139.5
115.8
96.0
82.5
79.7
82.2
93.5
115.4
t
50.0
48.8
47.2
44.0
37.9
33.3
26.5
18.5
9.9
1.3
–6.3#
h
209.3
175.2
147.0
123.4
106.7
102.6
103.8
114.0
134.5
163.5
195.0#
t
60.0
58.8
57.0
53.6
47.3
42.5
35.5
27.3
18.4
9.5
1.4#
h
251.1
211.7
179.1
151.4
131.7
125.8
125.8
134.7
153.7
181.4
211.9#
t
70.0
68.7
66.8
63.3
56.6
51.6
44.4
36.1
27.0
17.7
9.0#
h
293.0
247.7
210.5
178.8
155.7
148.9
148.0
155.6
173.2
199.4
228.8#
t
80.0
78.6
76.7
73.0
66.0
60.8
53.4
44.8
35.6
26.0
16.7#
h
334.9
287.8
243.6
207.3
181.0
172.8
170.0
176.2
192.6
217.2
245.7#
t
90.0
88.6
86.5
82.6
75.4
70.0
62.3
53.6
44.1
34.2
24.3#
h
376.9
321.1
275.6
235.4
206.1
195.8
192.3
197.1
212.2
235.6
262.9#
t
100.0
98.5
96.3
92.3
84.7
79.1
71.3
62.4
52.7
42.4
32.0
h
419.0
357.6
307.9
263.8
231.0
219.9
214.6
218.2
231.5
253.5
279.7
t
110.0
108.4
106.2
101.9
94.1
88.3
80.2
71.1
61.3
50.6
39.7
h
461.3
394.3
340.1
292.4
255.9
243.3
236.8
239.1
251.0
271.4
296.3
t
120.0*
118.3*
116.0*
111.6
103.4
97.5
89.2
79.9
69.8
58.9
47.3
h
503.7*
431.0*
372.5*
320.9
281.0
267.0
259.0
260.0
270.2
289.5
313.4
t
130.0*
128.3*
125.8*
121.3*
112.8
106.7
92.8
88.7
78.4
67.1
55.0
h
546.5*
468.4*
404.5*
349.6*
306.2
290.7
281.0
280.4
289.1
306.9
330.2
t
140.0*
138.2*
135.7*
130.9*
122.2*
115.8
107.1
97.4
87.0
75.3
62.7
h
589.1*
505.6*
437.8*
377.9*
331.3*
314.2
303.2
301.1
308.1
324.7
346.9
t
150.0*
148.1*
145.5*
140.6*
131.5*
125.0*
116.1*
106.2
95.5
83.5
70.3
h
632.2*
542.7*
470.5*
406.8*
356.6*
337.8*
325.5*
321.6
327.3
342.7
363.6
t
160.0*
158.1*
155.3*
150.3*
140.9*
134.2*
125.0*
115.0
104.1
91.8
78.9
h
675.6*
580.8*
503.1*
435.4*
381.9*
361.2*
347.7*
342.2
346.1
360.3
380.1
t
170.0*
168.0*
165.2*
159.9*
150.3*
143.3*
134.0*
123.7
112.7
100.0
85.7
h
719.2*
618.9*
536.1*
464.3*
406.8*
384.9*
369.9*
362.9
365.4
378.3
396.0
t
180.0*
177.9*
175.0*
169.6*
159.6*
152.5*
142.9*
132.5*
121.2*
108.2
93.3
h
763.2*
657.1*
569.4*
493.4*
432.1*
408.8*
392.1*
383.4*
384.3*
395.8
411.3
*Extensions of data above 115°C are well above the original data and should be used with care. #Supersaturated solution.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
30.72
Fig. 34 Enthalpy-Concentration Diagram for Water/Lithium Bromide Solutions
Licensed for single user. © 2017 ASHRAE, Inc.
30.73
Reprinted by permission of Carrier Corp
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
Fig. 35 Equilibrium Chart for Aqueous Lithium Bromide Solutions
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Licensed for single user. © 2017 ASHRAE, Inc.
30.74
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 36
Specific Density of Aqueous Solutions of Lithium Bromide
Fig. 37
Specific Heat of Aqueous Lithium Bromide Solutions
REFERENCES Tables added or revised for the 2017 edition [R-245fa, R1233zd(E), R-1234yf, and R-1234ze(E)] were calculated using Lemmon, E.W., M.L. Huber, and M.O. McLinden. 2016. NIST standard reference database 23, NIST reference fluid thermodynamic and transport properties—REFPROP, v. 9.2. Standard Reference Data Program, National Institute of Standards and Technology, Gaithersburg, MD.
The tables for R-125, R-170, R-290, R-600, and R-600a have not changed from the 2009 ASHRAE Handbook—Fundamentals; these tables are indicated with a * after the fluid name and were calculated using Lemmon, E.W., M.L. Huber, and M.O. McLinden. 2007. NIST standard reference database 23, NIST reference fluid thermodynamic and transport properties—REFPROP, v. 8.0. Standard Reference Data Program, National Institute of Standards and Technology, Gaithersburg, MD.
Many tables (indicated with a ‡ after the fluid name) have not changed from the 2005 ASHRAE Handbook—Fundamentals. These tables were calculated using Lemmon, E.W., M.O. McLinden, and M.L. Huber. 2002. NIST standard reference database 23, NIST reference fluid thermodynamic and transport properties—REFPROP, v. 7.0. Standard Reference Data Program, National Institute of Standards and Technology, Gaithersburg, MD.
Some tables (indicated with a † following the fluid name) have not changed from the 2001 ASHRAE Handbook—Fundamentals. These tables were calculated using McLinden, M.O., S.A. Klein, E.W. Lemmon, and A.P. Peskin. 2000a. NIST standard reference database 23: Thermodynamic and transport properties of refrigerants and refrigerant mixtures—REFPROP, v. 6.10. Standard Reference Data Program, National Institute of Standards and Technology, Gaithersburg, MD.
Fig. 38 Viscosity of Aqueous Solutions of Lithium Bromide
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Thermophysical Properties of Refrigerants
Licensed for single user. © 2017 ASHRAE, Inc.
Tables for R-704 have not changed from the 1997 ASHRAE Handbook—Fundamentals; these tables were calculated using the programs noted. The underlying sources for these computer packages are listed as follows by fluid and property. The reference listed under “Equation of state” was used for vapor pressure, liquid density, vapor volume, enthalpy, entropy, specific heat, and velocity of sound. R-12† Equation of state Marx, V., A. Pruss, and W. Wagner. 1992. Neue Zustandsgleichungen für R-12, R-22, R-11 und R-113. Beschreibung des thermodynamischen Zustandsverhaltens bei Temperaturen bis 525 K und Drücken bis 200 MPa. VDI-Fortschritt-Ber Wärmetechnik/Kältetechnik 19(57). VDI Verlag, Düsseldorf. Viscosity Klein, S.A., M.O. McLinden, and A. Laesecke. 1997. An improved extended corresponding states method for estimation of viscosity of pure refrigerants and mixtures. International Journal of Refrigeration 20:208217. Thermal conductivity McLinden, M.O., S.A. Klein, and R.A. Perkins. 2000b. An extended corresponding states model for the thermal conductivity of refrigerants and refrigerant mixtures. International Journal of Refrigeration 23:43-63. Surface tension Okada, M., and K. Watanabe. 1988. Surface tension correlations for several fluorocarbon refrigerants. Heat Transfer—Japanese Research 17:35-52. R-22† Equation of state Kamei, A., S.W. Beyerlein, and R.T. Jacobsen. 1995. Application of nonlinear regression in the development of a wide range formulation for HCFC-22. International Journal of Thermophysics 16(5):1155-1164. Viscosity Klein et al. 1997. op. cit. (See R-12.) Thermal conductivity McLinden et al. 2000b. op. cit. (See R-12.) Surface tension Okada, M., and K. Watanabe. 1988. op. cit. (See R-12.) R-23‡ Equation of state Penoncello, S.G., E.W. Lemmon, Z. Shan, and R.T. Jacobsen. 2003. An equation of state for the calculation of the thermodynamic properties of trifluoromethane (R-23). Journal of Physical and Chemical Reference Data 32:1473. Viscosity and thermal conductivity Shan, Z., S.G. Penoncello, and R.T. Jacobsen. 2000. A generalized model for viscosity and thermal conductivity of trifluoromethane (R-23). ASHRAE Transactions 106(1):757-767. Surface tension Penoncello, S.G. 1999. Thermophysical properties of trifluoromethane (R-23). ASHRAE Research Project RP-997, Final Report. R-32‡ Equation of state Tillner-Roth, R., and A. Yokozeki. 1997. An international standard equation of state for difluoromethane (R-32) for temperatures from the triple point at 136.34 K to 435 K and pressures up to 70 MPa. Journal of Physical and Chemical Reference Data 26:1273-1328. Viscosity Huber, M.L., A. Laesecke, and R.A. Perkins. 2003. Estimation of the viscosity and thermal conductivity of refrigerants including a new correlation for the viscosity of R134a. Industrial & Engineering Chemistry Research 42:3163-3178. Thermal conductivity Perkins, R.A. 2002. Personal communication, correlation to data as implemented in the NIST REFPROP Database. National Institute of Standards and Technology, Boulder, CO.
30.75 Surface tension Okada, M., and Y. Higashi. 1995. Experimental surface tensions for HFC-32, HCFC-124, HFC-125, HCFC-141b, HCFC-142b, and HFC152a. International Journal of Thermophysics 16(3):791-800. R-123† Equation of state Younglove, B.A., and M.O. McLinden. 1994. An international standard equation-of-state formulation of the thermodynamic properties of refrigerant 123 (2,2-dichloro-1,1,1-trifluoroethane). Journal of Physical and Chemical Reference Data 23(5):731-779. Viscosity Tanaka, Y., and T. Sotani. 1995. Thermodynamic and physical proper-ties, Chapter 2: Transport properties (thermal conductivity and viscos-ity), R-123. International Institute of Refrigeration, Paris. Thermal conductivity Laesecke, A., R.A. Perkins, and J.B. Howley. 1996. An improved correlation for the thermal conductivity of HCFC-123 (2,2-dichloro-1,1,1trifluoroethane). International Journal of Refrigeration 19:231-238. Surface tension Okada and Higashi. 1995. op. cit. (See R-32.) R-124‡ Equation of state de Vries, B., R. Tillner-Roth, and H.D. Baehr. 1995. Thermodynamic properties of HCFC-124. 19th International Congress of Refrigeration, International Institute of Refrigeration IVa:582-589. Viscosity and thermal conductivity Huber et al. 2003. op. cit. (See R-32.) Surface tension Okada and Higashi. 1995. op. cit. (See R-32.) R-125* Equation of state Lemmon, E.W., and R.T. Jacobsen. 2005. A new functional form and new fitting techniques for equations of state with application to pentafluoroethane (HFC-125). Journal of Physical and Chemical Reference Data 34(1):69-108. Viscosity Huber, M.L., and A. Laesecke. 2006. Correlation for the viscosity of pentafluoroethane (R-125) from the triple point to 500 K at pressures up to 60 MPa. Industrial & Engineering Chemistry Research 45(12):44474453. Thermal conductivity Perkins, R.A., and M.L. Huber. 2006. Measurement and correlation of the thermal conductivity of pentafluoroethane (R-125) from 190 K to 512 K at pressures to 70 MPa. Journal of Chemical & Engineering Data 51(3):898-904. Surface tension Okada and Higashi. 1995. op. cit. (See R-32.) R-134a† Equation of state Tillner-Roth, R., and H.D. Baehr. 1994. An international standard formulation of the thermodynamic properties of 1,1,1,2-tetrafluoroethane (HFC-134a) covering temperatures from 170 K to 455 K at pressures up to 70 MPa. Journal of Physical and Chemical Reference Data 23:657729. Viscosity Huber et al. 2003. op. cit. (See R-32.) Thermal conductivity Perkins, R.A., A. Laesecke, J. Howley, M.L.V. Ramires, A.N. Gurova, and L. Cusco. 2000. Experimental thermal conductivity values for the IUPAC round-robin sample of 1,1,1,2-tetrafluoroethane (R134a). NISTIR 6605. Surface tension Okada, M., and Y. Higashi. 1994. Surface tension correlation of HFC134a and HCFC-123. CFCs, the Day After: Proceedings of Joint Meeting of IIR Commissions B1, B2, E1, and E2, pp. 541-548. R-143a† Equation of state Lemmon, E.W., and R.T. Jacobsen. 2001. An international standard formulation for the thermodynamic properties of 1,1,1-trifluoroethane
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30.76
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(HFC- 143a) for temperatures from 161 to 500 K and pressures to 60 MPa. Journal of Physical and Chemical Reference Data 29(4):521-552. Viscosity Klein et al. 1997. op. cit. (See R-12.) Thermal conductivity McLinden et al. 2000b. op. cit. (See R-12.) Surface tension Schmidt, J.W., E. Carrillo-Nava, and M.R. Moldover. 1996. Partially halogenated hydrocarbons CHFCl-CF3, CF3-CH3, CF3-CHF-CHF2, CF3-CH2-CF3, CHF2-CF2-CH2F, CF3-CH2-CHF2, CF3-O-CHF2: Critical temperature, refractive indices, surface tension and estimates of liquid, vapor and critical densities. Fluid Phase Equilibria 122:187-206. R-152a‡ Equation of state Outcalt, S.L., and M.O. McLinden. 1996. A modified Benedict-WebbRubin equation of state for the thermodynamic properties of R-152a (1,1difluoroethane). Journal of Physical and Chemical Reference Data 25(2):605-636. Viscosity Klein et al. 1997. op. cit. (see R-12), ECS model of McLinden et al. (2000b) correlated to data as implemented in NIST REFPROP. Thermal conductivity Krauss, R., V.C. Weiss, T.A. Edison, J.V. Sengers, and K. Stephan. 1996. Transport properties of 1,1-difluoroethane (R-152a). International Journal of Thermophysics 17:731-757. Surface tension Okada and Higashi. 1995. op. cit. (See R-32.) R-245fa Equation of state Akasaka, R., Y. Zhou, and E.W. Lemmon. 2015. A fundamental equation of state for 1,1,1,3,3-pentafluoropropane (R-245fa). Journal of Physical and Chemical Reference Data 44:013104. Viscosity and thermal conductivity Perkins, R.A., M.L. Huber, and M.J. Assael. 2016. Measurements of the thermal conductivity of 1,1,1,3-3-pentafluoropropane (R-245fa) and correlations for the viscosity and thermal conductivity surfaces. Journal of Chemical and Engineering Data 61:3286-3294. Surface tension Mulero, A., I. Cachadiña, and M.I. Parra. 2015. Recommended correlations for the surface tension of common fluids. Journal of Physical and Chemical Reference Data 41:043105. R-1233zd(E) Equation of state Mondejar, M.E., M.O. McLinden, and E.W. Lemmon. 2015. Thermodynamic properties of trans-1-chloro-3,3,3-trifluoro-propene [R1233zd(E)]: Vapor pressure, p--T data, speed of sound measurements and equation of state. Journal of Chemical and Engineering Data 60: 2477-2489. Viscosity Hulse, R.J., R.S. Basu, R.R. Singh, and R.H.P. Thomas. 2012. Physical properties of HCFO-1233zd(E). Journal of Chemical and Engineering Data 57:3581-3586. Thermal conductivity Perkins, R.A., M.L. Huber, and M.J. Assael. 2017. Measurement and correlation of the thermal conductivity of trans-1-chloro-3,3,3-trifluoropropene (R1233zd(E)). Journal of Chemical and Engineering Data. dx.doi.org/10.1021/acs.jced.7b00106. Surface tension Kondou, C., R. Nagata, N. Nii, S. Koyama, and Y. Higashi. 2015 Surface tension of low GWP refrigerants R1243zf, R1234ze(Z), and R1233zd(E). International Journal of Refrigeration 53:80-89. R-1234yf Equation of state Richter, M., M.O. McLinden, and E.W. Lemmon. 2011. Thermodynamic properties of 2,3,3,3-tetrafluoroprop-1-ene (R-1234yf); p--T measurements and an equation of state. Journal of Chemical & Engineering Data 56:3254-3264. Viscosity Huber, M.L., and M.J. Assael. 2016. Correlations for the viscosity of 2,3,3,3-tetrafluoroprop-1-ene (R1234yf) and trans-1,2,2,2-tetrafluoropropene (R1234ze(E)). International Journal of Refrigeration 71:39-45.
2017 ASHRAE Handbook—Fundamentals (SI) Thermal conductivity Perkins, R.A., and M.L. Huber. 2011. Measurement and correlation of the thermal conductivity of 2,3,3,3-tetrafluoroprop-1-ene (R-1234fy) and trans-1,3,3,3-tetrafluoropropene (R-1234ze). Journal of Chemical & Engineering Data 56:4868-4874. Surface tension Mulero et al. 2015. op. cit. (See R-245fa). R-1234ze(E) Equation of state Thol, M., and E.W. Lemmon. 2016. Equation of state for the thermodynamic properties of trans-1,3,3,3-Tetrafluoropropene [R1234ze(E)]. International Journal of Thermophysics 37:28. Viscosity Huber and Assael. op. cit. (See R-1234yf.) Thermal conductivity Perkins and Huber. 2011. op. cit. (See R-1234yf). Surface tension Mulero et al. 2015. op. cit. (See R-245fa). R-404A‡ Equation of state Lemmon, E.W., and R.T. Jacobsen. 2004a. Equations of state for mixtures of R-32, R-125, R-134a, R-143a, and R-152a. Journal of Physical and Chemical Reference Data 33(2):593-620. Viscosity Klein et al. 1997. op. cit. (See R-12.) Thermal conductivity McLinden et al. 2000b. op. cit. (See R-12.) Surface tension Moldover, M.R., and J.C. Rainwater. 1988. Interfacial tension and vaporliquid equilibria in the critical region of mixtures. Journal of Chemical Physics 88:7772-7780. R-407C‡ Equation of state Lemmon and Jacobsen. 2004a. op. cit. (See R-404A.) Viscosity Klein et al. 1997. op. cit. (See R-12.) Thermal conductivity McLinden et al. 2000b. op. cit. (See R-12.) Surface tension Moldover and Rainwater. 1988. op. cit. (See R-404A.) R-410A‡ Equation of state Lemmon and Jacobsen. 2004a. op. cit. (See R-404A.) Viscosity Klein et al. 1997. op. cit. (See R-12.) Thermal conductivity McLinden et al. 2000b. op. cit. (See R-12.) Surface tension Moldover and Rainwater. 1988. op. cit. (See R-404A.) R-507A‡ Equation of state Lemmon and Jacobsen. 2004a. op. cit. (See R-404A.) Viscosity Klein et al. 1997. op. cit. (See R-12.) Thermal conductivity McLinden et al. 2000b. op. cit. (See R-12.) Surface tension Moldover and Rainwater. 1988. op. cit. (See R-404A.) R-717 (Ammonia)† Equation of state Tillner-Roth, R., F. Harms-Watzenberg, and H.D. Baehr. 1993. Eine neue Fundamentalgleichung für Ammoniak. DKV-Tagungsbericht 20(II):167181. Viscosity Fenghour, A., W.A. Wakeham, V. Vesovic, J.T.R. Watson, J. Millat, and E. Vogel. 1995a. The viscosity of ammonia. Journal of Physical and Chemical Reference Data 24:1649-1667.
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Thermophysical Properties of Refrigerants
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Thermal conductivity Tufeu, R., D.Y. Ivanov, Y. Garrabos, and B. Le Neindre. 1984. Thermal conductivity of ammonia in a large temperature and pressure range including the critical region. Berichte der Bunsen-Gesellschaft—Physical Chemistry 88:422-427. Surface tension Stairs, R.A., and M.J. Sienko. 1956. Surface tension of ammonia and of solutions of alkalai halides in ammonia. Journal of American Chemical Society 78:920-923. R-718 (Water/Steam)† Data computed using Harvey, A.H., S.A. Klein, and A.P. Peskin. 1999. NIST Standard Reference Database 10. NIST/ASME steam properties database, v. 2.2. Standard Reference Data Program. Equation of state Wagner, W., and A. Pruss. 2002. The IAPWS formulation 1995 for the thermodynamic properties of ordinary water substance for general and scientific use. Journal of Physical and Chemical Reference Data 31:387535. Viscosity and thermal conductivity Kestin, J., J.V. Sengers, B. Kamgar-Parsi, and J.M.H. Levelt Sengers. 1984. Thermophysical properties of fluid H2O. Journal of Physical and Chemical Reference Data 13:175. Surface tension IAPWS. 1995. Physical chemistry of aqueous systems: Meeting the needs of industry. Proceedings of the 12th International Conference on the Properties of Water and Steam, Orlando. Begell House, Inc., A139A142. International Association for the Properties of Steam. R-744 (Carbon Dioxide)† Equation of state Span, R., and W. Wagner. 1996. A new equation of state for carbon dioxide covering the fluid region from the triple-point temperature to 1100 K at pressures up to 800 MPa. Journal of Physical and Chemical Reference Data 26:1509-1596. Viscosity Fenghour, A., W.A. Wakeham, and V. Vesovic. 1995b. The viscosity of carbon dioxide. Journal of Physical and Chemical Reference Data 27: 31-44. Thermal conductivity Vesovic, V., W.A. Wakeham, G.A. Olchowy, J.V. Sengers, J.T.R. Watson, and J. Millat. 1990. The transport properties of carbon dioxide. Journal of Physical and Chemical Reference Data 19:763-808. Surface tension Rathjen, W., and J. Straub. 1977. Heat transfer in boiling, Chapter 18, Temperature dependence of surface tension, coexistence curve, and vapor pressure of CO2, CClF3, CBrF3, and SF6. Academic Press, New York. R-50 (Methane)† Equation of state Setzmann, U., and W. Wagner. 1991. A new equation of state and tables of thermodynamic properties for methane covering the range from the melting line to 625 K at pressures to 1000 MPa. Journal of Physical and Chemical Reference Data 20:1061-1151. Viscosity Younglove, B.A., and J.F. Ely. 1987. Thermophysical properties of fluids. II. Methane, ethane, propane, isobutane and normal butane. Journal of Physical and Chemical Reference Data 16:577-798. Thermal conductivity Friend, D.G., J.F. Ely, and H. Ingham. 1989. Thermophysical properties of methane. Journal of Physical and Chemical Reference Data 18(2): 583-638. Surface tension Somayajulu, G.R. 1988. A generalized equation for surface tension from the triple point to the critical point. International Journal of Thermophysics 9:559-566. R-170 (Ethane)* Equation of state Bücker, D., and W. Wagner. 2006. A reference equation of state for the thermodynamic properties of ethane for temperatures from the melting
30.77 line to 675 K and pressures up to 900 MPa. Journal of Physical and Chemical Reference Data 35:205. Viscosity, and thermal conductivity Friend, D.G., H. Ingham, and J.F. Ely. 1991. Thermophysical properties of ethane. Journal of Physical and Chemical Reference Data 20(2): 275-347. Surface tension Soares, V.A.M., B.d.J.V.S. Almeida, I.A. McLure, and R.A. Higgins. 1986. Surface tension of pure and mixed simple substances at low temperature. Fluid Phase Equilibria 32:9-16. Pressure-enthalpy diagram based on data of Friend, D.G., H. Ingham, and J.F. Ely. 1991. Thermophysical properties of ethane. Journal of Physical and Chemical Reference Data 20(2): 275-347. R-290 (Propane)* Equation of state Lemmon, E.W., W. Wagner, and M.O. McLinden. 2009. Thermodynamic properties of propane, III: Equation of state. Journal of Chemical & Engineering Data 54:3141-3180. Viscosity Vogel, E., C. Küchenmeister, E. Bich, and A. Laesecke. 1998. Reference correlation of the viscosity of propane. Journal of Physical and Chemical Reference Data 27:947-970. Thermal conductivity Marsh, K., R. Perkins, and M.L.V. Ramires. 2002. Measurement and correlation of the thermal conductivity of propane from 86 to 600 K at pressures to 70 MPa. Journal of Chemical & Engineering Data 47:932-940. Surface tension Baidakov, V.G., and I.I. Sulla. 1985. Surface tension of propane and isobutane at near-critical temperatures. Russian Journal of Physical Chemistry 59:551-554. Pressure-enthalpy diagram based on data of Miyamoto, H., and K. Watanabe. 2000. A thermodynamic property model for fluid-phase propane. International Journal of Thermophysics 21:1045-1072. R-600 (n-Butane)* Equation of state Bücker, D., and W. Wagner. 2006. Reference equations of state for the thermodynamic properties of fluid phase n-butane and isobutane. Journal of Physical and Chemical Reference Data 35:929. Viscosity Vogel, E., C. Kuchenmeister, and E. Bich. 1999. Viscosity for n-butane in the fluid region. High Temperatures—High Pressures 31:173-186. Thermal conductivity Perkins, R.A., M.L.V. Ramires, C.A. Nieto de Castro, and L. Cusco. 2002. Measurement and correlation of the thermal conductivity of butane. Journal of Chemical and Engineering Data 47:1263-1271. Surface tension Calado, J.C.G., I.A. McLure, and V.A.M. Soares. 1978. Surface tension for octafluorocyclobutane, n-butane and their mixtures from 233 K to 254 K, and vapour pressure, excess Gibbs function and excess volume for the mixture at 233 K. Fluid Phase Equilibria 2:199-213. Coffin, C.C., and O. Maass. 1928. The preparation and physical properties of -, - and -butylene and normal and isobutane. Journal of the American Chemical Society 50(5):1427-1437. Pressure-enthalpy diagram based on data of Miyamoto, H., and K. Watanabe. 2001. Thermodynamic property model for fluid-phase n-butane. International Journal of Thermophysics 22:459-475. R-600a (Isobutane)* Equation of state Bücker and Wagner. 2006. op. cit. (See R-600.) Viscosity Vogel, E., C. Küchenmeister, and E. Bich. 2000. Viscosity correlation for isobutane over wide ranges of the fluid region. International Journal of Thermophysics 21:343-356.
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Thermal conductivity Perkins, R.A. 2002. Measurement and correlation of the thermal conductivity of isobutane. Journal of Chemical Engineering Data 47:12721279. Surface tension Baidakov and Sulla. 1985. op. cit. (See R-290.) Pressure-enthalpy diagram based on data of Miyamoto, H., and K. Watanabe. 2002. A thermodynamic property model for fluid-phase isobutane. International Journal of Thermophysics 23:477-499.
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R-1150 (Ethylene)† Equation of state Smukala, J., R. Span, and W. Wagner. 2000. A new equation of state for ethylene covering the fluid region for temperatures from the melting line to 450 K and pressures up to 300 MPa. Journal of Physical and Chemical Reference Data 29:1053-1122. Viscosity and thermal conductivity Holland, P.M., B.E. Eaton, and H.J.M. Hanley. 1983. A correlation of the viscosity and thermal conductivity data of gaseous and liquid ethylene. Journal of Physical and Chemical Reference Data 12:917-932. Surface tension Soares et al. op. cit. (See R-170.) R-1270 (Propylene)‡ Equation of state Angus, S., B. Armstrong, and K.M. de Reuck. 1980. International thermodynamic tables of the fluid state—7: Propylene. Pergamon Press, Oxford, U.K. Viscosity and thermal conductivity Huber et al. 2003. op. cit. (See R-32.) Surface tension Maass, O., and C.H. Wright. 1921. Some physical properties of hydrocarbons containing two and three carbon atoms. Journal of the American Chemical Society 43:1098-1111. R-704 (Helium) Thermodynamic data computed using the ALLPROPS database, v. 4.0: Lemmon, E.W., R.T. Jacobsen, S.G. Penoncello, and S.W. Beyerlein. 1994. Computer programs for the calculation of thermodynamic properties of cryogenic and other fluids. Advances in Cryogenic Engineering 39:1891-1897. Transport data computed using the NIST 12 database, v. 3.0: Friend, D.G., R.D. McCarty, and V. Arp. 1992. NIST thermophysical properties of pure fluids database, v. 3.0. Standard Reference Data Program. Equation of state, viscosity, and thermal conductivity Arp, V.D., R.D. McCarty, and D.G. Friend. 1995. Thermophysical properties of helium-4 from 0.8 to 1500 K with pressures to 2000 MPa. NIST Technical Note 1334 (revised).
Surface tension Liley, P.E., and P.D. Desai. 1993. ASHRAE thermophysical properties of refrigerants. R-728 (Nitrogen)‡ Equation of state, viscosity, and thermal conductivity Span, R., E.W. Lemmon, R.T. Jacobsen, W. Wagner, and A. Yokozeki. 2000. A reference equation of state for the thermodynamic properties of nitrogen for temperatures from 63.151 to 1000 K and pressures to 2200 MPa. Journal of Physical and Chemical Reference Data 29:13611433. Viscosity and thermal conductivity Lemmon, E.W., and R.T. Jacobsen. 2004b. Viscosity and thermal conductivity equations for nitrogen, oxygen, argon, and air. International Journal of Thermophysics 25:21-69. Surface tension Lemmon, E.W., and S.G. Penoncello. 1994. The surface tension of air and air component mixtures. Advances in Cryogenic Engineering 39:1927-1934. R-729 (Air)‡ Equation of state Lemmon, E.W., R.T. Jacobsen, S.G. Penoncello, and D.G. Friend. 2000. Thermodynamic properties of air and mixtures of nitrogen, argon, and oxygen from 60 to 2000 K at pressures to 2000 MPa. Journal of Physical and Chemical Reference Data 29:331-385. Viscosity and thermal conductivity Lemmon and Jacobsen. 2004b. op. cit. (See R-728.) Surface tension Lemmon and Penoncello. 1994. op. cit. (See R-728.) R-732 (Oxygen)‡ Equation of state Schmidt, R., and W. Wagner. 1985. A new form of the equation of state for pure substances and its application to oxygen. Fluid Phase Equilibria 19:175-200. Viscosity and thermal conductivity Lemmon and Jacobsen. 2004b. op. cit. (See R-728) Surface tension Lemmon and Penoncello. 1994. op. cit. (See R-728.) R-740 (Argon)‡ Equation of state Tegeler, C., R. Span, and W. Wagner. 1999. A new equation of state for argon covering the fluid region for temperatures from the melting line to 700 K at pressures up to 1000 MPa. Journal of Physical and Chemical Reference Data 28:779-850. Viscosity and thermal conductivity Lemmon and Jacobsen. 2004b. op. cit. (See R-728.) Surface tension Lemmon and Penoncello. 1994. op. cit. (See R-728.)
Related Commercial Resources
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Related Commercial Resources CHAPTER 31
PHYSICAL PROPERTIES OF SECONDARY COOLANTS (BRINES) Salt-Based Brines .......................................................................................................................... 31.1 Inhibited Glycols ........................................................................................................................... 31.4 Halocarbons................................................................................................................................ 31.12 Nonhalocarbon, Nonaqueous Fluids .......................................................................................... 31.12
N many refrigeration applications, heat is transferred to a secondary coolant, which can be any liquid cooled by the refrigerant and used to transfer heat without changing state. These liquids are also known as heat transfer fluids, brines, or secondary refrigerants. Other ASHRAE Handbook volumes describe various applications for secondary coolants. In the 2014 ASHRAE Handbook— Refrigeration, refrigeration systems are discussed in Chapter 13, their uses in food processing in Chapters 23 and 28 to 42, and ice rinks in Chapter 44. In the 2015 ASHRAE Handbook—HVAC Applications, solar energy use is discussed in Chapter 35, and snow melting and freeze protection in Chapter 51. Thermal storage is covered in Chapter 51 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment. This chapter describes physical properties of the more common secondary coolants based on ethylene glycol, propylene glycol, sodium chloride, or calcium chloride and provides information on their use. Less widely used secondary coolants such as ethyl alcohol or potassium formate are not included in this chapter, but their physical properties are summarized in Melinder (2007). Physical
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I
property data for nitrate and nitrite salt solutions used for stratified thermal energy storage are presented by Andrepont (2012). The chapter also includes information on corrosion protection. Supplemental information on corrosion inhibition can be found in Chapter 49 of the 2015 ASHRAE Handbook—HVAC Applications and Chapter 13 of the 2014 ASHRAE Handbook—Refrigeration.
1.
SALT-BASED BRINES
Physical Properties Water solutions of calcium chloride and sodium chloride have historically been the most common refrigeration brines. Tables 1 and 2 list the properties of pure calcium chloride brine and sodium chloride brine. For commercial grades, use the formulas in the footnotes to these tables. For calcium chloride brines, Figure 1 shows specific heat, Figure 2 shows the ratio of mass of solution to that of water, Figure 3 shows viscosity, and Figure 4 shows thermal conductivity. Figures 5 to 8 show the same properties for sodium chloride brines.
Table 1 Properties of Pure Calcium Chloride* Brines Pure CaCl2, % Specific Heat at by Mass 15°C, J/(kg·K) 0 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 29.87 30 32 34
4184 3866 3824 3757 3699 3636 3577 3523 3464 3414 3364 3318 3259 3209 3163 3121 3084 3050 2996 2958 2916 2882 2853 2816 2782 2753 2741 2732 2678 2636
Crystallization Starts, °C 0.0 –2.4 –2.9 –3.4 –4.1 –4.7 –5.4 –6.2 –7.1 –8.0 –9.2 –10.3 –11.6 –13.0 –14.5 –16.2 –18.0 –19.9 –22.1 –24.4 –26.8 –29.4 –32.1 –35.1 –38.8 –45.2 –55.0 –46.0 –28.6 –15.4
Density at 16°C, kg/m3 CaCl2
Brine
0.0 52.2 63.0 74.2 85.5 96.9 108.6 120.5 132.5 144.8 157.1 169.8 182.6 195.7 209.0 222.7 236.0 249.6 264.3 278.7 293.5 308.2 323.1 338.5 354.0 369.9 378.8 358.4 418.1 452.0
999 1044 1049 1059 1068 1078 1087 1095 1104 1113 1123 1132 1141 1152 1161 1171 1180 1189 1201 1211 1223 1232 1242 1253 1264 1275 1289 1294 1316 1339
Density at Various Temperatures, kg/m3
–20°C
–10°C
0°C
10°C
1140 1150 1160 1170 1179 1189
1042 1051 1060 1070 1079 1088 1097 1107 1116 1126 1136 1145 1155 1165 1175 1185
1041 1050 1059 1068 1077 1086 1095 1104 1114 1123 1133 1142 1152 1162 1172 1182
1214
1210
1206
1202
1235
1231
1227
1223
Source: CCI (1953) *Mass of Type 1 (77% min.) CaCl2 = (mass of pure CaCl2)/(0.77). Mass of Type 2 (94% min.) CaCl2 = (mass of pure CaCl2)/(0.94).
________ The preparation of this chapter is assigned to TC 3.1, Refrigerants and Secondary Coolants.
31.1 Copyright © 2017, ASHRAE
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31.2
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 3 Viscosity of Calcium Chloride Brines (CCI 1953)
Fig. 1
Specific Heat of Calcium Chloride Brines (CCI 1953)
Fig. 2 Density of Calcium Chloride Brines (CCI 1953)
Fig. 4
Thermal Conductivity of Calcium Chloride Brines (CCI 1953)
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Physical Properties of Secondary Coolants (Brines) Table 2
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Pure NaCl, % by Mass 0 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 25.2 aMass
31.3
Properties of Pure Sodium Chloridea Brines
Specific Heat at 15°C, J/(kg·K)
Crystallization Starts, °C
4184 3925 3879 3836 3795 3753 3715 3678 3640 3607 3573 3544 3515 3485 3456 3427 3402 3376 3356 3330 3310 3289
0.0 –2.9 –3.6 –4.3 –5.0 –5.8 –6.6 –7.3 –8.2 –9.1 –10.1 –10.9 –11.9 –13.0 –14.1 –15.3 –16.5 –17.8 –19.1 –20.6 –15.7 –8.8 0.0
Density at 16°C, kg/m3 NaCl
Brine
0.0 51.7 62.5 73.4 84.6 95.9 107.2 118.8 130.3 142.2 154.3 166.5 178.9 191.4 204.1 217.0 230.0 243.2 256.6 270.0 283.7 297.5
1000 1035 1043 1049 1057 1065 1072 1080 1086 1094 1102 1110 1118 1126 1134 1142 1150 1158 1166 1174 1182 1190
Density at Various Temperatures, kg/m3 –10°C
–0°C
10°C
20°C
1119.4 1127.6 1135.8 1144.1 1153.4 1160.7 1169.1 1177.6 1186.1 1194.7
1038.1 1045.8 1053.7 1061.2 1069.0 1076.8 1084.8 1092.4 1100.3 1108.2 1116.2 1124.2 1132.2 1140.3 1148.5 1156.7 1165.0 1173.3 1181.7 1190.1
1036.5 1043.9 1051.4 1058.9 1066.4 1074.0 1081.6 1089.6 1097.0 1104.7 1112.5 1120.4 1128.3 1136.2 1144.3 1154.1 1160.5 1168.7 1177.0 1185.3
1034.0 1041.2 1048.5 1055.8 1063.2 1070.6 1078.1 1085.6 1093.2 1100.8 1108.5 1116.2 1124.0 1131.8 1139.7 1147.7 1155.8 1163.9 1172.0 1180.3
of commercial NaC1 required = (mass of pure NaCl required)/(% purity).
Fig. 5
Specific Heat of Sodium Chloride Brines (adapted from Carrier 1959)
Fig. 6
Density of Sodium Chloride Brines (adapted from Carrier 1959)
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31.4
2017 ASHRAE Handbook—Fundamentals (SI)
Brine applications in refrigeration are mainly in industrial machinery and in skating rinks. Corrosion is the principal problem for calcium chloride brines, especially in ice-making tanks where galvanized iron cans are immersed. Ordinary salt (sodium chloride) is used where contact with calcium chloride is intolerable (e.g., the brine fog method of freezing fish and other foods). It is used as a spray to air-cool unit coolers to prevent frost formation on coils. In most refrigerating work, the lower freezing point of calcium chloride solution makes it more convenient to use. Commercial calcium chloride, available as Type 1 (77% minimum) and Type 2 (94% minimum), is marketed in flake, solid, and solution forms; flake form is used most extensively. Commercial sodium chloride is available both in crude (rock salt) and refined grades. Because magnesium salts tend to form sludge, their presence in sodium or calcium chloride is undesirable.
Corrosion Inhibition All brine systems must be treated to control corrosion and deposits. Historically, chloride-based brines were maintained at neutral pH and treated with sodium chromate. However, using chromate as
a corrosion inhibitor is no longer deemed acceptable because of its detrimental environmental effects. Chromate has been placed on hazardous substance lists by several regulatory agencies. For example, the U.S. Agency for Toxic Substances and Disease Registry’s (ATSDR 2016) Priority List of Hazardous Substances ranks hexavalent chromium 17th out of 275 chemicals of concern (based on frequency, toxicity, and potential for human exposure at National Priorities List facilities). Consequently, hexavalent chrome and several chromates are also listed on several state right-to-know hazardous substance lists, including New Jersey, California, Minnesota, Pennsylvania and others. Instead of chromate, most brines use a sodium-nitrite-based inhibitor ranging from approximately 3000 mg/kg in calcium brines to 4000 mg/kg in sodium brines. Other, proprietary organic inhibitors are also available to mitigate the inherent corrosiveness of brines. Before using any inhibitor package, review federal, state, and local regulations concerning the use and disposal of the spent fluids. If the regulations prove too restrictive, an alternative inhibition system should be considered.
2.
INHIBITED GLYCOLS
Licensed for single user. © 2017 ASHRAE, Inc.
Ethylene glycol and propylene glycol, when properly inhibited for corrosion control, are used as aqueous-freezing-point depressants (antifreeze) and heat transfer media. Their chief attributes are their ability to efficiently lower the freezing point of water, their low volatility, and their relatively low corrosivity when properly inhibited. Inhibited ethylene glycol solutions have better thermophysical properties than propylene glycol solutions, especially at lower temperatures. However, the less toxic propylene glycol is preferred for applications involving possible human contact or where mandated by regulations. If a heat transfer fluid may have incidental food contact, then it should be made from propylene glycol that meets U.S. Pharmacopeia (USP 2016) Food Chemical Codex (FCC) specifications. Avoid other, less pure grades of propylene glycol: they can contain toxic or unwanted impurities that also adversely affect performance characteristics (e.g., foaming propensity, corrosion).
Physical Properties
Fig. 7 Viscosity of Sodium Chloride Brines (adapted from Carrier 1959)
Fig. 8 Thermal Conductivity of Sodium Chloride Brines (adapted from Carrier 1959)
Ethylene glycol and propylene glycol are colorless, practically odorless liquids that are miscible with water and many organic compounds. Table 3 shows properties of the pure materials. The freezing and boiling points of aqueous solutions of ethylene glycol and propylene glycol are given in Tables 4 and 5. Note that increasing the concentration of ethylene glycol above 60% by mass causes the freezing point of the solution to increase. Propylene glycol solutions above 60% by mass do not have freezing points. Instead of freezing, propylene glycol solutions supercool and become a glass (a liquid with extremely high viscosity and the appearance and properties of a noncrystalline amorphous solid). On the dilute side of the eutectic (the mixture at which freezing produces a solid phase of the same composition), ice forms on freezing; on the concentrated side, solid glycol separates from solution on freezing. The freezing rate of such solutions is often quite slow, but, in time, they set to a hard, solid mass. Physical properties (i.e., density, specific heat, thermal conductivity, and viscosity) for aqueous solutions of ethylene glycol can be found in Tables 6 to 9 and Figures 9 to 12; similar data for aqueous solutions of propylene glycol are in Tables 10 to 13 and Figures 13 to 16. Densities are for aqueous solutions of industrially inhibited glycols, and are somewhat higher than those for pure glycol and water alone. Typical corrosion inhibitor packages do not significantly affect other physical properties. Physical properties for the two fluids are similar, except for viscosity. At the same concentration, aqueous solutions of propylene glycol are more viscous than solutions of ethylene glycol. This higher viscosity accounts for the majority of the performance difference between the two fluids.
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Physical Properties of Secondary Coolants (Brines)
Fig. 9
Density of Aqueous Solutions of Industrially Inhibited Ethylene Glycol (vol. %)
Licensed for single user. © 2017 ASHRAE, Inc.
(Dow Chemical 2001b)
31.5
Fig. 10 Specific Heat of Aqueous Solutions of Industrially Inhibited Ethylene Glycol (vol. %) (Dow Chemical 2001b)
Table 3 Physical Properties of Ethylene Glycol and Propylene Glycol Property Relative molecular mass Density at 20°C, kg/m3 Boiling point, °C at 101.3 kPa at 6.67 kPa at 1.33 kPa Vapor pressure at 20°C, Pa Freezing point, °C Viscosity, mPa·s at 0°C at 20°C at 40°C Refractive index nD at 20°C Specific heat at 20°C, kJ/(kg·K) Heat of fusion at –12.7°C, kJ/kg Heat of vaporization at 101.3 kPa, kJ/kg Heat of combustion at 20°C, MJ/kg
Ethylene Glycol
Propylene Glycol
62.07 1113
76.10 1036
198 123 89 6.7 –12.7
187 116 85 9.3 Sets to glass below –51°C
57.4 20.9 9.5 1.4319 2.347 187 846 19.246
243 60.5 18.0 1.4329 2.481 — 688 23.969
Sources: Dow Chemical (2001a, 2001b)
The choice of glycol concentration depends on the type of protection required by the application. If the fluid is being used to prevent equipment damage during idle periods in cold weather, such as winterizing coils in an HVAC system, 30% by volume ethylene glycol or 35% by volume propylene glycol is sufficient. These concentrations allow the fluid to freeze. As the fluid freezes, it forms a slush that expands and flows into any available space. Therefore, expansion volume must be included with this type of protection. If the application requires that the fluid remain entirely liquid, use a concentration with a freezing point 3 K below the lowest expected temperature. Avoid excessive glycol concentration because it increases initial cost and adversely affects the fluid’s physical properties. Additional physical property data are available from suppliers of industrially inhibited ethylene and propylene glycol.
Fig. 11 Thermal Conductivity of Aqueous Solutions of Industrially Inhibited Ethylene Glycol (vol. %) (Dow Chemical 2001b)
Corrosion Inhibition Interestingly, ethylene glycol and propylene glycol, when not diluted with water, are actually less corrosive than water is with common construction metals. However, once diluted with water (as is typical), all aqueous glycol solutions are more corrosive than the water from which they are prepared: when uninhibited glycols thermally degrade and oxidize with use, they form acidic degradation products, which create an increasingly more corrosive environment if corrosion inhibitors and pH buffering compounds are not present. The amount of oxidation is influenced by temperature, degree of aeration, and type of metal components to which the glycol solution is exposed. In general, hydronic heating systems cause more degradation of glycol-based heat transfer fluids than do chilled-water systems. It is therefore necessary for glycol-based heat transfer fluids not only to use corrosion inhibitors that are effective at protecting common metals from corrosion, but also to contain additional additives to buffer or neutralize the acidic glycol degradation products that form during use. Corrosion inhibitors form a surface barrier that protects metal from attack, but their effectiveness is highly dependent on solution pH. Failure to compensate for glycol degradation
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31.6
2017 ASHRAE Handbook—Fundamentals (SI)
Table 4
Freezing and Boiling Points of Aqueous Solutions of Ethylene Glycol
Licensed for single user. © 2017 ASHRAE, Inc.
Percent Ethylene Glycol By Mass
By Volume
Freezing Point, °C
0.0 5.0 10.0 15.0 20.0 21.0 22.0 23.0 24.0 25.0 26.0 27.0 28.0 29.0 30.0 31.0 32.0 33.0 34.0 35.0 36.0 37.0 38.0 39.0 40.0 41.0 42.0 43.0 44.0 45.0 46.0 47.0 48.0 49.0 50.0 51.0 52.0 53.0 54.0 55.0 56.0 57.0 58.0 59.0 60.0 65.0 70.0 75.0 80.0 85.0 90.0 95.0
0.0 4.4 8.9 13.6 18.1 19.2 20.1 21.0 22.0 22.9 23.9 24.8 25.8 26.7 27.7 28.7 29.6 30.6 31.6 32.6 33.5 34.5 35.5 36.5 37.5 38.5 39.5 40.5 41.5 42.5 43.5 44.5 45.5 46.6 47.6 48.6 49.6 50.6 51.6 52.7 53.7 54.7 55.7 56.8 57.8 62.8 68.3 73.6 78.9 84.3 89.7 95.0
0.0 –1.4 –3.2 –5.4 –7.8 –8.4 –8.9 –9.5 –10.2 –10.7 –11.4 –12.0 –12.7 –13.3 –14.1 –14.8 –15.4 –16.2 –17.0 –17.9 –18.6 –19.4 –20.3 –21.3 –22.3 –23.2 –24.3 –25.3 –26.4 –27.5 –28.8 –29.8 –31.1 –32.6 –33.8 –35.1 –36.4 –37.9 –39.3 –41.1 –42.6 –44.2 –45.6 –47.1 –48.3 * * * –46.8 –36.9 –29.8 –19.4
Source: Dow Chemical (2001b) *Freezing points are below –50°C.
Table 5
Freezing and Boiling Points of Aqueous Solutions of Propylene Glycol
Percent Propylene Glycol
Boiling Point, °C at 100.7 kPa
By Mass
By Volume
Freezing Point, °C
Boiling Point, °C at 100.7 kPa
100.0 100.6 101.1 101.7 102.2 102.2 102.2 102.8 102.8 103.3 103.3 103.3 103.9 103.9 104.4 104.4 104.4 104.4 104.4 105.0 105.0 105.0 105.0 105.0 105.6 105.6 105.6 106.1 106.1 106.7 106.7 106.7 106.7 106.7 107.2 107.2 107.2 107.8 107.8 108.3 108.3 108.9 108.9 109.4 110.0 112.8 116.7 120.0 123.9 133.9 140.6 158.3
0.0 5.0 10.0 15.0 20.0 21.0 22.0 23.0 24.0 25.0 26.0 27.0 28.0 29.0 30.0 31.0 32.0 33.0 34.0 35.0 36.0 37.0 38.0 39.0 40.0 41.0 42.0 43.0 44.0 45.0 46.0 47.0 48.0 49.0 50.0 51.0 52.0 53.0 54.0 55.0 56.0 57.0 58.0 59.0 60.0 65.0 70.0 75.0 80.0 85.0 90.0 95.0
0.0 4.8 9.6 14.5 19.4 20.4 21.4 22.4 23.4 24.4 25.3 26.4 27.4 28.4 29.4 30.4 31.4 32.4 33.5 34.4 35.5 36.5 37.5 38.5 39.6 40.6 41.6 42.6 43.7 44.7 45.7 46.8 47.8 48.9 49.9 50.9 51.9 53.0 54.0 55.0 56.0 57.0 58.0 59.0 60.0 65.0 70.0 75.0 80.0 85.0 90.0 95.0
0.0 –1.6 –3.3 –5.1 –7.1 –7.6 –8.0 –8.6 –9.1 –9.6 –10.2 –10.8 –11.4 –12.0 –12.7 –13.4 –14.1 –14.8 –15.6 –16.4 –17.3 –18.2 –19.1 –20.1 –21.1 –22.1 –23.2 –24.3 –25.5 –26.7 –27.9 –29.3 –30.6 –32.1 –33.5 –35.0 –36.6 –38.2 –39.8 –41.6 –43.3 –45.2 –47.1 –49.0 –51.1 * * * * * * *
100.0 100.0 100.0 100.0 100.6 100.6 100.6 100.6 100.6 101.1 101.1 101.1 101.7 101.7 102.2 102.2 102.2 102.2 102.2 102.8 102.8 102.8 103.3 103.3 103.9 103.9 103.9 103.9 103.9 104.4 104.4 104.4 105.0 105.0 105.6 105.6 105.6 106.1 106.1 106.1 106.1 106.7 106.7 106.7 107.2 108.3 110.0 113.9 118.3 125.0 132.2 154.4
Source: Dow Chemical (2001a) *Above 60% by mass, solutions do not freeze but become a glass.
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Physical Properties of Secondary Coolants (Brines)
31.7
Table 6 Density of Aqueous Solutions of Ethylene Glycol Concentrations in Volume Percent Ethylene Glycol
Licensed for single user. © 2017 ASHRAE, Inc.
Temperature, °C
10%
–35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125
1018.73 1017.57 1016.28 1014.87 1013.34 1011.69 1009.92 1008.02 1006.01 1003.87 1001.61 999.23 996.72 994.10 991.35 988.49 985.50 982.39 979.15 975.80 972.32 968.73 965.01 961.17 957.21 953.12
20%
1036.85 1035.67 1034.36 1032.94 1031.39 1029.72 1027.93 1026.02 1023.99 1021.83 1019.55 1017.16 1014.64 1011.99 1009.23 1006.35 1003.34 1000.21 996.96 993.59 990.10 986.48 982.75 978.89 974.91 970.81 966.59
30%
1054.31 1053.11 1051.78 1050.33 1048.76 1047.07 1045.25 1043.32 1041.26 1039.08 1036.78 1034.36 1031.81 1029.15 1026.36 1023.45 1020.42 1017.27 1014.00 1010.60 1007.09 1003.45 999.69 995.81 991.81 987.68 983.43 979.07
40%
50%
60%
70%
80%
90%
1071.98 1070.87 1069.63 1068.28 1066.80 1065.21 1063.49 1061.65 1059.68 1057.60 1055.39 1053.07 1050.62 1048.05 1045.35 1042.54 1039.61 1036.55 1033.37 1030.07 1026.65 1023.10 1019.44 1015.65 1011.74 1007.71 1003.56 999.29 994.90 990.38
1089.94 1089.04 1088.01 1086.87 1085.61 1084.22 1082.71 1081.08 1079.33 1077.46 1075.46 1073.35 1071.11 1068.75 1066.27 1063.66 1060.94 1058.09 1055.13 1052.04 1048.83 1045.49 1042.04 1038.46 1034.77 1030.95 1027.01 1022.95 1018.76 1014.46 1010.03 1005.48 1000.81
1104.60 1103.54 1102.36 1101.06 1099.64 1098.09 1096.43 1094.64 1092.73 1090.70 1088.54 1086.27 1083.87 1081.35 1078.71 1075.95 1073.07 1070.06 1066.94 1063.69 1060.32 1056.83 1053.22 1049.48 1045.63 1041.65 1037.55 1033.33 1028.99 1024.52 1019.94 1015.23 1010.40
1118.61 1117.38 1116.04 1114.58 1112.99 1111.28 1109.45 1107.50 1105.43 1103.23 1100.92 1098.48 1095.92 1093.24 1090.43 1087.51 1084.46 1081.30 1078.01 1074.60 1071.06 1067.41 1063.64 1059.74 1055.72 1051.58 1047.32 1042.93 1038.43 1033.80 1029.05 1024.18 1019.19
1132.11 1130.72 1129.21 1127.57 1125.82 1123.94 1121.94 1119.82 1117.58 1115.22 1112.73 1110.13 1107.40 1104.55 1101.58 1098.48 1095.27 1091.93 1088.48 1084.90 1081.20 1077.37 1073.43 1069.36 1065.18 1060.87 1056.44 1051.88 1047.21 1042.41 1037.50 1032.46 1027.30
1141.87 1140.07 1138.14 1136.09 1133.91 1131.62 1129.20 1126.67 1124.01 1121.23 1118.32 1115.30 1112.15 1108.89 1105.50 1101.99 1098.36 1094.60 1090.73 1086.73 1082.61 1078.37 1074.01 1069.53 1064.92 1060.20 1055.35 1050.38 1045.29 1040.08 1034.74
Note: Density in kg/m3.
Source: Dow Chemical (2001b)
Table 7
Specific Heat of Aqueous Solutions of Ethylene Glycol Concentrations in Volume Percent Ethylene Glycol
Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125
10%
3.937 3.946 3.954 3.963 3.972 3.981 3.989 3.998 4.007 4.015 4.024 4.033 4.042 4.050 4.059 4.068 4.077 4.085 4.094 4.103 4.112 4.120 4.129 4.138 4.147 4.155
Source: Dow Chemical (2001b)
20%
3.757 3.769 3.780 3.792 3.803 3.815 3.826 3.838 3.849 3.861 3.872 3.884 3.895 3.907 3.918 3.930 3.941 3.953 3.964 3.976 3.987 3.999 4.010 4.022 4.033 4.045 4.056
30%
3.560 3.574 3.589 3.603 3.617 3.631 3.645 3.660 3.674 3.688 3.702 3.716 3.730 3.745 3.759 3.773 3.787 3.801 3.816 3.830 3.844 3.858 3.872 3.886 3.901 3.915 3.929 3.943
40%
50%
60%
70%
80%
90%
3.334 3.351 3.367 3.384 3.401 3.418 3.435 3.451 3.468 3.485 3.502 3.518 3.535 3.552 3.569 3.585 3.602 3.619 3.636 3.653 3.669 3.686 3.703 3.720 3.736 3.753 3.770 3.787 3.804 3.820
3.068 3.088 3.107 3.126 3.145 3.165 3.184 3.203 3.223 3.242 3.261 3.281 3.300 3.319 3.339 3.358 3.377 3.396 3.416 3.435 3.454 3.474 3.493 3.512 3.532 3.551 3.570 3.590 3.609 3.628 3.647 3.667 3.686
2.844 2.866 2.888 2.909 2.931 2.953 2.975 2.997 3.018 3.040 3.062 3.084 3.106 3.127 3.149 3.171 3.193 3.215 3.236 3.258 3.280 3.302 3.324 3.345 3.367 3.389 3.411 3.433 3.454 3.476 3.498 3.520 3.542
2.612 2.636 2.660 2.685 2.709 2.733 2.757 2.782 2.806 2.830 2.854 2.878 2.903 2.927 2.951 2.975 3.000 3.024 3.048 3.072 3.097 3.121 3.145 3.169 3.193 3.218 3.242 3.266 3.290 3.315 3.339 3.363 3.387
2.370 2.397 2.423 2.450 2.477 2.503 2.530 2.556 2.583 2.610 2.636 2.663 2.690 2.716 2.743 2.770 2.796 2.823 2.850 2.876 2.903 2.929 2.956 2.983 3.009 3.036 3.063 3.089 3.116 3.143 3.169 3.196 3.223
2.177 2.206 2.235 2.264 2.293 2.322 2.351 2.380 2.409 2.438 2.467 2.496 2.525 2.554 2.583 2.612 2.641 2.670 2.699 2.728 2.757 2.786 2.815 2.844 2.873 2.902 2.931 2.960 2.989 3.018 3.047
Note: Specific heat in kJ/(kg·K).
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31.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 8
Thermal Conductivity of Aqueous Solutions of Ethylene Glycol Concentrations in Volume Percent Ethylene Glycol
Licensed for single user. © 2017 ASHRAE, Inc.
Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125
10
0.512 0.520 0.528 0.535 0.543 0.550 0.556 0.563 0.569 0.574 0.579 0.584 0.588 0.592 0.596 0.599 0.602 0.605 0.607 0.609 0.610 0.612 0.613 0.614 0.614 0.615
20
0.458 0.466 0.472 0.479 0.486 0.492 0.498 0.503 0.509 0.514 0.518 0.523 0.527 0.530 0.534 0.537 0.540 0.542 0.544 0.546 0.548 0.549 0.550 0.551 0.552 0.552 0.552
30
0.411 0.417 0.423 0.429 0.435 0.440 0.445 0.450 0.455 0.459 0.463 0.467 0.471 0.474 0.477 0.480 0.483 0.485 0.487 0.489 0.490 0.491 0.493 0.493 0.494 0.495 0.495 0.495
40
0.366 0.371 0.376 0.381 0.386 0.391 0.395 0.400 0.404 0.408 0.412 0.415 0.419 0.422 0.425 0.427 0.430 0.432 0.434 0.436 0.438 0.439 0.440 0.441 0.442 0.443 0.443 0.444 0.444 0.444
Source: Dow Chemical (2001b)
50
60
70
80
90
0.328 0.332 0.336 0.340 0.344 0.348 0.352 0.356 0.360 0.363 0.366 0.370 0.373 0.376 0.378 0.381 0.383 0.385 0.387 0.389 0.391 0.392 0.394 0.395 0.396 0.396 0.397 0.398 0.398 0.398 0.398 0.398
0.300 0.303 0.306 0.310 0.313 0.316 0.319 0.322 0.325 0.328 0.331 0.334 0.336 0.338 0.341 0.343 0.345 0.347 0.348 0.350 0.351 0.352 0.354 0.355 0.355 0.356 0.357 0.357 0.358 0.358 0.358 0.358 0.358
0.279 0.282 0.284 0.287 0.289 0.292 0.294 0.297 0.299 0.301 0.303 0.305 0.307 0.309 0.311 0.312 0.314 0.315 0.316 0.317 0.318 0.319 0.320 0.321 0.322 0.322 0.322 0.323 0.323 0.323 0.323 0.323 0.323
0.262 0.264 0.266 0.268 0.270 0.271 0.273 0.275 0.277 0.278 0.280 0.281 0.283 0.284 0.285 0.286 0.288 0.289 0.289 0.290 0.291 0.292 0.292 0.293 0.293 0.294 0.294 0.294 0.294 0.294 0.294 0.294 0.294
0.252 0.253 0.255 0.256 0.257 0.259 0.260 0.261 0.262 0.263 0.264 0.265 0.266 0.267 0.268 0.268 0.269 0.270 0.270 0.271 0.271 0.271 0.272 0.272 0.272 0.272 0.272 0.272 0.272 0.272 0.271
Note: Thermal conductivity in W/(m·K).
Table 9
Viscosity of Aqueous Solutions of Ethylene Glycol Concentrations in Volume Percent Ethylene Glycol
Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125
10
2.08 1.79 1.56 1.37 1.21 1.08 0.97 0.88 0.80 0.73 0.67 0.62 0.57 0.53 0.50 0.47 0.44 0.41 0.39 0.37 0.35 0.33 0.32 0.30 0.29 0.28
Source: Dow Chemical (2001b)
20
3.65 3.02 2.54 2.18 1.89 1.65 1.46 1.30 1.17 1.06 0.96 0.88 0.81 0.74 0.69 0.64 0.59 0.55 0.52 0.49 0.46 0.43 0.40 0.38 0.36 0.34 0.33
30
6.19 5.03 4.15 3.48 2.95 2.53 2.20 1.92 1.69 1.50 1.34 1.21 1.09 0.99 0.90 0.83 0.76 0.70 0.65 0.60 0.56 0.52 0.49 0.46 0.43 0.41 0.38 0.36
40
50
60
70
80
90
15.75 11.74 9.06 7.18 5.83 4.82 4.04 3.44 2.96 2.57 2.26 1.99 1.77 1.59 1.43 1.29 1.17 1.06 0.97 0.89 0.82 0.76 0.70 0.65 0.60 0.56 0.53 0.49 0.46 0.43
66.93 43.98 30.50 22.07 16.53 12.74 10.05 8.09 6.63 5.50 4.63 3.94 3.39 2.94 2.56 2.26 2.00 1.78 1.59 1.43 1.29 1.17 1.07 0.98 0.89 0.82 0.76 0.70 0.65 0.60 0.56 0.53 0.49
93.44 65.25 46.75 34.28 25.69 19.62 15.25 12.05 9.66 7.85 6.46 5.38 4.52 3.84 3.29 2.84 2.47 2.16 1.91 1.69 1.51 1.35 1.22 1.10 1.00 0.92 0.84 0.77 0.71 0.66 0.61 0.57 0.53
133.53 96.57 70.38 51.94 38.88 29.53 22.76 17.79 14.09 11.31 9.18 7.53 6.24 5.23 4.42 3.76 3.23 2.80 2.43 2.13 1.88 1.67 1.49 1.33 1.20 1.09 0.99 0.90 0.82 0.76 0.70 0.64 0.60
191.09 141.02 102.21 74.53 55.09 41.36 31.56 24.44 19.20 15.29 12.33 10.05 8.29 6.90 5.79 4.91 4.19 3.61 3.12 2.72 2.39 2.11 1.87 1.66 1.49 1.34 1.21 1.10 1.00 0.91 0.83 0.77 0.71
196.87 128.43 87.52 61.85 45.08 33.74 25.84 20.18 16.04 12.95 10.59 8.77 7.34 6.21 5.30 4.56 3.95 3.45 3.03 2.67 2.37 2.12 1.90 1.71 1.54 1.40 1.27 1.16 1.07 0.98 0.90
Note: Viscosity in mPa/s.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Physical Properties of Secondary Coolants (Brines) Table 10
31.9
Density of Aqueous Solutions of Industrially Inhibited Propylene Glycol Concentrations in Volume Percent Propylene Glycol
Licensed for single user. © 2017 ASHRAE, Inc.
Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125
10%
1013.85 1012.61 1011.24 1009.75 1008.13 1006.40 1004.54 1002.56 1000.46 998.23 995.88 993.41 990.82 988.11 985.27 982.31 979.23 976.03 972.70 969.25 965.68 961.99 958.17 954.24 950.18 945.99
20%
1027.24 1025.84 1024.32 1022.68 1020.91 1019.01 1016.99 1014.84 1012.56 1010.16 1007.64 1004.99 1002.21 999.31 996.28 993.12 989.85 986.44 982.91 979.25 975.47 971.56 967.53 963.37 959.09 954.67 950.14
30%
1039.42 1037.89 1036.24 1034.46 1032.55 1030.51 1028.35 1026.06 1023.64 1021.09 1018.42 1015.62 1012.69 1009.63 1006.44 1003.13 999.69 996.12 992.42 988.60 984.65 980.57 976.36 972.03 967.56 962.97 958.26 953.41
40%
1050.43 1048.79 1047.02 1045.12 1043.09 1040.94 1038.65 1036.24 1033.70 1031.03 1028.23 1025.30 1022.24 1019.06 1015.75 1012.30 1008.73 1005.03 1001.21 997.25 993.17 988.95 984.61 980.14 975.54 970.81 965.95 960.97 955.86
50%
60%
70%
80%
90%
1062.11 1060.49 1058.73 1056.85 1054.84 1052.71 1050.44 1048.04 1045.52 1042.87 1040.09 1037.18 1034.15 1030.98 1027.69 1024.27 1020.72 1017.04 1013.23 1009.30 1005.24 1001.05 996.73 992.28 987.70 983.00 978.16 973.20 968.11 962.89 957.55
1072.92 1071.31 1069.58 1067.72 1065.73 1063.61 1061.37 1059.00 1056.50 1053.88 1051.13 1048.25 1045.24 1042.11 1038.85 1035.47 1031.95 1028.32 1024.55 1020.66 1016.63 1012.49 1008.21 1003.81 999.28 994.63 989.85 984.94 979.90 974.74 969.45 964.03 958.49
1079.67 1077.82 1075.84 1073.74 1071.51 1069.16 1066.69 1064.09 1061.36 1058.51 1055.54 1052.44 1049.22 1045.87 1042.40 1038.81 1035.09 1031.25 1027.28 1023.19 1018.97 1014.63 1010.16 1005.57 1000.86 996.02 991.06 985.97 980.76 975.42 969.96 964.38 958.67
1094.50 1090.85 1087.18 1083.49 1079.77 1076.04 1072.27 1068.49 1064.68 1060.85 1057.00 1053.12 1049.22 1045.30 1041.35 1037.38 1033.39 1029.37 1025.33 1021.27 1017.19 1013.08 1008.95 1004.79 1000.62 996.41 992.19 987.94 983.68 979.38 975.07 970.73 966.37
1092.46 1088.82 1085.15 1081.46 1077.74 1074.00 1070.24 1066.46 1062.65 1058.82 1054.96 1051.09 1047.19 1043.26 1039.32 1035.35 1031.35 1027.34 1023.30 1019.24 1015.15 1011.04 1006.91 1002.76 998.58 994.38 990.16 985.91 981.64 977.35 973.03 968.69 964.33
Note: Density in kg/m3.
Source: Dow Chemical (2001a)
Table 11 Specific Heat of Aqueous Solutions of Propylene Glycol Concentrations in Volume Percent Propylene Glycol Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120 125
10%
4.042 4.050 4.058 4.067 4.075 4.083 4.091 4.099 4.107 4.115 4.123 4.131 4.139 4.147 4.155 4.163 4.171 4.179 4.187 4.195 4.203 4.211 4.219 4.227 4.235 4.243
Source: Dow Chemical (2001a)
20%
3.918 3.929 3.940 3.951 3.962 3.973 3.983 3.994 4.005 4.016 4.027 4.038 4.049 4.060 4.071 4.082 4.093 4.104 4.115 4.126 4.136 4.147 4.158 4.169 4.180 4.191 4.202
30%
3.765 3.779 3.793 3.807 3.820 3.834 3.848 3.862 3.875 3.889 3.903 3.917 3.930 3.944 3.958 3.972 3.985 3.999 4.013 4.027 4.040 4.054 4.068 4.082 4.095 4.109 4.123 4.137
40%
3.586 3.603 3.619 3.636 3.652 3.669 3.685 3.702 3.718 3.735 3.751 3.768 3.784 3.801 3.817 3.834 3.850 3.867 3.883 3.900 3.916 3.933 3.949 3.966 3.982 3.999 4.015 4.032 4.049
50%
60%
70%
80%
90%
3.358 3.378 3.397 3.416 3.435 3.455 3.474 3.493 3.513 3.532 3.551 3.570 3.590 3.609 3.628 3.648 3.667 3.686 3.706 3.725 3.744 3.763 3.783 3.802 3.821 3.841 3.860 3.879 3.898 3.918 3.937
3.096 3.118 3.140 3.162 3.184 3.206 3.228 3.250 3.272 3.295 3.317 3.339 3.361 3.383 3.405 3.427 3.449 3.471 3.493 3.515 3.537 3.559 3.581 3.603 3.625 3.647 3.670 3.692 3.714 3.736 3.758 3.780 3.802
2.843 2.868 2.893 2.918 2.943 2.968 2.993 3.018 3.042 3.067 3.092 3.117 3.142 3.167 3.192 3.217 3.242 3.266 3.291 3.316 3.341 3.366 3.391 3.416 3.441 3.465 3.490 3.515 3.540 3.565 3.590 3.615 3.640
2.572 2.600 2.627 2.655 2.683 2.710 2.738 2.766 2.793 2.821 2.849 2.876 2.904 2.931 2.959 2.987 3.014 3.042 3.070 3.097 3.125 3.153 3.180 3.208 3.236 3.263 3.291 3.319 3.346 3.374 3.402 3.429 3.457
2.264 2.295 2.326 2.356 2.387 2.417 2.448 2.478 2.509 2.539 2.570 2.600 2.631 2.661 2.692 2.723 2.753 2.784 2.814 2.845 2.875 2.906 2.936 2.967 2.997 3.028 3.058 3.089 3.119 3.150 3.181 3.211 3.242
Note: Specific heat in kJ/(kg·K).
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
31.10
2017 ASHRAE Handbook—Fundamentals (SI) Table 12
Thermal Conductivity of Aqueous Solutions of Propylene Glycol Concentrations in Volume Percent Propylene Glycol
Licensed for single user. © 2017 ASHRAE, Inc.
Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120
10
0.510 0.518 0.526 0.534 0.541 0.548 0.555 0.561 0.567 0.573 0.578 0.583 0.587 0.591 0.595 0.598 0.601 0.604 0.606 0.608 0.609 0.611 0.612 0.613 0.613
20
0.449 0.456 0.463 0.470 0.477 0.483 0.489 0.494 0.500 0.505 0.509 0.514 0.518 0.521 0.525 0.528 0.531 0.533 0.535 0.537 0.538 0.540 0.541 0.542 0.542 0.543
30
40
0.346 0.351 0.356 0.361 0.366 0.371 0.376 0.380 0.384 0.388 0.392 0.396 0.399 0.402 0.405 0.408 0.410 0.413 0.415 0.416 0.418 0.419 0.420 0.421 0.422 0.423 0.423 0.423 0.423
0.397 0.403 0.409 0.415 0.421 0.426 0.431 0.436 0.441 0.445 0.450 0.453 0.457 0.460 0.463 0.466 0.469 0.471 0.473 0.474 0.476 0.477 0.478 0.479 0.480 0.480 0.480
Source: Dow Chemical (2001a)
50
60
70
80
90
0.302 0.306 0.311 0.315 0.319 0.323 0.327 0.331 0.334 0.338 0.341 0.344 0.347 0.350 0.353 0.355 0.358 0.360 0.362 0.363 0.365 0.366 0.367 0.368 0.369 0.370 0.370 0.371 0.371 0.371 0.371
0.269 0.272 0.275 0.278 0.282 0.285 0.288 0.291 0.294 0.297 0.299 0.302 0.304 0.307 0.309 0.311 0.313 0.314 0.316 0.317 0.319 0.320 0.321 0.321 0.322 0.323 0.323 0.323 0.323 0.323 0.323 0.323
0.242 0.245 0.247 0.250 0.252 0.254 0.256 0.259 0.261 0.263 0.265 0.267 0.268 0.270 0.272 0.273 0.274 0.275 0.277 0.277 0.278 0.279 0.280 0.280 0.281 0.281 0.281 0.281 0.281 0.281 0.281 0.280
0.220 0.222 0.224 0.226 0.227 0.229 0.230 0.232 0.233 0.235 0.236 0.237 0.239 0.240 0.241 0.242 0.242 0.243 0.244 0.244 0.245 0.245 0.246 0.246 0.246 0.246 0.246 0.246 0.246 0.246 0.245 0.245
0.203 0.204 0.205 0.206 0.207 0.208 0.209 0.210 0.211 0.212 0.213 0.214 0.215 0.215 0.216 0.216 0.217 0.217 0.218 0.218 0.218 0.218 0.218 0.218 0.218 0.218 0.218 0.218 0.218 0.217 0.217 0.216
Note: Thermal conductivity in W/(m·K).
Table 13 Viscosity of Aqueous Solutions of Propylene Glycol Concentrations in Volume Percent Propylene Glycol Temperature, °C –35 –30 –25 –20 –15 –10 –5 0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80 85 90 95 100 105 110 115 120
10
2.68 2.23 1.89 1.63 1.42 1.25 1.11 0.99 0.89 0.81 0.73 0.67 0.62 0.57 0.53 0.49 0.46 0.43 0.40 0.38 0.35 0.33 0.32 0.30 0.28
Source: Dow Chemical (2001a)
20
4.98 4.05 3.34 2.79 2.36 2.02 1.74 1.52 1.34 1.18 1.06 0.95 0.86 0.78 0.71 0.66 0.60 0.56 0.52 0.49 0.45 0.43 0.40 0.38 0.36 0.34
30
11.84 9.07 7.07 5.61 4.52 3.69 3.06 2.57 2.19 1.88 1.63 1.43 1.26 1.13 1.01 0.92 0.83 0.76 0.70 0.65 0.60 0.56 0.53 0.50 0.47 0.45 0.42
40
48.90 33.07 23.11 16.63 12.30 9.32 7.21 5.70 4.59 3.75 3.12 2.62 2.24 1.93 1.68 1.48 1.31 1.18 1.06 0.96 0.88 0.81 0.75 0.69 0.65 0.60 0.57 0.54 0.51
50
60
70
80
90
171.54 109.69 72.42 49.29 34.51 24.81 18.28 13.77 10.59 8.30 6.62 5.36 4.41 3.68 3.10 2.65 2.28 1.99 1.75 1.55 1.38 1.24 1.12 1.02 0.93 0.86 0.79 0.74 0.69 0.64 0.60
524.01 330.39 211.43 137.96 92.00 62.78 43.84 31.32 22.87 17.05 12.96 10.04 7.91 6.34 5.15 4.25 3.55 3.00 2.57 2.22 1.93 1.70 1.51 1.35 1.22 1.10 1.01 0.92 0.85 0.79 0.74 0.69
916.18 551.12 340.09 215.67 140.62 94.23 64.83 45.74 33.04 24.41 18.41 14.15 11.08 8.81 7.12 5.84 4.85 4.08 3.46 2.98 2.58 2.26 1.99 1.77 1.59 1.43 1.30 1.18 1.08 1.00 0.93 0.86
1434.22 908.47 575.92 368.77 239.86 159.02 107.64 74.45 52.63 37.99 28.00 21.04 16.10 12.55 9.94 7.99 6.52 5.39 4.51 3.82 3.28 2.83 2.47 2.18 1.94 1.73 1.56 1.42 1.29 1.19 1.09 1.02
3813.29 2071.34 1176.09 696.09 428.19 272.94 179.78 122.03 85.15 60.93 44.62 33.38 25.45 19.76 15.60 12.49 10.15 8.35 6.95 5.85 4.97 4.26 3.69 3.22 2.83 2.50 2.23 2.00 1.80 1.63 1.48 1.35
Note: Viscosity in mPa·s.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Physical Properties of Secondary Coolants (Brines)
31.11
Fig. 12 Viscosity of Aqueous Solutions of Industrially Inhibited Ethylene Glycol (vol. %) (Dow Chemical 2001b)
Fig. 14 Specific Heat of Aqueous Solutions of Industrially Inhibited Propylene Glycol (vol. %) Licensed for single user. © 2017 ASHRAE, Inc.
(Dow Chemical 2001b)
Fig. 13 Density of Aqueous Solutions of Industrially Inhibited Propylene Glycol (vol. %) (Dow Chemical 2001b)
Fig. 15 Thermal Conductivity of Aqueous Solutions of Industrially Inhibited Propylene Glycol (vol. %) (Dow Chemical 2001b)
leads to a downward shift in solution pH, which negates the usefulness of some corrosion inhibitors (particularly for cast iron and carbon steels, but also solders). Properly inhibited glycol products should be used.
Service Considerations Design Considerations. Inhibited glycols can be used at temperatures as high as 175°C. However, maximum-use temperatures vary from fluid to fluid, so the manufacturer’s suggested temperature-use ranges should be followed. In systems with a high degree of aeration, the bulk fluid temperature should not exceed 65°C; however, temperatures up to 175°C are permissible in a pressurized system if air intake is eliminated. Maximum film temperatures should not exceed 28 K above the bulk temperature. Nitrogen blanketing minimizes oxidation when the system operates at elevated temperatures for extended periods. Minimum operating temperatures for a recirculating fluid are typically –29°C for ethylene glycol solutions and –18°C for propylene glycol solutions. Operation below these temperatures is generally impractical, because the fluids’ viscosity builds dramatically,
Fig. 16 Viscosity of Aqueous Solutions of Industrially Inhibited Propylene Glycol (vol. %) (Dow Chemical 2001a)
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Licensed for single user. © 2017 ASHRAE, Inc.
31.12 thus increasing pumping power requirements and reducing heat transfer film coefficients. Standard materials can be used with most inhibited glycol solutions, except galvanized metals, which form insoluble zinc salts with the corrosion inhibitors. This depletes corrosion inhibitors below effective limits, and can cause excessive insoluble salt (sludge) formation. Because removal of sludge and other contaminants is critical, install suitable filters. If inhibitors are rapidly and completely adsorbed by such contamination, the fluid is ineffective for corrosion inhibition. Consider such adsorption when selecting filters. Storage and Handling. Inhibited glycol concentrates are stable, relatively noncorrosive materials with high flash points. These fluids can be stored in mild steel, stainless steel, or aluminum vessels. However, aluminum should be used only when the fluid temperature is below 65°C. Corrosion in the vapor space of vessels may be a problem, because the fluid’s inhibitor package cannot reach these surfaces to protect them. A protective coating may be necessary (e.g., novolac-based vinyl ester resins, high-bake phenolic resins, polypropylene, polyvinylidene fluoride). To ensure the coating is suitable for a particular application and temperature, consult the manufacturer. Because the chemical properties of an inhibited glycol concentrate differ from those of its dilutions, the effect of the concentrate on different containers should be known when selecting storage. Choose transfer pumps only after considering temperature/ viscosity data. Centrifugal pumps with electric motor drives are often used. Materials compatible with ethylene or propylene glycol should be used for pump packing material. Mechanical seals are also satisfactory. Bypass or in-line filters are recommended to remove suspended particles, which can abrade seal surfaces. Welded mild steel transfer piping with a minimum diameter is normally used in conjunction with the piping, although flanged and gasketed joints are also satisfactory. Preparation Before Application. Before an inhibited glycol is charged into a system, remove residual contaminants such as sludge, rust, brine deposits, and oil so the newly installed fluid functions properly. Avoid strong acid cleaners; if they are required, consider inhibited acids. Completely remove the cleaning agent before charging with inhibited glycol, because residual cleaning agents are known to cause system foaming problems. Foaming. Adding glycols to water reduces surface tension and increases viscosity, particularly for propylene glycol, and this affects the rate at which dispersed gas bubbles coalesce and release to the surface of the liquid. If relatively high amounts of impurities accumulate either from the original source of glycol-based heat transfer fluid or from thermal oxidative degradation of the glycol, severe foaming problems may occur. Dilution Water. Purified water such as distilled, deionized, or condensate water should be used whenever diluting or topping off a system. Nonpurified water contains corrosive ions and scale-promoting elements that reduce the effectiveness of the inhibited formulation. If purified water of this quality is unavailable, then ensure the source of water contains less than 25 mg/kg chloride, less than 25 mg/kg sulfate, and less than 100 mg/kg of total hardness before using. Fluid Maintenance. Glycol concentrations can be determined by refractive index, gas chromatography, or Karl Fischer analysis for water (assuming that the concentration of other fluid components, such as inhibitor, is known). Using density to determine glycol concentration is unsatisfactory because (1) density measurements are temperature sensitive, (2) inhibitor concentrations can change density, (3) values for propylene glycol are close to those of water, and (4) propylene glycol values exhibit a maximum at 70 to 75% concentration. An effective inhibitor monitoring and maintenance schedule is essential to keep a glycol solution relatively noncorrosive for a long
2017 ASHRAE Handbook—Fundamentals (SI) period. Inspection immediately after installation, and annually thereafter, is normally an effective practice. Visual inspection of solution and filter residue can often detect potential system problems. Many manufacturers of inhibited glycol-based heat transfer fluids provide analytical service to ensure that their product remains in good condition. This analysis may include some or all of the following: percent of ethylene and/or propylene glycol, freezing point, pH, reserve alkalinity, corrosion inhibitor evaluation, contaminants, total hardness, metal content, and degradation products. If maintenance on the fluid is required, recommendations may be given along with the analysis results. Properly inhibited and maintained glycol solutions provide better corrosion protection than brine solutions in most systems. A long, though not indefinite, service life can be expected. Avoid indiscriminate mixing of inhibited formulations.
3.
HALOCARBONS
Many common refrigerants are used as secondary coolants as well as primary refrigerants. Their favorable properties as heat transfer fluids include low freezing points, low viscosities, nonflammability, and good stability. Chapters 29 and 30 present physical and thermodynamic properties for common refrigerants. Tables 1 and 2 in Chapter 29 summarizes comparative safety characteristics for halocarbons. ACGIH has more information on halocarbon toxicity threshold limit values and biological exposure indices (see the Bibliography). Construction materials and stability factors in halocarbon use are discussed in Chapter 29 of this volume and Chapter 6 of the 2014 ASHRAE Handbook—Refrigeration.
4.
NONHALOCARBON, NONAQUEOUS FLUIDS
Numerous additional secondary refrigerants, used primarily by the chemical processing and pharmaceutical industries, have been used rarely in the HVAC and allied industries because of their cost and relative novelty. Before choosing these types of fluids, consider electrical classifications, disposal, potential worker exposure, process containment, and other relevant issues. Tables 14 to 16 list physical properties for a mixture of dimethylsiloxane polymers of various relative molecular masses (Dow Corning 1989) and d-limonene. Information on d-limonene is limited; it is based on measurements made over small temperature ranges or simply on standard physical property estimation techniques. The compound (molecular formula C10H16) is derived as an extract from orange and lemon oils. The mixture of dimethylsiloxane polymers can be used with most standard construction materials; d-limonene, however, can be quite corrosive, easily autooxidizing at ambient temperatures. This fact should be understood and considered before using d-limonene in a system.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. Andrepont, J.A. 2012. Applications of low temperature fluid in thermally stratified thermal energy storage. Paper CH-12-C063. Presented at ASHRAE Winter Conference, Chicago. ATSDR. 2016. Priority list of hazardous substances that will be the candidates for toxicological profiles. www.atsdr.cdc.gov/SPL/index.html. Agency for Toxic Substances and Disease Registry, Atlanta, GA. Carrier Air Conditioning Company. 1959. Basic data, Section 17M. Syracuse, NY.
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Physical Properties of Secondary Coolants (Brines)
31.13
Table 14 Properties of Polydimethylsiloxane Heat Transfer Fluid Vapor Heat Thermal Temperature, Pressure, Viscosity, Density, Capacity, Conductivity, °C kPa mPa·s kg/m3 kJ/(kg·K) W/(m·K)
Licensed for single user. © 2017 ASHRAE, Inc.
–73 –70 –60 –50 –40 –30 –20 –10 0 10 20 30 40 50 60 70 80 90
0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.01 0.03 0.08 0.16 0.32 0.61 1.09 1.85 3.02 4.76 7.25
12.4 11.2 8.26 6.24 4.83 3.81 3.07 2.51 2.09 1.76 1.49 1.29 1.12 0.98 0.86 0.77 0.69 0.62
924.6 922.1 913.5 905.0 896.4 887.9 879.3 870.7 862.0 853.3 844.5 835.5 826.5 817.3 807.9 798.4 788.7 778.8
1.410 1.418 1.443 1.469 1.495 1.520 1.546 1.572 1.597 1.623 1.649 1.674 1.700 1.726 1.751 1.777 1.803 1.828
0.1294 0.1288 0.1269 0.1251 0.1231 0.1212 0.1192 0.1171 0.1150 0.1129 0.1108 0.1086 0.1064 0.1042 0.1019 0.0996 0.0973 0.0949
Heat Vapor Thermal Temperature, Pressure, Viscosity, Density, Capacity, Conductivity, °C W/(m·K) kPa mPa·s kg/m3 kJ/(kg·K) 100 110 120 130 140 150 160 170 180 190 200 210 220 230 240 250 260
10.73 15.45 21.75 29.95 40.45 53.67 70.06 90.10 114.29 143.17 177.27 217.14 263.36 316.47 377.03 445.61 522.74
0.56 0.51 0.47 0.43 0.40 0.37 0.34 0.32 0.30 0.28 0.26 0.25 0.24 0.22 0.21 0.20 0.19
768.7 758.3 747.7 736.8 725.6 714.1 702.3 690.2 677.7 664.8 651.6 638.0 623.9 609.5 594.5 579.1 563.3
1.854 1.880 1.905 1.931 1.957 1.982 2.008 2.033 2.059 2.085 2.110 2.136 2.162 2.187 2.213 2.239 2.264
0.0925 0.0901 0.0877 0.0852 0.0827 0.0802 0.0777 0.0751 0.0725 0.0699 0.0673 0.0646 0.0620 0.0593 0.0566 0.0538 0.0511
Source: Dow Chemical (1998)
Table 15 Summary of Physical Properties of Polydimethylsiloxane Mixture and d-Limonene Polydimethylsiloxane Mixture d-Limonene Flash point, °C, closed cup Boiling point, °C Freezing point, °C Operational temperature range, °C
46.7 175 –111.1 –73.3 to 260
46.1 154.4 –96.7 None published
Source: Dow Corning (1989).
Table 16 Physical Properties of d-Limonene Temperature, Specific Heat, Viscosity, °C kJ/(kg·K) mPa·s –73 –50 –25 0 25 50 75 100 125 150
1.27 1.39 1.51 1.65 1.78 1.91 2.04 2.17 2.30 2.41
3.8 3.0 2.3 1.8 1.4 1.1 0.8 0.7 0.5 0.4
Density, kg/m3
Thermal Conductivity, W/(m·K)
914.3 897.1 878.3 859.2 839.8 820.1 800.0 779.5 758.4 736.6
0.137 0.133 0.128 0.124 0.119 0.114 0.110 0.105 0.100 0.096
Source: Dow Corning (1989) Note: Properties are estimated or based on incomplete data.
CCI. 1953. Calcium chloride for refrigeration brine. Manual RM-1. Calcium Chloride Institute. Dow Chemical. 1998. Syltherm XLT heat transfer fluid. Midland, MI. Dow Chemical USA. 2001a. Engineering and operating guideline for DOWFROST and DOWFROST HD inhibited propylene glycol heat transfer fluids. Midland, MI. Dow Chemical USA. 2001b. Engineering manual for DOWTHERM SR-1 and DOWTHERM 4000 inhibited ethylene glycol heat transfer fluids. Midland, MI. Dow Corning USA (now Dow Chemical). 1989. Syltherm heat transfer liquids. Midland, MI. Melinder, Å. 2007. Thermo-physical properties of aqueous solutions used as secondary working fluids. Ph.D. dissertation, Department of Energy Technology, Kungliga Tekniska Högskolan, Stockholm. urn.kb.se/resolve ?urn=urn:nbn:se:kth:diva-4406. USP. 2016. Food chemicals codex (FCC), 10th ed. U.S. Pharmacopeial Convention, Rockville, MD.
BIBLIOGRAPHY ACGIH. Annual. TLVs® and BEIs®. American Conference of Governmental Industrial Hygienists, Cincinnati. ASM. 2000. Corrosion: Understanding the basics. J.R. Davis, ed. ASM International, Materials Park, OH. Born, D.W. 1989. Inhibited glycols for corrosion and freeze protection in water-based heating and cooling systems. Midland, MI. Fontana, M.G. 1986. Corrosion engineering. McGraw-Hill, New York. NACE. 1973. Corrosion inhibitors. C.C. Nathan, ed. National Association of Corrosion Engineers, Houston. NACE. 2002. NACE corrosion engineer’s reference book, 3rd ed. R. Baboian, ed. National Association of Corrosion Engineers, Houston.
Related Commercial Resources
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Related Commercial Resources CHAPTER 32
SORBENTS AND DESICCANTS Desiccant Applications.................................................................................................................. Desiccant Cycle............................................................................................................................. Types of Desiccants....................................................................................................................... Desiccant Isotherms ...................................................................................................................... Desiccant Life................................................................................................................................ Cosorption of Water Vapor and Indoor Air Contaminants ..........................................................
ORPTION refers to the binding of one substance to another. Sorbents are materials that have an ability to attract and hold other gases or liquids. They can be used to attract gases or liquids other than water vapor, which makes them very useful in chemical separation processes. Desiccants are a subset of sorbents; they have a particular affinity for water. Virtually all materials are desiccants; that is, they attract and hold water vapor. Wood, natural fibers, clays, and many synthetic materials attract and release moisture as commercial desiccants do, but they lack holding capacity. For example, woolen carpet fibers attract up to 23% of their dry mass in water vapor, and nylon can take up almost 6% of its mass in water. In contrast, a commercial desiccant takes up between 10 and 1100% of its dry mass in water vapor, depending on its type and on the moisture available in the environment. Furthermore, commercial desiccants continue to attract moisture even when the surrounding air is quite dry, a characteristic that other materials do not share. All desiccants behave in a similar way: they attract moisture until they reach equilibrium with the surrounding air. Moisture is usually removed from the desiccant by heating it to temperatures between 50 and 205°C and exposing it to a scavenger airstream. After the desiccant dries, it must be cooled so that it can attract moisture once again. Sorption always generates sensible heat equal to the latent heat of the water vapor taken up by the desiccant plus an additional heat of sorption that varies between 5 and 25% of the latent heat of the water vapor. This heat is transferred to the desiccant and to the surrounding air. The process of attracting and holding moisture is described as either adsorption or absorption, depending on whether the desiccant undergoes a chemical change as it takes on moisture. Adsorption does not change the desiccant, except by addition of the mass of water vapor; it is similar in some ways to a sponge soaking up water. Absorption, on the other hand, changes the desiccant. An example of an absorbent is lithium chloride, which changes from a solid to a liquid as it absorbs moisture.
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S
1.
DESICCANT APPLICATIONS
Desiccants can dry either liquids or gases, including ambient air, and are used in many air-conditioning applications, particularly when the • Latent load is large in comparison to the sensible load • Energy cost to regenerate the desiccant is low compared to the cost of energy to dehumidify the air by chilling it below its dew point and reheating it • Moisture control level for the space would require chilling the air to subfreezing dew points if compression refrigeration alone were used to dehumidify the air The preparation of this chapter is assigned to TC 8.12, Desiccant Dehumidification Equipment and Components.
32.1 Copyright © 2017, ASHRAE
32.1 32.1 32.3 32.5 32.5 32.5
• Temperature control level for the space or process requires continuous delivery of air at subfreezing temperatures • Air delivered to a space or ductwork must be at less than 70% rh In any of these situations, the cost of running a vapor compression cooling system can be very high. A desiccant process may offer considerable advantages in energy, initial cost of equipment, and maintenance. Because desiccants can attract and hold more than simply water vapor, they can remove contaminants from airstreams to improve indoor air quality. Desiccants have been used to remove organic vapors and, in special circumstances, to control microbiological contaminants (Battelle 1971; Buffalo Testing Laboratory 1974). Hines et al. (1991) also confirmed their usefulness in removing vapors that can degrade indoor air quality. Desiccant materials can adsorb hydrocarbon vapors while collecting moisture from air. These cosorption phenomena show promise of improving indoor air quality in typical building HVAC systems. Desiccants are also used in drying compressed air to low dew points. In this application, moisture can be removed from the desiccant without heat. Desorption uses differences in vapor pressures compared to the total pressures of the compressed and ambient pressure airstreams. Finally, desiccants are used to dry the refrigerant circulating in air-conditioning and refrigeration systems. This reduces corrosion in refrigerant piping and prevents valves and capillaries from becoming clogged with ice crystals. In this application, the desiccant is not regenerated; it is discarded when it has adsorbed its limit of water vapor. This chapter discusses the water sorption characteristics of desiccant materials and explains some of the implications of those characteristics in ambient pressure air-conditioning applications. Information on other applications for desiccants can be found in Chapter 36 of this volume; Chapters 7, 8, 18, 39, and 44 of the 2014 ASHRAE Handbook—Refrigeration; Chapters 1, 2, 6, 10, 18, 20, 23, 30, and 46 of the 2015 ASHRAE Handbook—HVAC Applications; and Chapters 24 and 26 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment.
2.
DESICCANT CYCLE
Practically speaking, all desiccants function the same way: by moisture transfer caused by a difference between water vapor pressures at their surface and of the surrounding air. When the vapor pressure at the desiccant surface is lower than that of the air, the desiccant attracts moisture. When the surface vapor pressure is higher than that of the surrounding air, the desiccant releases moisture. Figure 1 shows the moisture content relationship between a desiccant and its surface vapor pressure. As the desiccant’s moisture content rises, so does the water vapor pressure at its surface. At some point, the vapor pressure at the desiccant surface is the same as that of the air: the two are in equilibrium. Then, moisture cannot move in
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32.2
Fig. 1
2017 ASHRAE Handbook—Fundamentals (SI)
Desiccant Water Vapor Pressure as Function of Moisture Content
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(Harriman 2003)
Fig. 3 Desiccant Cycle (Harriman 2003)
Table 1 Vapor Pressures and Dew-Point Temperatures Corresponding to Different Relative Humidities at 21°C Relative Humidity at 21°C, % 10 20 30 40 50 60 70 80 90 100
Dew Point, °C –10.5 –2.5 2.8 6.9 10.2 13.0 15.3 17.4 19.3 21.0
Vapor Pressure, kPa 0.25 0.50 0.75 1.00 1.24 1.49 1.74 1.99 2.24 2.49
Regeneration energy is equal to the sum of the heat
(Harriman 2003)
• Necessary to raise the desiccant to a temperature high enough to make its surface vapor pressure higher than that of the surrounding air • Necessary to vaporize the moisture it contains (about 2465 kJ/kg) • From desorption of water from the desiccant (a small amount)
either direction until some external force changes the vapor pressure at the desiccant or in the air. Figure 2 shows the effect of temperature on vapor pressure at the desiccant surface. Both higher temperature and increased moisture content increase surface vapor pressure. When surface vapor pressure exceeds that of the surrounding air, moisture leaves the desiccant (reactivation or regeneration). After the desiccant is dried (reactivated) by the heat, its vapor pressure remains high, so it has very little ability to absorb moisture. Cooling the desiccant reduces its surface vapor pressure so that it can absorb moisture again. The complete cycle is illustrated in Figure 3. The economics of desiccant operation depend on the energy cost of moving a given material through this cycle. Dehumidifying air (loading the desiccant with water vapor) generally proceeds without energy input other than fan and pump costs. The major portion of energy is invested in regenerating the desiccant (moving from point 2 to point 3) and cooling the desiccant (point 3 to point 1).
The cooling energy is proportional to the (1) mass of desiccant cycled and (2) difference between its temperature after regeneration and the lower temperature that allows the desiccant to remove water from the airstream again. The cycle is similar when desiccants are regenerated using pressure differences in a compressed air application. The desiccant is saturated in a high-pressure chamber (i.e., that of the compressed air). Then valves open, isolating the compressed air from the material, and the desiccant is exposed to air at ambient pressure. The saturated desiccant’s vapor pressure is much higher than ambient air at normal pressures; thus, moisture leaves the desiccant for the surrounding air. An alternative desorption strategy returns a small portion of dried air to the moist desiccant bed to reabsorb moisture, then vents that moist air to the atmosphere. Table 1 shows the range of vapor pressures over which the desiccant must operate in space-conditioning applications. It converts the relative humidity at 21°C to dew point and the corresponding vapor
Fig. 2 Desiccant Water Vapor Pressure as Function of Desiccant Moisture Content and Temperature
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Sorbents and Desiccants pressure. The greater the difference between the air and desiccant surface vapor pressures, the greater the ability of the material to absorb moisture from the air at that moisture content. The ideal desiccant for a particular application depends on the range of water vapor pressures likely to occur in the air, temperature of the regeneration heat source, and moisture sorption and desorption characteristics of the desiccant within those constraints. In commercial practice, however, most desiccants can be made to perform well in a wide variety of operating situations through careful engineering of mechanical aspects of the dehumidification system. Some of these hardware issues are discussed in Chapter 24 of the 2016 ASHRAE Handbook—HVAC Systems and Equipment.
3.
TYPES OF DESICCANTS
Desiccants can be liquids or solids and can hold moisture through absorption or adsorption, as described earlier. Most absorbents are liquids, and most adsorbents are solids.
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Liquid Absorbents Liquid absorption dehumidification can best be illustrated by comparison to air washer operation. When air passes through an air washer, its dew point approaches the temperature of water supplied to the machine. Air that is more humid is dehumidified, and air that is less humid is humidified. Similarly, a liquid absorption dehumidifier brings air into contact with a liquid desiccant solution. The liquid’s vapor pressure is lower than water at the same temperature, and air passing over the solution approaches this reduced vapor pressure; it is dehumidified. A liquid absorption solution’s vapor pressure is directly proportional to its temperature and inversely proportional to its concentration. Figure 4 illustrates the effect of increasing desiccant concentration on the water vapor pressure at its surface. The figure shows the vapor pressures of various solutions of water and triethylene glycol, a commercial liquid desiccant. As the mixture’s glycol
Fig. 4
Surface Vapor Pressure of Water/Triethylene Glycol Solutions (from data of Dow 1981)
32.3 content increases, its vapor pressure decreases. This lower pressure allows the glycol solution to absorb moisture from the air whenever the air’s vapor pressure is greater than that of the solution. Viewed another way, the vapor pressure of a given concentration of absorbent solution approximates the vapor pressure values of a fixed relative humidity line on a psychrometric chart. Higher solution concentrations give lower equilibrium relative humidities, which allow the absorbent to dry air to lower levels. Figure 5 illustrates the effect of temperature on vapor pressures of various solutions of water and lithium chloride (LiCl), another common liquid desiccant. A solution that is 25% lithium chloride has a vapor pressure of 1.25 kPa at a temperature of 21°C. If the same 25% solution is heated to 37.8°C, its vapor pressure more than doubles to 3.3 kPa. Expressed another way, the 21°C, 25% solution is in equilibrium with air at a 10.5°C dew point. The same 25% solution at 37.8°C is at equilibrium with an airstream at a 26°C dew point. The warmer the desiccant, the less moisture it can attract from the air. In standard practice, behavior of a liquid desiccant is controlled by adjusting its temperature, concentration, or both. Desiccant temperature is controlled by simple heaters and coolers. Concentration is controlled by heating the desiccant to drive moisture out into a waste airstream or directly to the ambient. Commercially available liquid desiccants have an especially high water-holding capacity. Each molecule of LiCl, for example, can hold two water molecules, even in the dry state. Above two water molecules per molecule of LiCl, the desiccant becomes a liquid and continues to absorb water. If the solution is in equilibrium with air at 90% rh, approximately 26 water molecules are attached to each molecule of LiCl. This represents a water absorption of more than 1000% on a dry-mass basis. As a practical matter, however, the absorption process is limited by the exposed surface area of desiccant and by the contact time allowed for reaction. More surface area and more contact time allow
Fig. 5 Surface Vapor Pressure of Water/Lithium Chloride Solutions (from data of Foote Mineral 1988)
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32.4 the desiccant to approach its theoretical capacity. Commercial desiccant systems stretch these limits by flowing liquid desiccant onto an extended surface, much like in a cooling tower.
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Solid Adsorbents Adsorbents are solid materials with a tremendous internal surface area per unit of mass; a single gram can have more than 4600 m2 of surface area. Structurally, adsorbents resemble a rigid sponge, and the surface of the sponge in turn resembles the ocean coastline of a fjord. This analogy indicates the scale of the different surfaces in an adsorbent. The fjords can be compared to the capillaries in the adsorbent. Spaces between the grains of sand on the fjord beaches can be compared to the spaces between the individual molecules of adsorbent, all of which have the capacity to hold water molecules. The bulk of adsorbed water is contained by condensation into the capillaries, and the majority of the surface area that attracts individual water molecules is in the crystalline structure of the material itself. Adsorbents attract moisture because of the electrical field at the desiccant surface. The field is not uniform in either force or charge, so specific sites on the desiccant surface attract water molecules that have a net opposite charge. When the complete surface is covered, the adsorbent can hold still more moisture because vapor condenses into the first water layer and fills capillaries throughout the material. As with liquid absorbents, an adsorbent’s ability to attract moisture depends on the difference in vapor pressure between its surface and the air. Capacity of solid adsorbents is generally less than the capacity of liquid absorbents. For example, a typical molecular sieve adsorbent can hold 17% of its dry mass in water when the air is at 21°C and 20% rh. In contrast, LiCl can hold 130% of its mass at the same temperature and relative humidity. Solid adsorbents have several advantages, however. For example, molecular sieves continue to adsorb moisture even when they are quite hot, allowing dehumidification of very warm airstreams. Also, several solid adsorbents can be manufactured to precise tolerances, with pore diameters that can be closely controlled. This means they can be tailored to adsorb molecules of a specific diameter. Water, for example, has an effective molecular diameter of 0.32 nm. A molecular sieve adsorbent with an average pore diameter of 0.40 nm adsorbs water but has almost no capacity for larger molecules, such as organic solvents. This selective adsorption characteristic is useful in many applications. For example, several desiccants with different pore sizes can be combined in series to remove first water and then other specific contaminants from an airstream. Adsorption Behavior. Adsorption behavior depends on (1) total surface area, (2) total volume of capillaries, and (3) range of capillary diameters. A large surface area gives the adsorbent a larger capacity at low relative humidities. Large capillaries provide a high capacity for condensed water, which gives the adsorbent a higher capacity at high relative humidities. A narrow range of capillary diameters makes an adsorbent more selective in the vapor molecules it can hold. In designing a desiccant, some tradeoffs are necessary. For example, materials with large capillaries necessarily have a smaller surface area per unit of volume than those with smaller capillaries. As a result, adsorbents are sometimes combined to provide a high adsorption capacity across a wide range of operating conditions. Figure 6 illustrates this point using three noncommercial silica gel adsorbents prepared for use in laboratory research. Each has a different internal structure, but because they are all silicas, they have similar surface adsorption characteristics. Gel 1 has large capillaries, making its total volume large but its total surface area small. It has a large adsorption capacity at high relative humidities but adsorbs a small amount at low relative humidities. In contrast, gel 8 has a capillary volume one-seventh the size of gel 1, but a total surface area almost twice as large. This gives it a
2017 ASHRAE Handbook—Fundamentals (SI) higher capacity at low relative humidities but a lower capacity to hold the moisture that condenses at high relative humidities. Silica gels and most other adsorbents can be manufactured to provide optimum performance in a specific application, balancing capacity against strength, mass, and other favorable characteristics (Bry-Air 1986). Types of Solid Adsorbents. General classes of solid adsorbents include • • • • • •
Silica gels Zeolites Synthetic zeolites (molecular sieves) Activated aluminas Carbons Synthetic polymers
Silica gels are amorphous solid structures formed by condensing soluble silicates from solutions of water or other solvents. Advantages include relatively low cost and relative simplicity of structural customizing. They are available as large as spherical beads about 5 mm in diameter or as small as grains of a fine powder. Zeolites are aluminosilicate minerals. They occur in nature and are mined rather than synthesized. Zeolites have a very open crystalline lattice that allows molecules such as water vapor to be held inside the crystal itself, like an object in a cage. Particular atoms of
Gel Number
Total Surface Area, m2/g
Average Capillary Diameter, nm
Total Volume of Capillaries, mm3/g
1 5 8
315 575 540
21 3.8 2.2
1700 490 250
Fig. 6
Adsorption and Structural Characteristics of Some Experimental Silica Gels (from data of Oscic and Cooper 1982)
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Sorbents and Desiccants an aluminosilicate determine the size of the openings between the “bars” of the cage, which in turn governs the maximum size of the molecule that can be adsorbed into the structure. Synthetic zeolites, also called molecular sieves, are crystalline aluminosilicates manufactured in a thermal process. Controlling process temperature and composition of ingredient materials allows close control of the adsorbent’s structure and surface characteristics. At a somewhat higher cost, this provides a much more uniform product than naturally occurring zeolites. Activated aluminas are oxides and hydrides of aluminum that are manufactured in thermal processes. Their structural characteristics can be controlled by the gases used to produce them and by the temperature and duration of the thermal process. Carbons are most frequently used for adsorption of gases other than water vapor because they have a greater affinity for the nonpolar molecules typical of organic solvents. Like other adsorbents, carbons have a large internal surface and especially large capillaries. This capillary volume gives them a high capacity to adsorb water vapor at relative humidities of 45 to 100%. Synthetic polymers have potential for use as desiccants, as well. Long molecules, like those found in polystyrenesulfonic acid sodium salt (PSSASS), are twisted together like strands of string. Each of the many sodium ions in the long PSSASS molecules has the potential to bind several water molecules, and the spaces between the packed strings can also contain condensed water, giving the polymer a capacity exceeding that of many other solid adsorbents.
4.
32.5 approximately 10 to 20% after three years of operation. If the concentration were 30 mg/kg, this reduction would occur after one year. In contaminated environments, equipment manufacturers often arrange filters to remove these products of reaction, and provide devices to replenish desiccant so that capacity stays constant. Solid adsorbents tend to be less chemically reactive and more sensitive to clogging, a function of the type and quantity of particulate material in the airstream. Also, certain types of silica gel can be sensitive to saturated airstreams or to liquid moisture carried over from cooling coils into the desiccant bed. In more challenging applications, thermally stabilized desiccants are used in place of less durable materials. In air-conditioning applications, desiccant equipment is designed to minimize the need for desiccant replacement in much the same way that vapor compression cooling systems are designed to avoid the need for compressor replacement. Unlike filters, desiccants are seldom intended to be frequently replaced during normal service in an air-drying application.
6.
COSORPTION OF WATER VAPOR AND INDOOR AIR CONTAMINANTS
Hines et al. (1991) confirmed that many desiccant materials can collect common indoor pollutants while they collect water vapor from ambient air. This characteristic promises to become useful in
DESICCANT ISOTHERMS
Figure 7 shows a rough comparison of sorption characteristics of different desiccants. Large variations from these isotherms occur because manufacturers use different methods to optimize materials for different applications. Suitability of a given desiccant to a particular application generally depends as much on the engineering of the mechanical system that presents the material to airstreams as on the characteristics of the material itself. Several sources give details of desiccant equipment design and information about desiccant isotherm characteristics. Brunauer (1945) considers five basic isotherm shapes. Sing (1985) added a sixth isotherm shape and defined four types of hysteresis loops that are often identified with specific pore structures. Each shape is determined by the dominant sorption mechanisms of the desiccant, which give rise to its specific capacity characteristics at different vapor pressures. Isotherm shape can be important in designing the optimum desiccant for applications where a narrow range of operating conditions can be expected. Collier (1986, 1988) illustrates how an optimum isotherm shape can be used to ensure a maximum coefficient of performance in one particular air-conditioning desiccant application.
5.
DESICCANT LIFE
The useful life of desiccant materials depends largely on the quantity and type of contamination in the airstreams they dry. In commercial equipment, desiccants last from 10 000 to 100 000 h or longer before they need replacement. Normally, three mechanisms cause loss of desiccant capacity: (1) change in desiccant sorption characteristics through chemical reactions with contaminants, (2) loss of effective surface area through clogging or hydrothermal degradation, or (3) clogging or masking the pore system’s surface by contaminants. Liquid absorbents are more susceptible to chemical reaction with airstream contaminants other than water vapor than are solid adsorbents. For example, certain sulfur compounds can react with LiCl to form lithium sulfate, which is not a desiccant. If the concentration of sulfur compounds in the airstream were below 10 mg/kg and the desiccant were in use 24 h a day, capacity reduction would be
Sources for isotherms presented in the figure include PSSASS: Czanderna (1988) Lithium chloride: Munters Corporation: Cargocaire Division and Kathabar, Inc. Triethylene glycol: Dow Chemical Corporation Silica gel: Davison Chemical Division of W.R. Grace Co. Activated carbon: Calgon Corporation Activated alumina: LaRoche Industries Inc. Molecular sieve: Davison Chemical Division of W.R. Grace Co.
Fig. 7
Sorption Isotherms of Various Desiccants
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32.6
2017 ASHRAE Handbook—Fundamentals (SI)
future air-conditioning systems where indoor air quality is especially important. The behavior of different desiccant and vapor mixtures is complex, but in general, pollutant sorption reactions can be classified into five categories: • • • • •
Humidity-neutral sorption Humidity-reduced sorption Humidity-enhanced sorption Humidity-pollutant displacement Desiccant-catalyzed pollutant conversion
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Humidity-reduced sorption is illustrated by the behavior of water vapor and chloroform on activated carbon. Sorption is humidityneutral until relative humidity exceeds 45%, when the uptake of chloroform is reduced. The adsorbed water blocks sites that would otherwise attract and hold chloroform. In contrast, water and carbonyl chloride mixtures on activated carbon demonstrate humidityenhanced sorption (i.e., sorption of the pollutant increases at high relative humidities). Hines et al. (1991) attribute this phenomenon to the high water solubility of carbonyl chloride.
Collier, R.K. 1986, 1988. Advanced desiccant materials assessment. Research Report 5084-243-1089. Phase I-1986, Phase II-1988. Gas Research Institute, Chicago. Czanderna, A.W. 1988. Polymers as advanced materials for desiccant applications. Research Report NREL/PR-255-3308. National Renewable Energy Laboratory, Golden, CO. Davidson Chemical Division of W.R. Grace Co. 1958. Adsorption and dehydration with silica gel. Technical Bulletin 202. Dow. 1981. Guide to glycols. Dow Chemical Corporation, Organic Chemicals Division, Midland, MI. Foote Mineral. 1988. Lithium chloride technical data. Bulletin 151. Foote Mineral Corporation, Exton, PA. Harriman, L.G., III. 2003. The dehumidification handbook, 2nd ed. Munters Corporation, Amesbury, MA. Hines, A.J., T.K. Ghosh, S.K. Loyalka, and R.C. Warder, Jr. 1991. Investigation of co-sorption of gases and vapors as a means to enhance indoor air quality. ASHRAE Research Project 475-RP and Gas Research Institute Project GRI-90/0194. Gas Research Institute, Chicago. Oscic, J., and I.L. Cooper. 1982. Adsorption. John Wiley & Sons, New York. Sing, K.S.W. 1985. Reporting physisorption data for gas/solid systems with special reference to the determination of surface area and porosity (Recommendations 1984). Pure and Applied Chemistry 57:603-620.
REFERENCES
BIBLIOGRAPHY
ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal .ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore.
Adamson, A.W., and A.P. Gast. 1997. The physical chemistry of surfaces, 6th ed. John Wiley & Sons, New York. Falcone, J.S., Jr., ed. 1982. Soluble silicates. Symposium Series 194. American Chemical Society, Washington, D.C. Lowel, S., J.E. Shields, M.A. Thomas, and M. Thommes (eds.). 2004. Characterization of porous solids and powders: Surface area, pore size and density. Kluwer Academic Publishers, Dordrecht, the Netherlands. Ruthven, D.M. 1984. Principles of adsorption and adsorption processes. John Wiley & Sons, New York. SUNY Buffalo School of Medicine. Effects of glycol solution on microbiological growth. Niagrara Blower Report No. 03188. Valenzuela, D., and A. Myers. 1989. Adsorption equilibrium data handbook. Simon & Schuster/Prentice-Hall, Englewood Cliffs, NJ.
Battelle. 1971. Project No. N-0914-5200-1971. Battelle Memorial Institute, Columbus, OH. Brunauer, S. 1945. The adsorption of gases and vapors, vol. I. Princeton University Press, Princeton, NJ. Quoted and expanded in The physical chemistry of surfaces, by Arthur W. Adamson. John Wiley & Sons, New York, 1982. Bry-Air. 1986. MVB series engineering data. Bry-Air Inc., Sunbury, OH.
Related Commercial Resources
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Related Commercial Resources CHAPTER 33
PHYSICAL PROPERTIES OF MATERIALS
V
properties vary with temperature, material density, and composition. The references document the source of the values and provide more detail or values for materials not listed here. The preparation of this chapter is assigned to TC 1.3, Heat Transfer and Fluid Flow.
ALUES in the following tables are in consistent units to assist the engineer looking for approximate values. For data on refrigerants, see Chapter 29; for secondary coolants, see Chapter 31. Chapter 26 gives more information on the values for materials used in building construction and insulation. Many
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Table 1 Properties of Vapor
Material
Relative Molecular Mass
Normal Boiling Point, °C
Critical Temperature, °C
Critical Pressure, kPa
Alcohol, ethyl Alcohol, methyl Ammonia Argon Acetylene Benzene Bromine Butane Carbon dioxide Carbon disulfide
46.07a 32.04a 17.03a 39.948a 26.04a 78.11a 159.82a 58.12a 44.01a 76.13h
78.6a 65.0a –33.2a –185.9* –83.7a 80.2a 58.8a –0.5a –78.5a 46.3h
243.2b 240.1b 132.6b –122.5* 36.1b 289.6d 58.8d 152.1d 31.1d 278.9h
6 394b 7 977b 11 300b 4 860b 6 280b 4 924d 10 340d 3 797d 7 384d 7 212h
Carbon monoxide Carbon tetrachloride Chlorine Chloroform Ethyl chloride Ethylene Ethyl ether Fluorine Helium Hydrogen
28.01a 153.84g 70.91a 119.39h 64.52h 28.03h 74.12h 38.00h 4.0026a 2.0159a
–191.5a 76.6h –34.7a 61.8h 12.4h –103.7h 34.7h –187.0h –269.0i –253.1i
–140.3d 283.3h 144.1d 263.4h 187.3h 10.0h 192.7h –129.2h –267.9h –240.0i
3 500d 4 560h 7 710d 5 470h 5 270h 5 120h 3 610h 5 580h 229i 1 316i
Hydrogen chloride Hydrogen sulfide Heptane (m) Hexane (m) Isobutane Methane Methyl chloride Naphthalene Neon Nitric oxide
36.461a 34.080a 100.21a 86.18a 58.12f 16.04a 50.49a 128.19a 20.183a 30.01a
–84.9a –60.8a 98.5a 66.9a –11.6* –164.0a –24.3a 218.0* –247.0a –152.0a
51.4d 100.4d 266.8b 234.8d 135.1j –81.8j 143.2j 469.1j –228.8j –92.9j
8 260d 9 012d 2 720b 3 030d 3 648j 4 641b 6 678b 3 972j 2 698j 6 546j
–195.8a –88.5a
–146.9j 36.4j 158.3j –118.6* 196.7j 418.9b 96.7* 91.8l 156.9b 374.0*
3 394b 7 235j 10 133j 5 043* 3 375j 6 130b 4 248* 4 622l 7 874b 22 064*
Nitrogen Nitrous oxide Nitrogen tetroxide Oxygen n-Pentane Phenol Propane Propylene Sulfur dioxide Water vapor
28.01a 44.01a 92.02a 31.9977* 72.53a 74.11b 44.09g 42.08b 64.06b 18.02b
–183.0a 36.1* 181.4b –42.1g –47.7l –10.0b 100.0m
Density, kg/m3
7.72b 1.785b 1.17b 2.68e (80) 6.1f (59) 2.69g 1.97g
1.25d 3.22d 2.872b 1.25b 1.637b 0.178i 0.0900i 1.640b 1.54b 3.4k 3.4k 2.47s (21) 0.718b 2.307b
2.6k 2.02g 1.92l 2.93b 0.598m
Specific Heat, J/(kg ·K) 1520j 1350j 2200aa 523c 1580a 1300e (80) 230f (100) 1580aa 840g 599.0p (27)
Copyright © 2017, ASHRAE
Viscosity, Pa·s
0.013a 0.0301r 0.0221b 0.016a 0.0187b 0.0071e 0.0061a 0.014a 0.015a
14.2j (289) 14.8j (272) 9.30aa 21.0a 9.34a 7.0a 17a 7.0a 14h
0.0230a
0.0254j 0.142aa 0.168aa
17a 16.0j 12a 16j 16.0q 9.60aa 11.3q 37j 19.0aa 8.40aa
800j 996j 1990j 1880j 1570aa 2180aa 770aa 1310q (25) 1030aa 996j
0.0131j 0.0130j 0.0185j 0.0168j 0.014aa 0.0310aa 0.0093aa
13.3j 11.6j 7.00j 7.52j 6.94aa 10.3aa 10.1aa
0.0464aa
30.0aa 29.4j
1040j 850j 842p (27) 913j 1680a (27) 1400k 1571j (4.5) 1460aa 607l 2050aa
0.0240aa 0.01731j (26.8) 0.0401r (55) 0.0244aa 0.0152j (26.8) 0.017k 0.015j 0.014aa 0.0085j 0.0247m
16.6aa 22.4j
1100f 862q (27) 490a 528j 1780r 1470aa 2470h (35) 812j 5192aa 14 200j
*Data source unknown. Notes: 1. Properties at 101.325 kPa and 0°C, or the saturation temperature if higher than 0°C, unless otherwise noted in parentheses. 2. Superscript letters indicate data source from the References section.
33.1
Thermal Conductivity, W/(m ·K)
0.0080a 0.014r 0.00872j 0.0176aa
19.1aa 11.7j 12k 7.40j 8.06aa 11.6j 12.1aa
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
33.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 2
Name or Description
Normal Enthalpy Specific Heat, Boiling of cp Point, VaporizaJ/ Temp., °C at tion, (kg·K) °C 101.325 kPa kJ/kg
Licensed for single user. © 2017 ASHRAE, Inc.
Acetic acid Acetone Allyl alcohol n-Amyl alcohol Ammonia Alcohol, ethyl Alcohol, methyl Aniline Benzene Bromine n-Butyl alcohol n-Butyric acid Calcium chloride brine (20% by mass) Carbon disulfide Carbon tetrachloride Chloroform n-Decane Ethyl ether Ethyl acetate Ethyl chloride Ethyl iodide Ethylene bromide Ethylene chloride Ethylene glycol Formic acid Glycerin (glycerol) Heptane Hexane Hydrogen chloride Isobutyl alcohol Kerosene Linseed oil Methyl acetate Methyl iodide Naphthalene Nitric acid Nitrobenzene
118.6a 56.3a 97.1a 138.2i –33.2a 78.6a 65.0a 184.4a 80.2a 58.8a 117.6a 163.6a
405.0b 532.4b 684.1b 503.1b 1357b 854.8b 1100b 434.0b 394.0h 185d 591.5h 504.7h
46.3a 76.7a
346.1h 195h
61.3v 174.1b 34.5v 77.2v 12.4j 72.3a 131.6a 83.6a 198.1a 99.8a
2180b 26 to 95 2150b 3 to 23 2740b 21 to 96 4601b 0 2840b 0 to 98 b 2510 15 to 20 2140b 8 to 82 1720h 20 448f 20 2350f 20 2150f 20 3110i 20
Octane Petroleum
124.8b
Toluene (C6H5CH3) Turpentine Water Xylene [C6H4(CH3)2] Ortho Meta Para Zinc sulfate and water 10% by mass 1% by mass
20 20 20 20 0 0 20 20
562v
20
306.3b 230 to 384w 357.3h 413.6f
230v 451v
20 20
9.90f 28.7f 14.0f
20 20 20
29.7f
20
17 800f
20 20 20
195b 98.0b 112b 322.40b 108b 99.3a 114b 126h 66.30d 125b 126a
202b 98.60v 119b 69.04a 57.73a 88.43a 181.10a 276.54a
20 20
409a 320d
486f 2000n
20 20
1950f
20
1680f 1700v 1450b
m.p. 20 20
3 910f 2 480b 42 900b 389f 500f 901b 910k 2 150b
20 20 20 20 20 m.p. 20 20
2100b 2000 to 3000w 2330h 1980h
20 20
20 20
180.70b
20 20
562b 7900 to 1.2106w 226d 1 102a
20 20
117h
3110x 3620x
20 20
1 570x 1 180x
20 20
3610b
20
1400b 1460v 1600v
20 20 20
22 000b 21 000v 25 000v 587v
363b
1690v
148.9a 100.0*
286v 2257m
1700b 4180m
20 20 20
142.9b 137.9b 136.9b
347b 342b 340b
1720b 1670b 1640b 3700b 3300b
1049a 791a 853.9a 817.9f 696.8b 789.2a 791.3a 1021a 879d 3119f 811a 964a 1180i
57.70d 1260d 29.80d 1590d
2220j 2250j
108.9b
*Data source unknown. †Approximate solidification temperature.
0.17b 20 53.3a 0.1761b 30 53.3a 0.180b 25 to 30 53.3a 0.16b 30 13.3a 0.50b –15 to 30 53.3a 0.182b 20 13.3a 0.215b 20 13.3a 0.173b –2 to 20 1.3a 0.147h 20 10d 0.122a 25 22.0d 0.15h 20 0.7d 0.16h 12 0.09d 0.574i 20
980v 2000b 351v 2260v 427.5v 1950v 385.9f (20) 1540f 191f (71) 1540f 231f (99) 729f 365.8f (153) 1260f 800.1f (344) 502.0f (216) 2200f
412f 192f 316f 628v 330b
n-Pentane 35.1a Propionic acid 140.2a Sodium chloride brine 20% by mass 103.9a 10% by mass 100.9a Sodium hydroxide and water (15% by mass) 100.7v Sulfuric acid and water 100% by mass 286.8v 95% by mass 300.9v 90% by mass 259.1v
20 20 20 23 –33 20 20 20 20 20 20 20 20 20 20
321f 337f 444f 579f
56.1a 41.6a 209.8a 85.1v 209.9b
1 222f 331f 1 363f 4 004f 266f 1 194f 592.8f 4 467.0f 653a 988a 2950f 1 540a 2 000i 360a 967a
179.9*
Vapor Pressure
W/ Temp., (m·K) °C
20 20
20
Thermal Conductivity
Enthalpy Density of Temp., Temp., Fusion, °C kJ/kg kg/m3 Pa·s °C Viscosity
1000i 842f
247v
97.5a 65.9a –85.9a 107.1a 204 to 293b
Properties of Liquids
140b 150b 54.9f
151b 166v 93.69v
20 20 20 71.90b
20 20 20 15 –45 20 20 20 20 20 20 20 20
kPa
Temp., °C
Freezing Point, °C
99 40 80 86 –45 35 21 69 20 20 20 20
16.7a –95.4a –129.0a –79.0a –77.8a 117.3a –97.8a –6.2a 5.9a –7.2a –90.2a –6.2a –16.2i
20 20
0.16b 0.11j
30 20
39.3d 12d
20 20
–111.2a –22.8a
1489v 730b 714.6v 838v 897.8a 1935.8a 2179.3a 1235a 1109a 1219a
20 20 20 20 20 20 20 20 20 20
0.13v 0.15b 0.14b 0.175b 0.310f 0.370f
20 20 20 20 1 30
0.173f 0.180a
20 –2
21.3v 0.17b 58.7v 9.6b 53.3y 13.3y 1.3y 8.0y 0.1y 5.3y
20 20 20 20 12 18 19 18 53 23
–63.3v –29.8b –116.3v –82.4v –136.4a –108.0* 9.6a –35.4a –10.8a 7.4a
1261a
20
0.195a
20
0.1a
51
18.9a
684a 658a 1190d 801f 820a 920d 971a 2270a 976y 1512v 1200b
20 20 b.p. 20 20 20 20 20 m.p. 20 20
0.128j 0.125j
20 20
4.73y 16.00y
20 20
0.14f 0.15n
20 20
1.3y
20
–92.2a –96.2a –115.8a –109.0a
0.16f
20
0.28v 1.7b
20 20
22.64y 42.7y 0.291b 0.236v 0.001b
20 20 20 20 20
–24.9a –99.2†a –67.5a 79.3a –42.7v 4.8b
20 20
0.15b
20
0.056b
20
–57.5b
20 20
0.11h 0.173*
20 12
20 20
–130.8a –21.8a
1150x 1070x
20 20
0.583x 0.593x
20 20
20 20
–17.4x –7.4x
1150b
20
1833v 1836v 1816v
20 20 20
0.38b
20
867b
20 20 20
0.16b 0.13b 0.602m
20 20 20
703b 640 to 1000w 626a 992a
56.7d 0.4d 0.076x 0.087x
–22.0b
546b 988m
20 20 20
333.8b
863b 998.20m
20 20 20
831b 628b 670b
20 20 20
128b 109b 161b
881b 867b 862b
20 0 20
1.6b 1.6b
20 20
20 20
1 570a 1 100a
20 20
1110r 1010r
20 20
0.583a 0.598a
20 20
0.001b 0.001v 0.001v
20 20 20
9.6b –29.2v –10.5v
0.12b
20
–96.0b
2.34*
20
–1.0m
0.0260b 0.0290b 0.0300b
20 20 20
–26.2b –48.2b 11.9b
Notes: Superscript letters indicate data source from the section on References. m.p. = melting point b.p. = boiling point
–2.3a –1.2a
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Physical Properties of Materials
33.3 Table 3 Properties of Solids
Material Description Aluminum (alloy 1100)
Licensed for single user. © 2017 ASHRAE, Inc.
Aluminum bronze (76% Cu, 22% Zn, 2% Al) Asbestos: Fiber Insulation Ashes, wood Asphalt Bakelite Bell metal Bismuth tin Brick, building
Specific Heat, J/(kg·K)
Density, kg/m3
896b
2 740u
221u
8 280u 2 400u 580b 640b 2 110b 1 300u
100u 0.170u 0.16b 0.071b (50) 0.74b 17u
1 970u
65.0* 0.7b
400n 1050b 800t 800t 920b 1500b 360t (50) 170* 800b
Emissivity
Brass: Red (85% Cu, 15% Zn) Yellow (65% Cu, 35% Zn) Bronze Cadmium Carbon (gas retort) Cardboard Cellulose Cement (Portland clinker) Chalk Charcoal (wood) Chrome brick
400u 400u 435t 230a 710a
8 780u 8 310u 8 490t 8 650f
1300b 670b 900t 840t 710b
54t 1 920i 2 290t 240a 3 200b
Clay Coal Coal tars Coke (petroleum, powdered) Concrete (stone) Copper (electrolytic) Cork (granulated) Cotton (fiber) Cryolite (AlF3 ·3NaF) Diamond
920b 1000b 1500b (40) 1500b (400) 653b (200) 390u 2030t 1340u 1060b 616b
1 000t 1 400t 1 200b 990b 2 300b 8 910u 86t 1 500u 2 900b 2 420t
Earth (dry and packed) Felt Fireclay brick Fluorspar (CaF2) German silver (nickel silver) Glass: Crown (soda-lime) Flint (lead) Heat-resistant “Wool” Gold Graphite: Powder Impervious Gypsum Hemp (fiber)
1 829b (100) 880b 400u 750b 490b 840b 657b 131u 691* 670u 1080b 1352.3u
Ice: 0°C –20°C Iron: Cast Wrought Lead Leather (sole) Limestone Linen Litharge (lead monoxide) Magnesia: Powdered Light carbonate Magnesite brick
2040t 1950t 500v (100)
Magnesium Marble Nickel, polished Paints: White lacquer White enamel Black lacquer Black shellac Flat black lacquer Aluminum lacquer
1000b 880b 440u
129u 909b 230b 980b (100) 930b (100)
500t
330b 1 790t 3 190v 8 730u 2 470u 4 280u 2 230t 52.0t 19 350u 1 870u 1 200b 1 500u 921b 210t
7 7 700b 11 300u 1 000b 1 650b 7 850b 796b 210b 2 530b 1 730u 2 600b 8 890u
1
000u
*Data source unknown. Notes: 1. Values are for room temperature unless otherwise noted in parentheses.
Thermal Conductivity, W/(m ·K)
150u 120u 29d (0) 92.9b 0.35b (–17) 0.07b 0.057t 0.029i 0.83* 0.05a (200) 1.2b 0.17f (0) 0.1b 0.95b (400) 0.93b 393u 0.048t (–5) 0.042u
Ratio
Surface Condition
0.09n 0.20n
Commercial sheet Heavily oxidized
0.93b
“Paper”
0.93* 0.030b 0.033b
Highly polished Highly polished
0.02d 0.81a
0.34*
About 120°C
0.072n
commercial, shiny
47t 0.064* 0.05b 1b (200) 1.1v 33u 1.0t (93) 1.4r 1.0t (93) 0.038t 297t 0.183* 130u 0.43b
0.41*
2.24b 2.44* 47.7b (54) 60.4b 34.8u 0.16b 0.93b 0.09b
0.95*
0.75n
At 1000°C
0.135n 0.94n
Polished Smooth
0.02n
Highly polished
0.75n 0.903b
On a smooth plate
0.435b 0.94b 0.28n 0.36* to 0.90
Freshly turned Dull, oxidized Gray, oxidized At 63 to 193°C
0.61b (47) 0.059b 3.8b (204) 160u 2.6b 59.5u
0.26u
0.55n 0.931b 0.045n 0.80n 0.91n 0.80n 0.91n 0.96n 0.39n
Oxidized Light gray, polished Electroplated On rough plate “Matte” finish On rough plate
2. Superscript letters indicate data source from the section on References.
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33.4
2017 ASHRAE Handbook—Fundamentals (SI) Table 3 Properties of Solids (Continued)
Material Description
Licensed for single user. © 2017 ASHRAE, Inc.
Paper Paraffin Plaster Platinum Porcelain Pyrites (copper) Pyrites (iron) Rock Salt Rubber, vulcanized: Soft Hard Sand Sawdust Silica Silver Snow: Freshly fallen At 0°C Steel (mild) Stone (quarried) Tar: Pitch Bituminous Tin Tungsten Wood: Hardwoods Ash, white Elm, American Hickory Mahogany Maple, sugar Oak, white Walnut, black Softwoods Fir, white Pine, white Spruce Wool: Fiber Fabric Zinc: Cast Hot-rolled Galvanizing
Specific Heat, J/(kg·K) 1300* 1670bb 130u 750* 549b 569b (69) 917u 2000* 800b 1320b 235u 500b 800b 2500v 233u 130u 1900/2700b
2390b See Table 4, Chapter 25 1360u 390u 390b
Emissivity Density, kg/m3 930b 749bb 2 110b 21 470u 260u 4 200b 4 970v 2 180u 1 100t 1 190t
0.13b 0.24b (0) 0.74b (75) 69.0u 2.2u
1 520b 190b 2 240v 10 500u 100y 500t 7 830b 1 500t 1 100u 1 200t 7 290u 19 400u
0.33b 0.05b 1.4t (93) 424u 0.598t 2.2t 45.3b
370/1100z 690z 580z 800z 550u 720z 750z 630z 350/740z 430z 430z 420z 1 300u 110/330u 7 130u 7 130b
0.11/0.255z 0.172z 0.153z
*Data source unknown. Notes: 1. Values are for room temperature unless otherwise noted in parentheses.
of chemistry and physics, 63rd ed. 1982-83. Chemical Rubber Publishing Co., Cleveland, OH. R.H. Chemical engineers’ handbook, 2nd ed., 1941, 5th ed., 1973. McGraw-Hill, New York. cTables of thermodynamic and transport properties of air, argon, carbon dioxide, carbon monoxide, hydrogen, nitrogen, oxygen and steam. 1960. Pergamon Press, Elmsford, NY. dAmerican Institute of Physics handbook, 3rd ed. 1972. McGraw-Hill, New York. eOrganick and Studhalter. 1948. Thermodynamic properties of benzene. Chemical Engineering Progress (November):847. fLange. 1972. Handbook of chemistry, rev. 12th ed. McGraw-Hill, New York. gASHRAE. 1969. Thermodynamic properties of refrigerants. hReid and Sherwood. 1969. The properties of gases and liquids, 2nd ed. McGraw-Hill, New York. iChapter 19, 1993 ASHRAE Handbook—Fundamentals. jT.P.R.C. data book. 1966. Thermophysical Properties Research Center, W. Lafayette, IN. kEstimated. lCanjar, L.N., M. Goldman, and H. Marchman. 1951. Thermodynamic properties of propylene. Industrial and Engineering Chemistry (May):1183. mASME steam tables. 1967. American Society of Mechanical Engineers, New York. nMcAdams, W.H. 1954. Heat transmission, 3rd ed. McGraw-Hill, New York. bPerry,
0.1t 0.16t
0.88v 0.71u 64.9u 201u
0.13u 0.187z 0.176z
Ratio
Surface Condition
0.92b
Pasted on tinned plate
0.91b 0.054b 0.92b
Rough Polished Glazed
0.86b 0.95b
Rough Glossy
0.02n
Polished and at 227°C
0.12n
Cleaned
0.06h 0.032n
Bright and at 50°C Filament at 27°C
0.90n
Planed
0.05n
Polished
0.23n
Fairly bright
0.11/0.16z 0.12z 0.11z 0.11z 0.036/0.063u 110u 110b
2. Superscript letters indicate data source from the section on References. oStull,
REFERENCES aHandbook
Thermal Conductivity, W/(m ·K)
D.R. 1947. Vapor pressure of pure substances (organic compounds). Industrial and Engineering Chemistry (April):517. pJANAF thermochemical tables. 1965. PB 168 370. National Technical Information Service, Springfield, VA. qPhysical properties of chemical compounds. 1955–61. American Chemical Society, Washington, D.C. rInternational critical tables of numerical data. 1928. National Research Council of USA, McGraw-Hill, New York. sMatheson gas data book, 4th ed. 1966. Matheson Company, Inc., East Rutherford, NJ. tBaumeister and Marks. 1967. Standard handbook for mechanical engineers. McGraw-Hill, New York. uMiner and Seastone. Handbook of engineering materials. John Wiley and Sons, New York. vKirk and Othmer. 1966. Encyclopedia of chemical technology. Interscience Division, John Wiley and Sons, New York. wGouse and Stevens. 1960. Chemical technology of petroleum, 3rd ed. McGraw-Hill, New York. xSaline water conversion engineering data book. 1955. M.W. Kellogg Co. for U.S. Department of Interior. yTimmermans, J. Physicochemical constants of pure organic compounds, 2nd ed. American Elsevier, New York. zWood handbook. 1955. Handbook No. 72. Forest Products Laboratory, U.S. Department of Agriculture. aaASHRAE. 1976. Thermophysical properties of refrigerants. bbLane, G. ed. 1986. Solar heat storage: Latent heat materials, Vol II—Technology. CRC Press, Chicago.
Related Commercial Resources
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Related Commercial Resources CHAPTER 34
ENERGY RESOURCES CHARACTERISTICS OF ENERGY AND ENERGY RESOURCE FORMS ........................................... On-Site Energy/Energy Resource Relationships ..................... Summary .................................................................................. ENERGY RESOURCE PLANNING......................................... Integrated Resource Planning (IRP) .......................................
34.1 34.2 34.3 34.3 34.3
NERGY used in buildings and facilities is responsible for 30 to 40% of the world's energy use, significantly impacting world energy resources. ASHRAE’s work to reduce energy consumption in the built environment is equally as important as research on new, more sustainable energy sources in helping ensure a reliable and secure supply of energy for future generations. Many governmental agencies regulate energy conservation, often through the procedures to obtain building permits. Required efficiency values for building energy use strongly influence selection of HVAC&R systems and equipment and how they are applied. More information on sustainable design is available in the ASHRAE GreenGuide (2013) and in Chapter 35.
Licensed for single user. © 2017 ASHRAE, Inc.
E
Tradable Emission Credits....................................................... OVERVIEW OF GLOBAL ENERGY RESOURCES ................ World Energy Resources.......................................................... Carbon Emissions .................................................................... U.S. Energy Use ....................................................................... U.S. Agencies and Associations ...............................................
34.4 34.4 34.4 34.7 34.7 34.9
water. The conversion equipment is a boiler or a furnace in lieu of a generator, and conversion usually occurs on a project site rather than off site. (District heating or cooling is an exception.) Inefficiencies of fossil fuel conversion occur on site, whereas inefficiencies of most electricity generation occur off site, before the electricity arrives at the building site. (Cogeneration is an exception.) Sustainability is an important consideration for energy use. The United Nations’ Brundtland Report (UN 1987) stated that the development of the built environment is sustainable if it “meets the needs of the present without compromising the ability of future generations to meet their own needs.” More information is in Chapter 35.
Forms of On-Site Energy
1. CHARACTERISTICS OF ENERGY AND ENERGY RESOURCE FORMS The HVAC&R industry deals with energy forms as they occur on or arrive at a building site. Generally, these forms are fossil fuels (natural gas, oil, and coal) and electricity. Solar and wind energy are also available at most sites, as is biomass. Geothermal energy is available at some locations.
Fossil Fuels and Electricity Most on-site energy for buildings in developed countries involves electricity and fossil fuels as energy sources. Both fossil fuels and electricity can be described by their energy content (joules). This implies that energy forms are comparable and that an equivalence can be established. In reality, however, they are only comparable in energy terms when they are used to generate heat. Fossil fuels, for example, cannot directly drive motors or energize light bulbs. Conversely, electricity gives off heat as a by-product regardless of whether it is used for running a motor or lighting a light bulb, and regardless of whether that heat is needed. Thus, electricity and fossil fuels have different characteristics, uses, and capabilities aside from any differences in their derivation. Other differences between energy forms include methods of extraction, transformation, transportation, and delivery, and characteristics of the resource itself. Natural gas arrives at the site in virtually the same form in which it was extracted from the earth. Oil is processed (distilled) before arriving at the site; having been extracted as crude oil, it arrives at a given site as, for example, No. 2 oil or diesel fuel. Electricity is created (converted) from a different energy form, often a fossil fuel, which itself may first be converted to a thermal form. The total electricity conversion, generation, and distribution process includes energy losses governed largely by the laws of thermodynamics. Fossil fuels undergo a conversion process by combustion (oxidation) and heat transfer to thermal energy in the form of steam or hot The preparation of this chapter is assigned to TC 2.8, Building Environmental Impacts and Sustainability.
34.1 Copyright © 2017, ASHRAE
Fossil fuels and electricity are commodities that are usually metered or measured for payment at the facility’s location. Solar or wind energy is freely available but does incur cost for the means to use it. Geothermal energy, which is not universally available, may or may not be a sold commodity, depending on the particular locale and local regulations. Chapter 34 of the 2015 ASHRAE Handbook— HVAC Applications has more information on geothermal energy. The term energy source refers to on-site energy in the form in which it arrives at or occurs on a site (e.g., electricity, gas, oil, coal). Energy resource refers to the raw energy that (1) is extracted from the earth (wellhead or mine-mouth), (2) is used to generate the energy source delivered to a building site (e.g., coal used to generate electricity), or (3) occurs naturally and is available at a site (solar, wind, or geothermal energy). Some on-site energy forms require further processing or conversion into more suitable forms for the particular systems and equipment in a building or facility. For instance, natural gas or oil is burned in a boiler to produce steam or hot water, which is then distributed to various use points (e.g., heating coils in air-handling systems, unit heaters, convectors, fin-tube elements, steam-powered cooling units, humidifiers, kitchen equipment) throughout the building. Although the methods and efficiencies of these processes fall within the scope of the HVAC&R designer, how an energy source arrives at a given facility site is not under direct control. On-site energy choices, when available, may be controlled by the designer based in part on the present and projected future availability of the resources. Energy sources used for heating may be natural gas, oil, coal, or electricity. Cooling may be produced by electricity, thermal energy, or natural gas. If electricity is generated on site, the generator may be driven by an engine or fuel cell that consumes fossil fuels or hydrogen on site, by a turbine using steam or gas directly, or by on-site renewable sources.
Nonrenewable and Renewable Energy Resources From the standpoint of energy conservation, energy resources can be classified as either (1) nonrenewable resources, which have definite, although sometimes unknown, limitations; or (2) renewable resources, which have the potential to regenerate in a reasonable
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34.2
2017 ASHRAE Handbook—Fundamentals (SI)
period. Resources used most in industrialized countries are nonrenewable (ASHRAE 2003). Note that renewable does not mean an infinite supply. For instance, hydropower is limited by rainfall and appropriate sites, usable geothermal energy is available only in limited areas, and crops are limited by the available farm area and competing non-energy land uses. Other forms of renewable energy also have supply limitations. Nonrenewable resources of energy include • • • •
Coal Crude oil Natural gas Uranium or plutonium (nuclear energy) Renewable resources of energy include
Licensed for single user. © 2017 ASHRAE, Inc.
• • • • •
Hydropower Solar Wind Geothermal Biomass (wood, wood wastes, and municipal solid waste, landfill methane, etc.) • Tidal power • Ocean thermal • Crops (for alcohol production or as boiler fuel)
Environmental Considerations The most widely recognized environmental impact from energy use in buildings is greenhouse gas emissions; carbon dioxide emissions is usually the most important greenhouse gas resulting from energy use in buildings. In this area, use of renewable resources and nuclear power generally results in no net greenhouse gas emissions, whereas fossil fuel energy use generally results in substantial greenhouse gas emissions. However, note that the important issue is the amount of greenhouse gas emissions released into the air, not the amount of a fuel used. It has been argued that some biomass energy sources are not really carbon-free sources, because they result in carbon dioxide releases that are not directly offset by carbon dioxide capture through growing vegetation to replace the biomass fuel. Even for fossil fuel energy use, research is ongoing for carbon capture and sequestration for emissions from fossil fuel electric power plants. If the carbon dioxide from a fossil fuel energy source is not released into the atmosphere, there are no greenhouse gas emissions resulting from the use of that fuel. In addition to greenhouse gases, there are also local air pollution issues from combustion of fuels. These include emissions of carbon monoxide, nitrogen oxides, sulfur dioxide, heavy metals, and particulates. These occur in the combustion of fossil fuels and biomass, and do not occur from the use of renewable energy (other than biomass) or nuclear power. Emissions of local air pollutants vary greatly, depending on design of the combustion equipment and controls technology used. Note that biomass energy sources may require similar mitigation measures to reduce local air emissions as would be required for fossil fuel energy sources.
1.1
ON-SITE ENERGY/ENERGY RESOURCE RELATIONSHIPS
An HVAC&R designer must select one or more forms of energy. Most often, these are fossil fuels and electricity, although installations are sometimes designed using a single energy source (e.g., only a fossil fuel or only electricity). Solar energy normally impinges on the site (and on the facilities to be put there), so it affects the facility’s energy consumption. The designer must account for this effect and may have to decide
whether to make active use of solar energy. Other naturally occurring and distributed renewable forms such as wind power and geothermal (if available) might also be considered. The designer should be aware of the relationship between on-site energy sources and raw energy resources, including how these resources are used and what they are used for. The relationship between energy sources and energy resources involves two parts: (1) quantifying the energy resource units expended and (2) considering the societal effect of depletion of one energy resource (caused by on-site energy use) with respect to others.
Quantifiable Relationships As on-site energy sources are consumed, a corresponding amount of resources are consumed to produce that on-site energy. For instance, for every volume of No. 2 oil consumed by a boiler at a building site, some greater volume of crude oil is extracted from the earth. On leaving the well, the crude oil is transported and processed into its final form, perhaps stored, and then transported to the site where it will be used. Even though natural gas often requires no significant processing, it is transported, often over long distances, to reach its final destination, which causes some energy loss. Electricity may have as its raw energy resource a fossil fuel, uranium, or an elevated body of water (hydroelectric generating plant). Data are available to help determine the amount of resource use per delivered on-site energy source unit. In the United States, data are available from entities within the U.S. Department of Energy and from the agencies and associations listed at the end of this chapter. A resource utilization factor (RUF) is the ratio of resources consumed to energy delivered (for each form of energy) to a building site. Specific RUFs may be determined for various energy sources normally consumed on site, including nonrenewable sources such as coal, gas, oil, and electricity, and renewable sources such as solar, geothermal, waste, and wood energy. With electricity, which may derive from several resources depending on the particular fuel mix of the generating stations in the region served, the overall RUF is the weighted combination of individual factors applicable to electricity and a particular energy resource. Grumman (1984) gives specific formulas for calculating RUFs. There are great differences in the efficiency of equipment used in buildings. Although electricity incurs losses in its production, it is often much more efficient than direct fuel use at the building site, particularly for lighting or heat pump applications. Minimizing both energy cost and the amount of energy resources needed to accomplish a task effectively should be a major design goal, which requires consideration of both RUFs and end-use efficiency of building equipment. Although a designer is usually not required to determine the amount of energy resources attributable to a given building or building site for its design or operation, this information may be helpful when assessing the long-range availability of energy for a building or the building’s effect on energy resources. Fuel-quantity-to-energyresource ratios or factors are often used, which suggests that energy resources are of concern to the HVAC&R industry.
Intangible Relationships Energy resources should not simply be converted into common energy units (e.g., gigajoule) because the commonality gives a misleading picture of the equivalence of these resources. Other differences and limitations of each of the resources defy easy quantification. For instance, electricity used on a site can be generated from coal, oil, natural gas, uranium, or hydropower. The end result is the same: electricity at x kV, y Hz. However, the societal impact of a kilowatt-hour electricity generated by hydropower may not equal that of a kilowatt-hour generated by coal, uranium, domestic oil, or imported oil.
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Intangible factors such as safety, environmental acceptability, availability, and national interest also are affected in different ways by the consumption of each resource. Heiman (1984) proposes a procedure for weighting the following intangible factors: National/Global Considerations • Balance of trade • Environmental impacts • International policy • Employment • Minority employment • Availability • Alternative uses • National defense • Domestic policy • Effect on capital markets Local Considerations • Exterior environmental impact • Air • Solid waste • Water resources • Local employment • Local balance of trade • Use of distribution infrastructure • Local energy independence • Land use • Exterior safety Site Considerations • Reliability of supply • Indoor air quality • Aesthetics • Interior safety • Anticipated changes in energy resource prices
1.2
SUMMARY
In HVAC&R system design, the need to address immediate issues such as economics, performance, and space constraints often prevents designers from fully considering the energy resources affected. Today’s energy resources are less certain because of issues such as availability, safety, national interest, environmental concerns, and the world political situation. As a result, the reliability, economics, and continuity of many common energy resources over the potential life of a building being designed are unclear. For this reason, the designer of building energy systems must consider the energy resources on which the long-term operation of the building will depend. If the continued viability of those resources is reason for concern, the design should provide for, account for, or address such an eventuality.
2.
ENERGY RESOURCE PLANNING
The energy supplier (or suppliers) in a particular jurisdiction must plan for that jurisdiction’s future energy needs. For competitive energy markets where these decisions do not have high societal costs, these plans are made by energy suppliers and are not revealed to governmental authorities or the public more than is absolutely necessary, because of the advantage competitors could gain by this knowledge. For electricity (and, to a lesser extent, natural gas), significant societal issues are involved in energy resource planning decisions that cannot be made by energy suppliers without approval by many different groups. Issues include • Reliability, which is affected by the diversity of supply sources available. For gas, this includes the number of geographic supply sources and pipelines; for electricity, it includes the percentage of
generation from various fuel sources. Consider the projected future supply and reliability of energy resources, including the possibility of supply disruption by natural or political events, and the likelihood of future supply shortages, which could reduce reliability. • Reserve margins, or the ratio of total supply sources to expected peak supply source needs. Reserve levels that are too high result in waste of resources, higher environmental costs, and possibly poor financial health of the energy suppliers. Reserves that are too low result in volatile and very high peak energy prices and reduced reliability. • Land use. Energy production and transmission often require governmental cooperation to condemn private property for energy production and transmission facilities. Construction and maintenance are also regulated to protect wetlands, prevent toxic waste releases, and other environmental issues. Note that some energy regulation plans provide no guidance at all on energy supplies, through integrated resource planning (IRP) or other methods. Energy suppliers choose whether to expand their capacity, and what types of fuel those facilities use, based on their own assessment of the future profitability of that investment. In these markets, decisions are made with little societal input other than permitting and pollution control regulations, just as a decision might be made by a manufacturer in an industry such as steel or paper. This does not mean that decisions are made independent of larger societal issues, because laws and regulations (e.g., tradable emissions credits, renewable portfolio standards) factor into the economic considerations of competing suppliers. There is simply more direct planning by governmental authorities, and more opportunities for public input, in the integrated resource planning process typically done by regulated utility providers, as described in more detail in the following.
2.1
INTEGRATED RESOURCE PLANNING (IRP)
In regulated utility markets, integrated resource planning is commonly used for planning significant new energy facilities, especially for electricity. Steps include (1) forecasting the amount of new resources needed and (2) determining the type and provider of this resource. Traditionally, the local utility provider forecasts future needs of a given energy resource, then either builds the necessary facility with the approval of regulators or uses a standard offer bid to determine what nonutility provider (or the utility itself) would provide the new energy resource. Supplying new energy resources through either a standard bid process by a supplier or traditional utility regulation usually results in selection of the lowest-cost supply option, without regard for environmental costs or other societal needs. IRP allows a greater variety of resource options and allows environmental and other indirect societal costs to be given greater consideration. IRP addresses a wider population of stakeholders than most other planning processes. Many regulatory agencies involve the public in the formulation and review of integrated resource plans. Customers, environmentalists, and other public interest groups are often prominent in these proceedings. In deregulated energy markets, supplying markets with new energy resources is typically left up to competitive market forces. This has sometimes resulted in excessive reliance on one form of energy, such as natural gas generation. Another result has been highly volatile prices, when supply is not provided because of insufficient price signals, followed by much higher prices and energy shortages until new supply sources can be obtained (which may not be for several years because of the time required for construction and environmental approval processes). Energy efficiency and demand response programs are increasingly treated as an energy resource on a par with energy production options, with incentives and compensation provided for participants in these programs.
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Demand-side management (DSM) is a common option for providing new energy resources, especially for electricity. These are actions taken to reduce the demand for energy, rather than increase the supply of energy. DSM is desirable because its environmental costs are almost always lower than those of building new energy facilities. However, the following factors have caused a decline in the number of DSM programs:
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• Building and equipment codes and standards are a highly efficient form of DSM, reducing energy use with much lower administrative costs than programs that reward installation of more efficient equipment at a single site. However, they are more subtle than traditional DSM programs and may not always be recognized as a form of DSM. • Opening markets to competing suppliers makes it more difficult to administer and implement DSM programs. However, they are still possible if regulators wish to continue them, and set appropriate rules and regulations for the market to allow implementation of DSM programs.
(19%), the United States (16%), Russia (10%), and Saudi Arabia (7%). Together, they produced about 52% of the world’s energy production. (Note: for this and for similar graphs, the “Europe and Eurasia” region consists of the nations of western Europe plus the countries comprised by the former Soviet Union.) Total world energy production by resource type for 2004 and 2014 is shown in Figure 2. The greatest growth in energy production among major sources has been hydroelectricity, up nearly 33% in usage from 2004 to 2014, and coal, which has increased 30.3%, and natural gas, up 10.3%. Petroleum use only rose 7.1%. Nonhydroelectric renewables use more than doubled, but is still a small percentage of total world energy production (BP 2015). Crude Oil. World crude oil production was 88.7 million barrels (10 577 000 m3) per day in 2014. The biggest crude-oil-producing region in 2014 was the Middle East, with 32% of the total. Among individual countries, the United States (13.1%), Saudi Arabia (13.0%), and Russia (12.2%) were the leading countries, following by China and Canada, each with 4.8% of total production. Since
Many IRP participants may be interested in only one aspect of the process. For example, the energy industry’s main interest may be cost minimization, whereas environmentalists may want to minimize pollutant emissions and prevent environmental damage from construction of energy facilities. Participation by all affected interest groups helps provide the best overall solution for society, including indirect costs and benefits from these energy resource decisions.
2.2
TRADABLE EMISSION CREDITS
Increasingly, quotas and limits apply to emissions of various pollutants. Often, a market-based system of tradable credits is used with these quotas. A company is given the right to produce a given level of emissions, and it earns a credit, which can be sold to others, if it produces fewer emissions than that level. If one company can reduce its emissions at a lower cost than another, it can do so and sell the emissions credit to the second company and earn a profit from its pollution control efforts. In the United States, emissions quota and trading programs currently include sulfur dioxide (SO2) and nitrogen oxides (NOx), with plans to implement carbon dioxide (CO2) trading now under consideration, as well. In Europe, emissions trading for CO2 has been active for several years. Designers must be aware of any regulations concerning pollutant emissions; failure to comply with these regulations may result in civil or criminal penalties for designers or their clients. However, understand the options available under these regulations. The purchase or sale of emissions credits may allow reduced construction or building operations costs if the equipment can overcomply at a lower cost than the cost of another source of emissions to comply, or vice versa. In some cases, documentation of energy savings beyond what codes and regulations require can result in receiving emissions credits that may be sold later.
3.
Fig. 1 Energy Production Trends: 2004-2014 (Basis: EIA 2015)
OVERVIEW OF GLOBAL ENERGY RESOURCES 3.1
WORLD ENERGY RESOURCES
Data in this section are from the Statistical Review of World Energy 2015 (BP 2015).
Production Energy production trends, by leading producers and world regions, from 2004 to 2014 are shown in Figure 1. World primary energy production increased 21.2% from 2004 to 2014, because strong economic growth occurred in countries such as China, which increased its energy production more than 60% since 2004. The largest total energy producers in 2014 were China
Fig. 2 World Primary Energy Production by Resource: 2004 Versus 2014 (Basis: EIA 2015)
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2004, oil production increased 56% in the United States and about 25% in both China and Russia. Natural Gas. World production reached 3460.6 billion m3 in 2014, up 14.4% from the 2004 level. The biggest producers in 2014 were the United States (22%), Russia (17%), Qatar (5%), Iran (5%), and Canada (5%). Since 2004, natural gas production in Qatar more than quadrupled, and was up 79% in Iran and 39% in the United States. Coal. At 164.7 EJ in 2014, coal production was up 38.7% since 2004. Leading producers of coal were China (47%), the United States (13%), India (7%), and Indonesia (6%). Since 2004, Indonesia more than tripled coal production, China increased coal production 67%, and India increased 56%, whereas U.S. production fell by 11%.
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Reserves On January 1, 2015, estimated world reserves of crude oil and gas were distributed by world region as shown in Figures 3 and 4. Countries with the largest reported crude oil reserves are Venezuela (17.5%), Saudi Arabia (15.7%), Canada (10.2%), Iran (9.3%), and Iraq (8.8%). For natural gas, the countries with the largest reported reserves are Iran (18.2%), Russia (17.4%), Qatar (13.1%), Turkmenistan (9.3%), and the United States (5.2%). Note that several factors make crude oil and natural gas reserve estimates less precise than some other energy statistics: • For nations with national oil companies, reserve estimates are not verifiable by third parties, as is typically the case with the reserve estimates of publicly traded corporations. • Technologies used for extracting hydrocarbons from nontraditional reserves (e.g., tar sands, shale deposits) are rapidly changing. This can result in very large revisions to reserve estimates that often lag production changes. For example, approximately 10 years ago, Canadian oil reserves rose substantially due to inclusion of tar sands in the total. More recently, reported crude oil and natural gas reserves in the United States have lagged production
growth, because of uncertainty concerning the size of the total available resource that can be economically extracted. • The definition of proved resource includes a requirement that it can be economically extracted, which can result in substantial changes due to fluctuations in the price of oil and natural gas. World coal reserves as of January 1, 2015, are shown by region in Figure 5. The most plentiful reserves, as a percent of total, were in the United States (27%), Russia (18%), China (13%), Australia (9%), and India (7%). An important factor is the relative amount of these energy resources that has not yet been consumed. A standard measure is called proved energy reserves, which is the remaining known deposits that could be recovered economically given current economic and operating conditions. Dividing proved reserves by the current production rate gives the number of years of the resource remaining. Using this measure, the reserve-to-production ratio at the end of 2014 for crude oil was 56.9 years; for natural gas, 55.0 years; and for coal, 109.2 years. This does not mean that these resources will be depleted in that length of time: additional resources may be discovered in new areas, and improved technology may increase the amount of a resource that may be economically extracted. Also, the future rate of production and consumption may be higher or lower than current levels, which would decrease or increase the remaining years of a resource. However, reserve-to-production ratios provide insights into the limited nature of nonrenewable energy resources and the need to find alternatives, especially for resources with fewer years of remaining reserves.
Consumption
(Basis: BP 2016)
Data on world energy consumption are available only by type of resource rather than by total energy consumed. Petroleum. Consumption trends of the leading consumers from 1965 to 2014 are depicted in Figure 6. In 2014, the United States consumed far more petroleum than any other country: 19.9% of the world total. Other major petroleum-consuming countries were China (12.4%), Japan (4.7%), Russia (3.5%), Brazil (3.4%), and Saudi Arabia (3.4%). Natural Gas. In 2014, the two biggest natural gas producers (the United States and Russia) were also the two biggest consumers. Figure 7 depicts natural gas consumption by the leading consumer countries as a percentage of world consumption. The United States in 2014 consumed only 4% more than it produced, and Russia consumed 29% less than it produced. Of the other major natural-gas-consuming nations, China, Mexico, Germany, and the United Kingdom consumed at least 25% more than they produced. Iran, Saudi Arabia, and the United Arab Emirates were net consumers, but consumed less than 25% more than their production. Canada and Russia were the only major consuming nations that produced more than they consumed, but given current trends the United States is expected to be a net exporter of natural gas within a few years.
Fig. 4 World Natural Gas Reserves: 2015
Fig. 5 World Recoverable Coal Reserves: 2015
(Basis: BP 2016)
(Basis: BP 2016)
Fig. 3
World Crude Oil Reserves: 2015
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Fig. 6 World Petroleum Consumption: 2015 (Basis: BP 2016)
Fig. 9 Coal Consumption in United States, China, and India, 1980-2014
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(Basis: BP 2015)
Fig. 7 World Natural Gas Consumption: 2014 (Basis: BP 2015)
Fig. 10 World Electricity Generation by Resource: 2002 and 2012 (EIA 2012)
Fig. 8 World Coal Consumption: 2014 (Basis: BP 2015)
Coal. The two largest coal producers in 2014 (China and the United States) were also the two largest consumers. China is by far the largest coal consumer, with consumption more than four times that of the United States in 2014. Figure 8 depicts the percentage of world consumption by the leading consumers during 2014. Since 1980, world coal consumption has doubled, largely because of remarkable growth in coal consumption in China. Figure 9 shows the change in coal consumption since 1980 for the United States, China, and India. Over this span, consumption by China increased 544%, in India 535%, and in the United States 17%. However, coal consumption has been declining in the United States over the past 10 years, and growth in coal consumption in China has slowed dramatically since 2010. Growth of coal consumption in India has continued. Although not shown on the graph, coal consumption is declining in most European countries. Electricity. Figure 10 shows the world’s electricity generation by energy resource in 2002 and 2012. (Note: unlike other energy statistics, which generally include 2014 data, the most recent year in
Fig. 11 World Electricity Generation 2014 (Basis: BP 2015)
which generation by fuel type is available is 2012.) Fossil fuel generation increased 45.3%, hydroelectric generation increased 39.9%, and nuclear generation decreased 7.9%, although this level of decrease may be temporary, driven by the shutdown of almost all nuclear generation in Japan during 2012. Nonhydroelectric renewable generation increased by the largest percentage (262.3%), but is still substantially less than the other resources. Figure 11 shows total electricity generation by geographic region for 2014. China and the United States produced the most electricity, followed by India, Russia, and Japan. Per Capita. Figure 12 compares the per capita energy consumption of selected countries for 2014. As is apparent, per capita energy
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Fig. 12 Per Capita Energy Consumption by Selected Countries: 2011
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(Basis: BP 2015; CIA 2016)
consumption in countries with extreme weather, whether cold or hot, tends to be highest; also, the level in more developed countries is vastly different from that in less developed countries and differs considerably even among the more developed countries. Note that, although China’s total energy use has grown very rapidly in recent years, on a per capita basis it is still substantially below the levels of more developed countries.
3.2
CARBON EMISSIONS
Worldwide carbon emissions from burning and flaring fossil fuels rose 21.8% from 2004 to 2014. Other sources of greenhouse gas emissions are included under international treaties; however, this section only shows the portion from marketed fossil fuel production, and so does not include greenhouse gas emissions from energy production, methane emission from various sources, or releases of high-global-warming-potential (GWP) chemicals such as refrigerants. Total carbon emissions were 35.499 billion tonnes of carbon dioxide in 2014, up from 29.143 billion tonnes in 2004. Figure 13A shows the changes in carbon emissions from burning fossil fuels from 2004 to 2014 for the total world and for selected countries. Greenhouse gas emissions dropped 7.4% in the United States, and dropped 18.8% in the European Union. Countries such as China, India, and Saudi Arabia showed the largest increases. Note that, although developing countries have the highest growth rates, their per capita carbon emissions are much less than in wealthier nations. A graph of per capita carbon emissions would look very similar to Figure 12, which shows per capita energy consumption of selected countries. Carbon emissions from fossil fuel use alone were not readily available for the EU as a whole. The data shown are for 22 of the 28 countries in the EU for which fossil fuel data was available by country. The six countries not included are very small, so the difference between the emissions shown and the EU total is not substantial.
3.3
Fig. 13 World Carbon Emissions (Basis: EIA 2015)
U.S. ENERGY USE
Per Capita Energy Consumption Figure 14, based on data from BP (2012), shows the growth in per capita energy use since 1965 for the world and for the United States. As can be seen in the graph, United States per capita energy use was approximately six times the world average in 1965, dropping to five times in 1985, and is now about four times the world average. Although the world average per capita energy use has steadily risen, U.S. per capita energy use peaked in the late 1960s, rose at a slow rate in the 1990s, and has generally declined in recent years.
Fig. 14 Per Capita United States Energy Consumption (Basis: BP 2015)
Projected Overall Energy Consumption The International Energy Outlook is a projection of total world energy use developed by the United States Department of Energy Information Administration (EIA 2016). Figure 15 shows projected energy consumption through the year 2040.
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Fig. 15 Projected World Energy Consumption by Resource Petroleum consumption is projected to increase by 1.10% percent per year, resulting in increased usage of 30% by 2040. Natural gas usage is projected to increase 1.90% per year, for a 64% increase by 2040. Coal is projected to increase 0.60% per year, for a 12% total increase. Nuclear power is projected to grow 2.3% yearly, for an 87% increase. “Other” energy sources, which includes renewables, are projected to grow at the highest rate, 2.60% annually, for an 87% increase by 2040. The Annual Energy Outlook is the basic source of data for projecting energy use in the United States (EIA 2015). Figures 16 and 17 summarize data from this source. Information is available in greater detail for United States projected energy use than for world energy use. EIA (2016) forecasts energy trends based on macroeconomic growth scenarios, which include a variety of energy price and economic growth assumptions. Figures 16 and 17 (the baseline or reference case) assume average annual growth of the real gross domestic product (GDP) at 2.2%, of the labor force at 0.7%, and of productivity at 1.7%. To be policy neutral, the forecast also assumes that all federal, state, and local laws and regulations in effect at the end of 2015 remain unchanged through 2040. Note: based on laws in effect at the end of 2015, this forecast assumes the implementation of the United States’ Clean Power Plan, which proposed to reduce the allowable greenhouse gas emissions for existing electric power plants by 30% compared to 2005 levels (EPA 2016). However, later in 2016 this proposal was suspended, pending review by the United States Supreme Court. An alternative scenario developed by EIA assumed that the Clean Power Plan was not implemented, and projected less renewable energy growth and a lower rate of decline in coal consumption. Figure 16 shows energy use by major end-use sector (i.e., residential, commercial, industrial, and transportation). HVAC&R engineers are primarily concerned with the first three sectors. Figure 17 shows energy consumption by type of resource. The following observations apply to the overall picture of projected energy use in the United States over the next two decades (Figures 16 and 17): • As discussed previously, the base forecast assumes implementation of the Clean Power Plan for reducing greenhouse gas emissions for existing electric power plants, which is currently under legal review. Carbon emissions from energy use are projected to decline slowly, decreasing by an average of 0.2% per year through 2040 because of projected shifts in fuel use from coal to less carbon-intensive fuels (gas and renewables), and because of higher efficiency standards for appliances and commercial equipment, stronger building codes, and higher required mileage from vehicles. Per capita carbon emissions are projected to decline 0.8% per year.
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Fig. 16
Projected Total U.S. Energy Consumption by End-Use Sector (Basis: EIA 2016)
Fig. 17
Projected Total U.S. Energy Consumption by Resource (Basis: EIA 2016)
• Crude oil prices are expected to rise at an annual rate of 3.9% more than inflation. This high rate of increase is due to a low starting point for crude oil prices in 2015, rather than a fundamental change in projected supply and demand conditions. • The wellhead price of natural gas was projected to rise at an annual rate of 2.5% more than inflation from 2015 through 2040. However, the starting point is about $2.46/gigajoule wellhead price, well below historic prices of recent years. Natural gas prices are projected to reach $4.74 per gigajoule in 2024, and stay stable near that price until 2040. This forecast assumes a long-term stable price for natural gas, tracking the rate of general inflation, due to abundant supplies of natural gas that result in 20% of United States natural gas production being exported by 2040. For coal used as a fuel, consumption is expected to decrease by 1.6% per year, resulting in a decline of about 40% in consumption compared to 2014 levels. The price of coal used as a fuel is expected to grow at an annual rate of 0.3% more than inflation over the same period. This is less than previous forecasts, and is a result of lower forecasted demand. • Commercial electricity prices are projected to increase at the rate of inflation over the forecast period to 2040. Lower fuel costs, including increased usage of renewable energy, will be offset by investments in new generation and replacement of aging assets. • Nuclear power generation is expected to be stable with a 0% growth rate over the forecast period, with construction of new nuclear power plants projected to offset retirements of existing units. Some older nuclear units have applied for a second 20-year life extension permit, which would allow operation for a total of 80 years.
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• Electricity generation using renewable sources (which includes cogenerators) is expected to increase by 3.6% per year. Total renewables, including hydroelectric power, are projected to grow from their current 12% share of electric generation to 25% in 2040. • Petroleum consumption will grow by 0.2% annually, led by the transportation sector, where most of it (72%) is used. • The share of petroleum consumption met by net imports is projected to be 41% in 2040, down from 51% in 2015. The U.S. reduction in oil imports is remarkable: as recently as 2005, imports were 60% of liquid fuel use. A steady, slow decline in net imports is expected across the forecast period. • Natural gas consumption will increase by 0.9% per year in all sectors, driven mostly by increased electric generation using natural gas. • Coal consumption will decrease at an average annual rate of 1.4%, as coal used for electric generation is replaced by other energy sources. • Electricity consumption is projected to grow by 0.7% annually, with efficiency gains offset by increased use of electricity-using equipment and an increasing population. • Total energy consumption in the residential sector is projected to decline by 0.1% per year, and commercial energy consumption is projected to grow by 0.5% per year, as increased efficiency and other factors approximately balance population growth. • Energy use by the transportation sector is projected to decline by 0.2% per year, with variations from this average depending heavily on prevailing fuel prices. Transportation energy use is declining, which reflects increased fuel efficiency during the forecast period. • Per capita energy use is projected to decline by 0.2% annually, because increases in efficiency more than offset population growth and new energy-consuming products. • Total energy use per dollar of gross domestic product (energy intensity) will continue to fall at an average rate of about 1.7% per year through 2040. • Total carbon emissions are projected to decline by 0.2% annually through 2040, based on current regulations in place, including the Clean Power Plan to restrict electric power sector carbon emissions.
Outlook Summary In general, the following key issues will dominate energy matters in the next two decades: • Reduced U.S. dependency on imported oil. • Increased use of natural gas and renewables, with particularly strong growth in the use of renewables. • Role of technology developments, including energy conservation and energy efficiency as alternatives to energy production. • Substantial increases in use of renewable energy, rising from 7.1% of total U.S. production in 2015 to 13.0% in 2040. For electric generation, renewables increase from 12.8% in 2015 to 27.8% by 2040. • Continued growth in total worldwide carbon emissions, and debate over actions to deal with the issue. • Relative merits of various energy alternatives, including nuclear power and different renewable energy options. • Population growth, coupled with the shift of large population segments into retirement.
3.4
U.S. AGENCIES AND ASSOCIATIONS
American Gas Association (AGA), Washington, D.C. American Petroleum Institute (API), Washington, D.C. Bureau of Mines, Department of Interior, Washington, D.C. Council on Environmental Quality (CEQ), Washington, D.C. Edison Electric Institute (EEI), Washington, D.C.
Electric Power Research Institute (EPRI), Palo Alto, CA Energy Information Administration (EIA), Washington, D.C. Gas Research Institute (GRI), Des Plaines, IL National Coal Association (NCA), Washington, D.C. North American Electric Reliability Council (NAERC), Princeton, NJ Organization of Petroleum Exporting Countries (OPEC), Vienna, Austria United States Green Building Council (USGBC), Washington, D.C.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ASHRAE. 2003. ASHRAE energy position document. ASHRAE. 2013. ASHRAE greenguide: Design, construction, and operation of sustainable buildings, 4th ed. BP. 2015. BP statistical review of world energy June 2015. www.bp.com /content/dam/bp/en/corporate/pdf/bp-statistical-review-of-world-energy -2015-full-report.pdf. BP. 2016. BP statistical review of world energy June 2016. www.bp.com /content/dam/bp/en/corporate/pdf/bp-statistical-review-of-world-energy -2016-full-report.pdf. CIA. 2016. The world factbook. U.S. Central Intelligence Agency, Washington, D.C. www.cia.gov/library/publications/resources/the-world-factbook /index.html. EIA. 2012. Annual energy outlook 2012 with projections to 2035. U.S. Energy Information Administration, Washington, D.C. www.eia.gov /outlooks/aeo/pdf/0383(2012).pdf. EIA. 2015. Annual energy outlook 2015 with projections to 2040. U.S. Energy Information Administration, Washington, D.C. www.eia.gov /outlooks/aeo/pdf/0383(2015).pdf. EIA. 2016. International energy outlook 2016. U.S. Energy Information Administration, Washington, D.C. www.eia.gov/outlooks/ieo/. EPA. 2016. Fact sheet: Clean Power Plan overview: Cutting carbon pollution from power plants. U.S. Environmental Protection Agency, Washington, D.C. www.epa.gov/cleanpowerplan/fact-sheet-clean-power -plan-overview. Grumman, D.L. 1984. Energy resource accounting: ASHRAE Standard 90C1977R. ASHRAE Transactions 90(1B):531-546. Heiman, J.L. 1984. Proposal for a simple method for determining resource impact factors. ASHRAE Transactions 90(1B):564-570. UN. 1987. Our common future: Report of the World Commission on Environment and Development. Annex to General Assembly document A/42/ 427, Development and International Co-operation: Environment. United Nations. www.un-documents.net/wced-ocf.htm.
BIBLIOGRAPHY DOE. 1979. Impact assessment of a mandatory source-energy approach to energy conservation in new construction. U.S. Department of Energy, Washington, D.C. EIA. 2001. Annual energy review 2000. DOE/EIA-0384(2000). Energy Information Administration, U.S. Department of Energy, Washington, D.C. EIA. 2006. Annual energy review 2005. DOE/EIA-0384(2005). Energy Information Administration, U.S. Department of Energy, Washington, D.C. www.eia.gov/totalenergy/data/annual/archive/038405.pdf. EIA. 2011. International energy statistics. U.S. Energy Information Administration, Washington, D.C. EISA. 2007. Energy independence and security act of 2007. HR-6. 110th Congress, 1st session. www.gpo.gov/fdsys/pkg/BILLS-110hr6enr/pdf /BILLS-110hr6enr.pdf. Pacific Northwest Laboratory. 1987. Development of whole-building energy design targets for commercial buildings phase 1 planning. PNL-5854, vol. 2. U.S. Department of Energy, Washington, D.C. Simmons, M.R. 2006. Twilight in the desert: The coming world oil shock and the world economy. John Wiley & Sons. USGBC. 2013. LEED™ reference guides, v.4. U.S. Green Building Council, San Francisco.
Related Commercial Resources
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Related Commercial Resources CHAPTER 35
SUSTAINABILITY Definition....................................................................................................................................... Characteristics of Sustainability ................................................................................................... Factors Impacting Sustainability .................................................................................................. Primary HVAC&R Considerations in Sustainable Design ........................................................... Factors Driving Sustainability into Design Practice.................................................................... Designing for Effective Energy Resource Use ..............................................................................
USTAINABILITY is today a goal that just about every organization, institution, business, or individual claims to be striving for, and sometimes claims to have achieved. Given the profound impact of buildings on the environment, the work of HVAC&R design engineers is inextricably linked to sustainability. The engineering sector has seminal influence on building performance, and HVAC&R designers’ work is inherently related to overall sustainability in buildings. HVAC&R engineering design on projects concerned with performance and sustainability requires understanding of and involvement with more than just HVAC, including projected energy and water demands, stormwater runoff generation, waste generation, and air quality impacts. This chapter is intended to provide key information and identify reference sources for further resources on
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S
• Defining the energy, water, and other resource-consuming aspects of projects • Quantifying the relative environmental impacts of competing design alternatives These aspects of sustainability are addressed with respect to energy and water conservation, greenhouse gas and air quality impacts, and other impacts of buildings, such as stormwater runoff and potable water use. The need to address sustainability in the built environment is being accelerated by external concerns such as environmental and resource issues, rising energy prices, indoor environmental quality, climate change, international pressure, natural disasters, and energy security. While economies transition from carbon-based to other forms of more sustainable energy, engineers will be challenged to meet an ever-increasing tide of regulation, demand, and expectations.
1.
DEFINITION
Sustainability is defined in the ASHRAE GreenGuide (ASHRAE 2013), in general terms, as “providing for the needs of the present without detracting from the ability to fulfill the needs of the future,” a definition very similar to that developed in 1987 by the United Nations’ Brundtland Commission (UN 1987). Others have defined sustainability as “the concept of maximizing the effectiveness of resource use while minimizing the impact of that use on the environment” (ASHRAE 2006) and an environment in which “… an equilibrium … exists between human society and stable ecosystems” (Townsend 2006). Sustaining (i.e., keeping up or prolonging) those elements on which humankind’s existence and that of the planet depend, such as energy, the environment, and health, are worthy goals. The preparation of this chapter is assigned to TC 2.8, Building Environmental Impacts and Sustainability.
35.1 Copyright © 2017, ASHRAE
2.
35.1 35.1 35.2 35.2 35.5 35.8
CHARACTERISTICS OF SUSTAINABILITY
Sustainability Addresses the Future Sustainability is focused on the distant future. Any actions taken under the name of sustainability must address the impact of present actions on conditions likely to prevail in that future time frame. In designing the built environment, the emphasis has often been on the present or the near future, usually in the form of capital (or first-cost) impact. As is apparent when life-cycle costing analysis is applied, capital cost assumes less importance the longer the future period under consideration. This emphasis on the distant future can differentiate sustainable design from green design. Whereas green design addresses many of the same characteristics as sustainable design, it may also emphasize near-term impacts such as indoor environmental quality, operation and maintenance features, and meeting current client needs. Thus, green design may focus more on the immediate future (i.e., starting when the building is first constructed and then occupied). Sustainable design is of paramount importance to the global environment in the long term while still incorporating features of green design that focus on the present and near future.
Sustainability Has Many Contributors Sustainability is not just about energy, carbon emissions, pollution, waste disposal, or population growth. Although these are central ideas in thinking about sustainability, it is an oversimplification to think that addressing one factor, or even any one set of factors, can result in a sustainable future for the planet. It is likewise a mistake to think that HVAC&R design practitioners, by themselves and just through activities within their purview, can create a sustainable result. To be sure, their activities can contribute to sustainability by creating a sustainable building, development, or other related project. But they cannot by themselves create global sustainability. Such an endeavor depends on many outside factors that cannot be controlled by HVAC&R engineers; however, they should make their fair-share contribution to sustainability in all their endeavors, and encourage other individuals and entities to do the same.
Sustainability Is Comprehensive Sustainability has no borders or limits. A good-faith effort to make a project sustainable does not mean that sustainability will be achieved globally. A superb design job on a building with sustainability as a goal will probably not contribute much to the global situation if a significant number of other buildings are not so designed, or if the transportation sector makes an inadequate contribution, or if only a few regions of the world do their fair share toward making the planet sustainable. A truly sustainable outcome thus depends on comprehensive efforts in all sectors the world around.
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35.2
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Technology Plays Only a Partial Role It may well be that in due time technology will have the theoretical capability, if diligently applied, to create a sustainable future for the planet and humankind. Having the capability to apply technology, however, does not guarantee that it will be applied; that must come from attitude or mindset. As with all things related to comprehensive change, there must be the will. For example, automobile companies have long had the technical capability to make cars much more efficient; some developed countries highly dependent on imported oil have brought their transportation sectors close to self sufficiency. Until recently, that has not been the case in the United States. Part of the change is because of increased customer demand, but more of it is driven by government regulation (efficiency standards). The technology is available, but the will is not there; large-scale motivation is absent, what exists being mostly driven by regulation and the motivated few. Similarly, HVAC&R designers know how to design buildings that are much more energy efficient than they have been in the past, but such buildings are still relatively rare, especially in the general commercial market (as opposed to those owned by high-profile entities). ASHRAE’s long-standing guidance in designing energyefficient (now green and/or sustainable) buildings, and the motivation provided by its own and other entities’ programs, have pointed the way technologically for the built environment and related industries to make their fair-share contribution to sustainability. Such programs include (1) ASHRAE’s net-zero energy buildings (NZEB) thrust; (2) the U.S. Green Building Council’s (USGBC) Leadership in Energy and Environmental Design (LEED®) Green Building Rating System™; (3) the American Institute of Architects’ (AIA) 2030 Challenge (AIA 2011); (4) the Green Building Institute’s (GBI) Green Globes (www.thegbi.org/greenglobes); and (5) the U.S. Environmental Protection Agency’s (EPA) ENERGY STAR® program (www.energystar.gov/). The European Union (E.U.) has also taken a lead in the fight against climate change and promoting a low-carbon economy, although unsustainable trends persist in many areas. ASHRAE’s mission “to advance the arts and sciences of HVAC&R to serve humanity and promote a sustainable world” and its vision to “be the global leader, [and] the foremost source of technical and educational information” (www.ashrae.org/about-ashrae) has already had significant impact. However, more can be done, both within its technological purview and in overcoming other, nontechnological barriers. ASHRAE has set a good example in its area of expertise and can also encourage, advise, and inspire other sectors to do their part to move towards sustainability. Examples include ASHRAE’s guidance provided to the U.S. government on effective building energy efficiency programs, as well as its many publications such as the Advanced Energy Design Guides (AEDGs), the ASHRAE GreenGuide, and its numerous standards and guidelines.
3.
FACTORS IMPACTING SUSTAINABILITY
The major factors impacting global sustainability are the following: • • • • • • •
Population growth and migration Food supply Disease control and amelioration Energy resource availability Material resource availability and management Fresh water supply, both potable and nonpotable Effective and efficient usage practices for energy resources and water • Air and water pollution • Solid and liquid waste disposal • Land use
The preceding are only broad categories, yet they encompass many subsidiary factors that have received public attention recently. For instance, climate change/global warming, carbon emissions, acid rain, deforestation, transportation, and watershed management are important factors as well. However, each of these can be viewed as a subset of one or more of the listed major areas.
4.
PRIMARY HVAC&R CONSIDERATIONS IN SUSTAINABLE DESIGN
The main areas falling within an HVAC&R designer’s (and ASHRAE’s) purview on most projects are those dealing with energy and water use, material resources, air and water pollution, and solid waste disposal. Although HVAC&R professionals’ expertise may impact issues such as land use and food supply on certain specialized projects, these more typically fall under the purview of other professionals and their organizations.
Energy Resource Availability Although conventional energy resources and their availability largely fall beyond the scope of HVAC&R designers’ work, an understanding of these topics is often required for participation in project discussions or utility programs relating to projects. Chapter 34 has more information on energy resources. Some renewable energy resources, in contrast with traditional energy and fuels, are ubiquitous by nature and are thus available on many building sites. Wind and solar energy are widely distributed (if not always continuously available) on almost any site for use in active or passive ways. Chapter 35 of the 2015 ASHRAE Handbook—HVAC Applications and Chapter 37 of the 2016 ASHRAE Handbook—Systems and Equipment provide information on solar energy resources, passive and active space heating and cooling, domestic hot water, and applications of photovoltaics. High-level (high-temperature) geothermal energy is only present at limited sites, and may thus be unavailable as a direct energy source on most projects. Low-level geothermal, on the other hand, depends on the nearly constant temperature of the near-surface earth for use as an energy source or sink, and thus can be used on almost any project if other factors align in its favor. See Chapter 34 of the 2015 ASHRAE Handbook—HVAC Applications for more information. Climatic conditions may often provide another source of “renewable” energy. In arid climates, air systems using evaporative cooling (both direct and indirect) can supplement conventionally powered cooling and refrigeration systems. Designers should be familiar with the characteristics of common traditional (nonrenewable) energy resources (natural gas, heating oil, electricity) from the standpoint of their use in relevant building applications. Designers are typically very familiar with the relative per-unit cost as it affects the operating cost of the building being designed. Other energy characteristics traditionally taken into account by the designer might also include ease of handling and use, cleanliness, emissions produced, and local availability, because these also have a direct effect on design and installation. Until recently, designers had little reason to consider an energy resource’s characteristics beyond the site line of the project at hand. However, recent public focus on the impacts of building energy use on the environment has changed that approach. Designers now must consider a resource’s broader characteristics that may affect the regional, national, and global environment, such as its origin (domestic or foreign), security, future availability, emissions characteristics, broad economics, generation/use limitations [gravimetric energy density (J/kg) and areal power density (W/m2)] and social acceptability. Though responsible designers may not be able to do much about such factors, they should be aware of them; indeed, that awareness may affect decisions within the designer’s control.
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Sustainability For instance, familiarity with an energy resource’s emissions characteristics, whether at the well head, mine mouth, or generating station, may influence the designer to make the building more energy efficient, or provide the designer with arguments to convince the owner that energy-saving features in the building would be worth additional capital cost. Furthermore, as owners and developers of buildings become more aware of sustainability factors, designers must stay informed of the latest information and impacts. One way to reduce a project’s use of nonrenewable energy, beyond energy-efficient design itself, is to replace such energy use with renewable energy. Designers should develop familiarity with how projects might incorporate and benefit from renewable energy. Many kinds of passive design features can take advantage of naturally occurring energy. Increasingly common examples of nonpassive approaches are solar systems, whether photovoltaic (electricity-generating) or solar thermal (hot-fluid generating). Low-level geothermal systems take advantage of naturally occurring and widely distributed earth-embedded energy. Wind systems are increasingly applied to supplement electric power grids, and are also sometimes incorporated on a smaller scale into on-site or distributed generation approaches. Some large power users, such as municipalities or large industries, require that a minimum percentage of power they purchase be from renewable sources. Also, renewable portfolio standards are being imposed on electric utility companies by regulators.
Fresh Water Supply HVAC&R systems can impact potable and nonpotable water supplies both directly and indirectly. First, some building systems (e.g., evaporative cooling towers) use potable water. Second, some building systems can discharge treated water or other waste streams with contaminants of concern that can impact local watersheds and water supplies. Indirect impacts include water consumption for electricity generation and in mineral and fuel extraction.
Effective and Efficient Use of Energy Resources and Water This area is where HVAC&R engineers can have a profound impact on achieving sustainability goals. Impacts of building consumption can be at least partially mitigated through overall system performance improvement, as well as through increased use of on-site renewable energy and certain off-site energy resources. See the section on Designing for Effective Energy Resource Use for more information on addressing energy efficiency in the design process. Water Consumption. Building systems’ water use can be reduced by reusing clean water from on site, such as condensate drain water, or by using less potable water. For example, hybrid cooling towers can operate as water-to-air heat exchangers when run dry, and can operate their water sprays for additional evaporative capacity only when conditions require. (See also the section on Energy Resource Availability.) In process control and refrigeration systems, similar opportunities exist. The U.S. EPA’s WaterSense program (www3.epa.gov/water sense/) rates products on their water use efficiency; similarly to the EPA’s ENERGY STAR program, products are certified by an outside third party before they can claim the WaterSense label. ASHRAE is also developing a standard to provide minimum requirements for the design of mechanical systems that limit the volume of water required to operate HVAC systems (Proposed Standard 191P). In Europe, the U.K. Building Regulations (U.K. 2015) requires that design water consumption be reduced in new homes, with a combined hot- and cold-water consumption of no more than 125 L per person per day of potable water. Alternative sources of lowergrade water, such as harvested rainwater and reclaimed gray water,
35.3 may also be used for functions such as toilet flushing, subject to specific measures. The 2015 edition introduced an optional requirement of 110 L per person per day where required by planning permission, and an alternative fittings-based approach to demonstrating compliance instead of the prescribed calculation method. Discharge from building systems can be reduced through careful design, proper sequences and control, and choosing lower-impact chemical or nonchemical water treatment. These techniques may not eliminate chemical treatment in all applications, but negative effects from such usage can be substantially reduced. Water/Energy Nexus. The water/energy nexus refers to the interdependent and inseparable nature of these two important resources. From large-scale utilities to the built environment, water production requires energy to extract and deliver for consumption, and electricity generation and energy sources (e.g., thermal and nuclear power generation, hydraulic fracturing, biofuels) demand significant amounts of water for production. With approximately 8% of the global energy generation used for pumping, treatment, and transportation of water resources and approximately 15% of the world’s total water withdrawal used for energy production, each resource will continue to face rising demands and constraints as a consequence of economic and population growth and climate change. Increasing energy demands, as well as naturally occurring water constraints such as droughts, heat waves, or human-induced shortages, mean that demands on water resources can be expected to increase. In addition, changing temperatures, shifting precipitation patterns, increasing variability, and more extreme weather add significant uncertainty about water availability. Water and energy, in their various classifications, are generally viewed in individual silos, which has limited adoption of integrated solutions. To properly address the challenges and opportunities around the water/ energy nexus, emphasis on policy incentives and sustainable engineering solutions promoting optimized, efficient use of each resource, as well as advancement in technologies promoting both water and energy conservation, are needed.
Material Resource Availability and Management Environmentally conscious design and construction practices are increasingly motivating design teams to apply life-cycle thinking and look in to the embodied impacts (e.g., embodied energy, embodied CO2, other equivalent environmental impact indicators) of their systems’ design, although this is not yet a common practice. For example, within the LEED framework, building systems under the purview of HVAC&R designers are currently excluded from credits for locally procured building materials and resources. However, the same concepts can be applied in selection and procurement of HVAC&R system components. For example, recycled steel content in system components could be required to be stated in HVAC&R product submittals. In some areas, locally assembled or manufactured components may be available that can reduce transportation impacts.
Embodied Energy As buildings become more energy efficient and their operational energy is reduced, more emphasis will shift toward reducing their embodied energy (UNEP 2014). Materials used directly in the construction and the components, equipment, and systems for building operation embody the energy used during their manufacturing, transportation, and installation (ASHRAE 2013). Material selection should also consider the environmental impact of demolition and disposal after the service life of the products. Building life-cycle assessment (LCA) focuses on the environmental impact of a product system (from materials acquisition to manufacturing, use, and final disposition) and plays a major role in promoting sustainability. This is a cradle-to-grave approach that evaluates all stages of individual materials and the product’s life to determine their cumulative
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35.4 environmental impact. Srinivasan et al. (2014) reviews various building assessment methods that support environmental decision making. Designers should give preference to resource-efficient materials and reduce waste by recycling and reusing whenever possible. LCA is an internationally standardized methodology (ISO Standard 14040). Consecutive parts of an LCA include a life-cycle inventory (e.g., collection and analysis of air and water emissions, waste generation, and resource consumption over the product’s life cycle) and life-cycle impact assessment (LCIA), which is an estimation of indicators of the environmental pressures in terms of climate change, summer smog, resource depletion, acidification, human health effects, etc., attributable to the product’s life cycle. More information on the LCA approach is available online from the EPA (www.epa .gov/saferchoice/design-environment-life-cycle-assessments) and the European Commission (ec.europa.eu/environment/ipp/lca.htm). Various LCA databases and tools available from commercial as well as governmental or public domain sources can be used to calculate and compare the embodied energy (e.g., total energy per unit mass), related embodied emissions (e.g., total mass of CO2 per unit mass of material or product), or other embodied environmental impacts of common building materials and products. A challenge for modeling life-cycle energy use in buildings is using consistent system boundaries and data collection (Srinivasan et al. 2014). Available software to address these concerns includes Building for Environmental and Economic Sustainability (BEES; www.nist.gov/services-resources/software/bees) and the Tool for Reduction and Assessment of Chemicals and Other Environmental Impacts (TRACI), which focus on chemical releases and raw materials usage in products. The Athena Sustainable Materials Institute’s decision support tool provides a cradle-to-grave, process-based LCA, including regional data such as energy mix for power generation, transportation modes, etc. (www.athenasmi.org/what-we-do /lca-data-software/). Regional data are important because conversion factors to primary energy and GHG emissions can differ by country, depending on energy sources used (e.g., coal- or oil-fired power plants versus solar or natural-gas-based generation). LCA-based information may be in the form of environmental product declarations (EPD) based on CEN Standard 15804, or product environmental footprints (PEFs) inspired by ISO Standards 14040 and 14044 and voluntary environmental declarations (ISO Standard 14025). They have been gaining in popularity beyond building construction materials, given the growing criteria of green building certifications such as LEED v4 credit, which now rewards selection of HVAC products with EPDs, based on the updated LEED credit interpretation in early 2015. Another example is the French EPD program and national decree that regulates EPDs for construction products, which will also address HVAC equipment and other technical installations by mid-2017 (Passer et al. 2015). LCA can also assess specific refurbishments intended to improve energy performance of systems in existing buildings. For example, replacing electric or gas water heaters with solar hot-water systems can provide net emissions savings compared with the conventional systems after 0.6 month to 2.5 years, depending on the auxiliary fuel (Crawford et al. 2003). For solar domestic hot-water systems and solar central space heating, the energy consumed by producing and installing the solar systems is recovered in about 1.2 years, and the payback time for the systems’ embodied energy emissions varies from a few months (for solar domestic water heating) to 9.5 years (for solar central space heating), again depending on the energy carrier for the conventional system and the specific environmental emission indicators considered (Kalogirou 2004).
2017 ASHRAE Handbook—Fundamentals (SI) Air, Noise, and Water Pollution HVAC&R systems and equipment can interact with both local and global environments. On a local scale, HVAC&R systems interact with the environment in ways such as acoustical noise generated by heat rejecting equipment (e.g., condensing units, cooling tower). Occasionally, this may require the addition of special barriers to prevent sound migration from the site, as shown in Figure 1. Local impacts of combustion from on-site heat or electricity generation can be mitigated to an extent through careful consideration of the location of sources (emitters) with respect to nearby receptors, including outdoor air intakes and residences or other buildings with operable windows. On a larger scale, air and water pollution occurs indirectly through the consumption of energy to operate building systems. This occurs in generating the electricity (whether from fossil fuel, nuclear, or hydroelectric resources), steam, or hot water for building heating or cooling. In this sense, improved efficiency is an approach to partial mitigation.
Solid and Liquid Waste Disposal The solid waste disposal burden from installation and operations of building systems can be substantially reduced. Competing alternatives can be assessed through life-cycle analysis. For example, an air-cooled unitary system with a shorter service life than a costlier water-cooled alternative could, over the course of the building’s life, increase the solid waste burden when it is discarded. Reuse options should also be considered for locally available materials or process by-products. An example of an HVAC&R design impacting liquid waste disposal is using glycol to protect coils from freezing, where the glycol must be eliminated in summer to provide required capacity. Because reusing glycol is not a common practice, such a design would likely result in an annual glycol discharge. In many locations, water quality regulations and agencies essentially limit or prohibit liquid waste disposal. Other approaches to pursue in reducing liquid waste disposal are discussed in the section on Effective and Efficient Use of Energy Resources and Water.
Fig. 1 Cooling Tower Noise Barrier (Courtesy Neil Moiseev)
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Sustainability 5.
FACTORS DRIVING SUSTAINABILITY INTO DESIGN PRACTICE
HVAC&R designers face many challenges as they assimilate sustainability into their engineering practices. These challenges include climate change, a fast-changing regulatory and legal environment, and evolving standards of care. New tools, technologies, and approaches are required for well-prepared HVAC&R engineers. The challenges and the responses are creating new opportunities, just as changing project processes are allowing or requiring engineers to participate in projects in new ways.
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Climate Change In addition to their causal role, energy systems are exposed to significant vulnerabilities resulting from climate change (Bruckner et al. 2014). Increased volatility in weather profoundly affects HVAC&R practice. Historical weather data and extremes may inadequately describe conditions faced by a project built today, even over a modest building lifespan. In 1988, the United Nations Environment Programme (UNEP) and the World Meteorological Organization (WMO) established the Intergovernmental Panel on Climate Change (IPCC) (www.ipcc.ch) to study and report on the scientific issues, potential impacts, and mitigation methods associated with climate change. A series of publications discusses the possible outcomes and interventions required to mitigate the impacts of anthropogenic emissions. During the United Nations Framework Convention on Climate Change in Paris, 195 nations reached a historic agreement to combat climate change and encourage actions and investment toward a lowcarbon, resilient, and sustainable future (UNFCC 2015); the agreement governs greenhouse gas emissions measures from 2020 and limits average global warming to 2 K above preindustrial temperatures, while striving for a limit of 1.5 K. It also aims to strengthen the ability to deal with the impacts of climate change. Crucial areas include • Mitigation: reducing emissions quickly enough to achieve the temperature goal • Transparency system and global stock-taking: accounting for climate action • Adaptation: strengthening countries’ ability to deal with climate impacts • Loss and damage: strengthening ability to recover from climate impacts • Support (including financial support): for nations to build clean, resilient futures (e.g., work to define a clear roadmap on ratcheting up climate finance to USD 100 billion by 2020 for developing nations) The agreement will come into force after 55 countries that account for at least 55% of global emissions have deposited their instruments of ratification. Accordingly, each country should set up a bottom-up system, setting its own goals for nationally determined contribution and a coherent plan for reaching these objectives. Starting in 2018, each country will have to increase their pledges over time and submit new plans every five years. (See also the discussion in the section on Regulatory Environment.) Responsible designers are concerned with multiple dimensions of climate change: not only what they can do to reduce their designs’ contribution, but also whether and how their designs should anticipate the future. It is the first that is the focus of this chapter and a majority of the available information on sustainable design. Warming trends currently occurring have been observed with certainty. As a result, historical weather data may not be the best source for load calculations. Depending on the rate of change, anticipating future weather may become more significant in its impact on the climate control of building systems.
35.5 Responsible designers are concerned with two dimensions of climate change: not only what they can do to reduce their designs’ contribution, but also whether and how their designs should anticipate the future. It is the first that is the focus of this chapter and a majority of the available information on sustainable design. Warming trends currently occurring have been observed with certainty. As a result, historical weather data may not be the best source for load calculations. Depending on the rate of change, anticipating future weather may become more significant in its impact on the climate control of building systems.
Regulatory Environment The global community has responded to two major environmental issues during the past two decades. In the late 1980s, the Montreal Protocol (UNEP 2003) regulated the manufacture and trade of refrigerants that had been shown to damage the stratosphere by depleting stratospheric ozone. The effect on the HVAC&R industry was to require research and investment in alternative materials to those that had become the mainstays of the industry, as shown in Figure 2. Next, in the early 1990s, came the much more controversial issue of greenhouse gas (GHG) emissions and their potential for causing global warming. In response to these threats, some countries signed and accepted the Kyoto Protocol (UNFCCC 1998), which placed future limits on these emissions, but most large-emitter countries did not. By 2011, when follow-up climate talks occurred in Durban, South Africa, overall global GHG emissions not only had not been reduced but had increased. No new GHG emission reduction targets came out of those talks, although the countries agreed to look at the limits issue again in 2020 and to set up a “green fund” to help poor nations deal with climate change. Despite the lack of effective global action, evidence of climate change is compelling. The Synthesis Report that integrated the findings of the three working group contributions to the Fifth Assessment Report of the Intergovernmental Panel on Climate Change (IPCC 2014) confirmed that “human influence on the climate system is clear and growing, with impacts observed across all continents and oceans. . . . Many of the observed changes since the 1950s are unprecedented over decades to millennia. The IPCC is now 95% certain that humans are the main cause of current global warming.” Similar conclusions have also been supported by the National Academy of Sciences (NAS 2010), which concluded that “Climate change is occurring, is caused largely by human activities, and poses significant risk for—and in many cases is already affecting—a broad range of human and natural systems.” The predominant greenhouse gas pollutant is carbon dioxide, which is mainly a by-product of fossil fuel combustion in the
Fig. 2 Effect of Montreal Protocol on Global Chlorofluorocarbon (CFC) Production
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35.6 transportation, power, industrial, residential, and commercial sectors. According to the EPA (2016), CO2 contributed about 82% of total U.S. emissions in 2013. Methane is the next highest contributor, accounting for about 10% of the total U.S. emissions. Sources of methane emissions include oil and gas systems, enteric fermentation, landfills, coal mines, etc. Nitrous oxide (N2O) and fluorocarbon gases are other contributors to GHG emissions. In 2014, the top four emitting countries/regions, accounting for almost two-thirds (61%) of the total global CO2 emissions, are China (30%), the United States (15%), the 28 E.U. Member States (10%), and India (6.5%). On a per capita basis, emissions in the United States (16.5 tonnes CO2 per capita) are twice as high as those of China (7.6) and the European Union (6.7) (Olivier et al. 2015). Various standards, policies, and regulations under way that target reduction of U.S. GHGs in various sectors: examples include state renewable portfolio standards (RPS) programs, corporate average fuel economy (CAFE) standards, state emission performance standards for power plants [e.g., California (2006) Senate Bill SB 1368], and cap-and-trade programs in the Regional Greenhouse Gas Initiative (RGGI; www.rggi.org) states and California. In the United States, a consolidation of green building codes is planned: the merger of ASHRAE Standard 189.1 and ICC’s International Green Construction Code (see also the section on Evolving Standards of Care) is on track to occur in 2018. Energy efficiency levels in U.S. codes will continue to improve, with the green building codes evolving toward a net energy intensity level that, by 2025, is 20 to 25% of the code minimums that existed at the turn of the 21st century. Under its 2020 strategy, the E.U. has committed to an ambitious plan and introduced binding legislation for the E.U. Member States to meet three climate and energy targets by the end of 2020: (1) 20% cut in GHG emissions from 1990 levels, (2) obtain 20% of E.U. energy from renewables, and (3) 20% improvement in energy efficiency (ec.europa.eu/clima/policies/strategies/index_en.htm). The E.U. emissions trading system (E.U. ETS) is a cornerstone of the policy to combat climate change and a key tool for reducing industrial GHG cost effectively. The ETS covers more than 11 000 power stations and industrial plants in 31 countries, as well as airlines, and accounts for about 55% of total E.U. emissions. For sectors not in the ETS, the E.U. Member States have taken on binding annual targets until 2020 for cutting emissions under the effort-sharing decision (between 2013 and 2020) according to national wealth, measured by gross domestic product per capita; these targets range from a 20% cut for the richest countries (e.g., Luxembourg, Denmark) to a maximum 20% increase for the least wealthy (e.g., Bulgaria). Moreover, a new E.U. framework has set three key targets for 2030: (1) at least 40% cuts in GHG, (2) at least 27% share for renewable energy, and (3) at least 27% improvement in energy efficiency. The 2050 roadmap suggests that the E.U. should cut emissions to 80 to 95% below 1990 levels.
Evolving Standards of Care This section is based on Lawrence et al. (2016). Litigation relating to sustainability and global climate issues has increased. For example, a consortium of states successfully sued, and the U.S. Supreme Court agreed in 2007, that the U.S. EPA may act to consider CO2 a pollutant that is harming the environment and thus take measures to regulate its emissions. This ruling is one of several developments in the continued and broadened response to CO2 emissions by society at large. Building design and construction industries are already being impacted. Sustainability is being adopted into building codes at different levels of government and with varying motivation. Different approaches reflect local societal perceptions, political priorities, national policies and economic factors [see, e.g., Lawrence et al.
2017 ASHRAE Handbook—Fundamentals (SI) (2016) and references therein for an overview of current status and trends of sustainable building codes adopted in the United States and the E.U.]. U.S. Initiatives. In the United States, a combination of methods and programs have gradually increased focus on sustainability in new commercial and residential construction. The methods range from voluntary programs (including rating systems and guidelines) to standards and mandatory enforceable codes. One early voluntary program was U.S. Green Building Council’s Leadership in Energy and Environmental Design (USGCB’s LEED), which was at the forefront of including sustainable design practices in building codes worldwide. LEED ratings can be applied to a wide variety of building types and aspects (e.g., interior design and construction, neighborhood development, operations and maintenance); see www .usgbc.org/resources/grid/leed for specific ratings systems. At least 16 U.S. states now require LEED Silver certification for public buildings, and in most states some level of LEED certification is a required option for state buildings. Most states’ building codes follow the International Green Construction Code™ (IgCC™) for sustainability guidance. One of IgCC’s compliance options involves following ASHRAE Standard 189.1’s mandatory criteria in all topical areas (e.g., site, construction, materials, energy, indoor environmental quality, water). In 2010, California adopted its own statewide green buildings code, CALGreen, an updated version of which came into effect January 2014 (CALGreen 2013). Unlike most other U.S. codes, this approach includes criteria for both residential and nonresidential buildings in the same program. ASHRAE’s Building Energy Quotient® (Building EQ®; www .buildingenergyquotient.org/) focuses on lowering building operating cost and increasing value. Both as-designed and in-operation energy use ratings are available. As with green building codes, cities are beginning to require energy reporting and benchmarking for commercial buildings. As of early 2016, 14 major cities (Atlanta, Austin, Berkeley, Boston, Boulder, Cambridge, Chicago, Kansas City, New York, Philadelphia, Portland, San Francisco, Seattle, and Washington, D.C.) had adopted some form of energy reporting in local ordinances, although the size and type of buildings involved may differ. The motivation for these requirements is usually increasing overall citywide energy efficiency, as well as creating local energy auditing jobs. E.U. Initiatives. Outside the United States, the most notable voluntary program is the BREEAM method (www.breeam.com), an environmental assessment method and rating system introduced in the U.K. in 1990. This is a consensus-based, market-oriented assessment and sustainability benchmarking program for any type of building or large-scale community worldwide. It has one mandatory assessment area (the building’s potential environmental impact) and two optional assessment areas (design process, and operation/maintenance). Over 425,000 buildings have been certified. Specific national schemes are available in Germany, Norway, Sweden, Spain, and The Netherlands. Another U.K. rating system is the Global Environmental Method (GEM) (www.greenglobes.com/existing/homeuk.asp), which is a version of the U.S. and Canadian program Green Globes (www .greenglobes.com). In Germany, the passive house concept has evolved to an international association for standardizing the design and construction of low-energy buildings (passiv.de/en/). Similar approaches and labels, especially for residential buildings, are also used in France (e.g., Effinergie) and Switzerland (e.g., Minerigie). European regulatory efforts have also introduced several mandatory directives towards sustainability at all stages of the energy chain, targeting the buildings sector that plays a major role. The European Union’s Sustainable Development Strategy (SDS)
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Sustainability (ec.europa.eu/environment/eussd/index.htm) addresses climate change and clean energy, sustainable consumption and production, conservation and management of natural resources, public health, social inclusion, global poverty and sustainable development challenges, and education and training. The main E.U. sustainable consumption and production (ESCP) initiatives support efforts to meet the goals of SDS (ec.europa.eu/environment/eussd/escp_en .htm) and build on international and E.U.-wide initiatives and tools [e.g., the United Nations’ Marrakech Process (esa.un.org/marrakech process/index.shtml)]. Beyond operational energy issues for sustainable buildings, E.U. ESCP policy focuses on resources such as materials (including waste), water, and embodied energy. The E.U. Waste Framework Directive (WFD; E.U. Directive 2008/98/EC) aims for 70% for reuse, recycling, and others forms of material recovery (excluding energy recovery), and is the main European policy driver toward better recycling of construction and demolition waste. The WFD is expected to reduce burdens on the waste stream, because construction and demolition waste (CDW) accounts for 25 to 30% of the waste generated in Europe. CDW has a high potential for recycling and reuse, averaging about 46% across Europe (depending on value of the materials and availability of well-developed technology and infrastructure) (Dodd et al. 2015). Energy efficiency can be increased at all stages of the energy chain, from generation to consumption (ec.europa.eu/energy/en /topics/energy-efficiency). The main European legislative tool for improving buildings’ energy efficiency and reducing carbon emissions is the Energy Performance of Buildings Directive (EPBD; E.U. Directive 2010/31/EC): E.U. Member States must apply minimum energy performance requirements on all new and existing residential and nonresidential buildings when undergoing major renovation (25% of building surface or value). The EPBD requires that all new buildings must be nearly zero energy as of January 2021; new buildings occupied or owned by public authorities must comply as of January 2019. Other major policies include • Renewable Energy Directive (RED; E.U. Directive 2009/28/EC), which establishes an overall policy for the production and promotion of energy from renewables • Ecodesign (ED; E.U. Directive 2009/125/EC) and Energy Labelling Directives (ELD; E.U. Directive 2010/30/E.U.), which establish minimum energy efficiency standards for various products (including air-conditioners, boilers, circulators, motors, fans, pumps) • Energy Efficiency Directive (EED; E.U. Directive 2012/27/E.U.), which establishes a common framework for promotion of energy efficiency, and sets energy savings requirements for buildings In accordance with the EPBD, energy performance certification (EPC) of European buildings is an ongoing process for several years; see ASHRAE (2013) for examples. EPCs document the building’s energy performance, usually using an easy-to-understand indicator expressed as a ranking energy label (building class) with an index in terms of primary or final energy use, carbon dioxide emissions, or energy cost per unit of conditioned floor area. EPCs also may include an assessment of indoor environmental quality. In January 2013, The Netherlands became the first E.U. Member State to require measurement of GHG in buildings (Dodd et al. 2015): the Dutch Building Decree requires reporting of GHG emissions and depletion of natural resources for structural components of residential and office buildings (over 100 m2) on application for a building permit.
35.7 many resources and approaches cited in this chapter. (See the section on Designing for Effective Energy Resource Use.) ASHRAE, in partnership with the Illuminating Engineering Society (IES) and USGBC, developed Standard 189.1 for highperformance green buildings, which calls for a determination of annual CO2 equivalent emissions in addition to overall energy savings and other requirements. The component of such emissions from electricity use depends on the mix of fuels used to generate that electricity. In addition to regional variations, the overall fuel mix is projected to change, as shown in Figure 3. Emissions considerations alone are not the only driver for design decision making. Energy prices and societal pressures continue to mount. Examples of recent drivers include • Antiquated electric transmission and distribution infrastructure and plans to develop a smart grid to improve it • Power plants being forced to become cleaner and more efficient, expediting closure of cheap, dirty generators • Mandates imposed on utilities to provide more renewable energy to customers • Influence of commodities trading markets on spot and future prices • Constrained natural gas reserves and growth in demand continuing to increase volatility in the natural gas market • Climate change, through environmental pressures to reduce carbon emissions in the face of increased demand for electricity, and infrastructure damage from more frequent storms • Growing impatience from some elements, both domestically and internationally, over the perceived slow pace of acceptance of sustainable design, leading proponents to push harder for seriously addressing climate and energy resource issues These and other pressures are changing project teams and their work; those teams are being asked to • Incorporate sustainable design guidance, standards, and rating systems into their work • Add a variety of new team members to bring additional expertise to address sustainability • Gather quantitative data related to energy, water, occupant satisfaction, greenhouse gas emissions, etc. • Use new analysis tools (e.g., daylighting modeling) to help maximize sustainability Opportunities relating to sustainability for the well-prepared engineer are growing. The increased focus on sustainability in the built environment allows for more integrated, effective, and efficient ways to meet the nexus between environment, economy, regulation, and societal pressure. The challenge for the industry is how quickly
Changing Design Process Even in jurisdictions without regulatory action, change is happening in the HVAC&R industry. Today’s engineer can contribute value to projects that have sustainability goals, using some of the
Fig. 3 Electricity Generation by Fuel, 1980–2030 (EIA 2008)
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35.8
2017 ASHRAE Handbook—Fundamentals (SI)
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it can adapt to these new opportunities and grow in an increasingly regulated environment. At the very least, the standard of care for engineers must be tracked and implemented to manage liability. Sustainability can provide an opportunity for engineers and others to increase market share while exceeding current regulatory constraints and anticipating future regulations. More details on design considerations are provided in the section on Designing for Effective Energy Resource Use. Integrating sustainability into HVAC&R system design can result in built environments that respect the greater environment and provide safe and comfortable indoor environments. The three occurrences of the letter i in sustainability can be thought of as representing key concepts in sustainable design: interactive, iterative, and integrated. Design processes that require greater interaction between team members and more iterative analysis to improve design solutions can be undertaken by teams through what has become known as integrated design. Sustainability is inherently multidisciplinary. Recognizing this, teams often assemble a broad array of experts in a collaborative, interdisciplinary approach to achieve the highest levels of sustainability possible. This integrated design approach is addressed in Chapter 58 of the 2015 ASHRAE Handbook—HVAC Applications and in ASHRAE (2013).
Other Opportunities In addition to designing HVAC&R systems, engineers may increasingly be called upon to help address issues ranging from transportation to irrigation to on-site renewable energy. The approach to sustainable design alternatives opens the door for creativity and innovation in the design process. Rather than taking a one-size-fits-all approach to design, engineers can provide a range of available solutions and facilitate flexible implementation. Often, engineers are asked to develop and evaluate measures based on both economic and environmental performance. Success may require several design iterations to achieve the desired performance.
6.
DESIGNING FOR EFFECTIVE ENERGY RESOURCE USE
Most energy used in buildings is from nonrenewable resources, the cost of which historically has not considered replenishment or environmental impact. Thus, consideration of energy use in design has been based primarily on economic advantages, which are weighted to encourage more rather than less use. As resources become less readily available and more exotic, and replenishable sources are investigated, the need to operate buildings effectively using less energy becomes paramount. Extensive studies since the mid-1970s [see, e.g., Doris et al. (2009) and references therein] have shown that building energy use can be significantly reduced by applying the fundamental principles discussed in the following sections.
Energy Ethic: Resource Conservation Design Principles The basic approach to energy-efficient design is reducing loads (power), improving transport systems, and providing efficient components and “intelligent” controls. Important design concepts include understanding the relationship between energy and power, maintaining simplicity, using self-imposed budgets, and applying energy-smart design practices.
Energy and Power From an economic standpoint, more energy-efficient systems need not be more expensive than less efficient systems. Quite the opposite is true because of the simple relationship between energy and power, in which power is simply the time rate of energy use (or,
conversely, energy is power times time). Power terms such as kilowatts are used in expressing the size of a motor, chiller, boiler, or transformer. Generally, the smaller the equipment, the less it costs. Other things being equal, as smaller equipment operates over time, it consumes less energy. Thus, in designing for energy efficiency, the first objective is always to reduce the power required to the bare minimum necessary to provide the desired performance, starting with the building’s heating and cooling loads (a power term, in kilowatts) and continuing with the various systems and subsystems.
Simplicity Complex designs to save energy seldom function in the manner intended unless the systems are continually managed and operated by technically skilled individuals. Experience has shown that longterm, energy-efficient performance with a complex system is seldom achievable. Further, when complex systems are operated by minimally skilled individuals, both energy efficiency and performance suffer. Most techniques discussed in this chapter can be implemented with great simplicity.
Self-Imposed Budgets Just as an engineer must work to a cost budget with most designs, self-imposed power budgets can be similarly helpful in achieving energy-efficient design. The Advanced Energy Design Guide series from ASHRAE are a source for guidance on achievable design budgets. For example, the following are possible categories of power (or power-affecting) design budgets for a mid-rise office building: • • • • • • • • • • •
Installed lighting (overall) Space sensible cooling Space heating load Electric power (overall) Thermal power (overall) Hydronic system head Water chiller (water-cooled) Chilled-water system auxiliaries Unitary air-conditioning systems Annual electric energy Annual thermal energy
W/m2 W/m2 W/m2 W/m2 W/m2 kPa kW/kW cooling (COP) kW/kW cooling kW/kW cooling (COP) MJ/(m2 ·yr) kJ/(m2 ·yr·K·day)
As the building and systems are designed, all decisions become interactive as each subsystem’s power or energy performance is continually compared to the budget.
Design Process for Energy-Efficient Projects Consider energy efficiency at the beginning of the building design process, because energy-efficient features are most easily and effectively incorporated at that time. Seek the active participation of all members of the design team, including the owner, architect, engineer, and often the contractor, early in the design process. Consider building attributes such as building function, form, orientation, window/wall ratio, and HVAC system types early in the process, because each has major energy implications. Identify meaningful energy performance benchmarks suited to the project, and set project-specific goals. Energy benchmarks for a sample project are shown in Table 1. Consider energy resources, on-site energy sources, and use of renewable energy, credits, utility rebates, or carbon offsets to mitigate environmental impacts of energy use. Address a building’s energy requirements in the following sequence: 1. Minimize the impact of the building’s functional requirements by analyzing how the building relates to its external environment. Advocate changes in building form, aspect ratio, and other attributes that reduce, redistribute, or delay (shift) loads. The load calculation should be interactive so that the effect of those factors can be seen immediately.
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Sustainability
35.9 Table 1
Example Benchmark and Energy Targets for University Research Laboratory
Building area, m2
Electric
Gross
Lit/ Conditioned
15 793
10 266
Electricity Electricity for Electricity for Electricity Electricity for for Ventilation In-Building for Plug Unidentified Total Cogenerated Grid Lighting (Fans) Pumps Loads Loads Electricity Electricity Electricity
Design load, W/m2
5.60
5.38
6.46
10.4
—
28.0
—
Peak demand, W/m2
4.52
5.38
4.52
7.86
0.0017
22.3
—
71
85
72
124
20
372
—
218 154
346 598
191 245
891 503
175 200
1 823 000
966 000
Peak demand, kW (Projected submetered peak) Annual consumption, kWh/yr (Projected submetered reading) Annual use index goal, kWh/yr
1.28
2.04
1.12
5.24
1.03
10.72
Annual use index goal, site MJ/m2 gross·yr
4378
6956
3838
17 893
3516
36 583
27.0 to 35.7
48.2 to 74.0
included elsewhere
47.3 to 61.0
NA
158.7 to 192.8
97.3
173.6
—
170.1
—
571.1 to 694.0
Annual use index, kWh/m2 gross·yr* Annual use index, site MJ/m2 gross·yr*
857 000
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*From Labs21 program of U.S. Environmental Protection Agency (EPA) and U.S. Department of Energy (DOE). See labs21benchmarking.lbl.gov.
2. Minimize loads by analyzing external and internal loads imposed on the building’s energy-using subsystems, both for peak- and part-load conditions. Design for efficient and effective operation off-peak, where the majority of operating hours and energy use typically occurs. 3. Maximize subsystem efficiency by analyzing the diversified energy and power requirements of each energy-using subsystem serving the building’s functional requirements. Consider static and dynamic efficiencies of energy conversion and energy transport subsystems, and consider opportunities to reclaim, redistribute, and store energy for later use. 4. Study alternative ways to integrate subsystems into the building by considering both power and time components of energy use. Identify, evaluate, and design each of these components to control overall design energy consumption. Consider the following when integrating major building subsystems: • Address more than one problem at a time when developing design solutions, and make maximum use of the building’s advantageous features (e.g., windows, structural mass). • Examine design solutions that consider time (i.e., when energy use occurs), because sufficient energy may already be present from the environment (e.g., solar heat, night cooling) or from internal equipment (e.g., lights, computers) but available at times different from when needed. Active (e.g., heat pumps with water tanks) and passive (e.g., building mass) storage techniques may need to be considered. • Examine design solutions that consider the anticipated use of space. For example, in large but relatively unoccupied spaces, consider task or zone lighting. Consider transporting excess energy (light and heat) from locations of production and availability to locations of need instead of purchasing additional energy. • Never reject waste energy at temperatures usable for space conditioning or other practical purposes without calculating the economic benefit of energy recovery or treatment for reuse. • Consider or advocate design solutions that provide more comfortable surface temperatures or increase the availability of controlled daylight in buildings where human occupancy is a primary function. • Use easily understood design solutions, because they have a greater probability of use by building operators and occupants.
• Where the functional requirements of a building are likely to change over time, design the installed environmental system to adapt to meet anticipated changes and to provide flexibility in meeting future changes in use, occupancy, or other functions. • Develop energy performance benchmarks, metrics, and targets that allow building owners and operators to better achieve the design intent. The effectiveness of these benchmarks was studied in ASHRAE research project 1627, which examines performance of K-12 schools and small office buildings that use the AEDGs (Jones et al. 2016). • Differentiate between peak loads for system design and selection and lower operating loads that determine actual energy use.
Building Energy Use Elements Envelope. Control thermal conductivity by using insulation (including movable insulation), thermal mass, and/or phase-change thermal storage at levels that minimize net heating and cooling loads on a time-integrated (annual) basis. • Minimize unintentional or uncontrolled thermal bridges, and include them in energy-related calculations because they can radically alter building envelope conductivity. Examples include wall studs, balconies, ledges, and extensions of building slabs. • Minimize infiltration so that it approaches zero. (An exception is when infiltration provides the sole means of ventilation, such as in small residential units.) This minimizes fan energy consumption in pressurized buildings during occupied periods and minimizes heat loss (or unwanted heat gain, in warm climates) during unoccupied periods. In warm, humid climates, a tight envelope also improves indoor air quality. Reduce infiltration through design details that enhance the fit and integrity of building envelope joints in ways that may be readily achieved during construction (e.g., caulking, weatherstripping, vestibule doors, revolving doors), with construction meeting accepted specifications. Building envelope commissioning or testing can help verify these design and construction targets. • Consider operable windows to allow occupant-controlled ventilation. This requires careful design of the building’s mechanical system to minimize unnecessary HVAC energy consumption, and building operators and occupants should be cautioned about improper use of operable windows. CIBSE (2005) provides comprehensive design considerations for natural ventilation.
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35.10 • Strive to maintain occupant radiant comfort regardless of whether the building envelope is designed to be a static or dynamic membrane. Design opaque surfaces so that average inside surface temperatures remain within 3 K of room temperature in the coldest anticipated weather (i.e., winter design conditions) and so that the coldest inside surface remains within 14 K of room temperature (but always above the indoor dew point). In a building with timevarying internal heat generation, consider thermal mass for controlling radiant comfort. In the perimeter zone, thermal mass is more effective when it is positioned inside the envelope’s insulation. • Effective control of solar radiation is critical to energy-efficient design because of the high level of internal heat production in most commercial buildings. In some climates, lighting energy consumption savings from daylighting techniques can be greater than the heating and cooling energy penalties that result from additional glazed surface area required, if the building envelope is properly designed for daylighting and lighting controls are installed and used. (In other climates, there may not be net savings.) Daylighting designs are most effective if direct solar beam radiation is not allowed to cause glare in building spaces. • Design transparent parts of the building envelope to prevent solar radiant gain above that necessary for effective daylighting and solar heating. On south-facing facades (in the northern hemisphere), using low shading coefficients is generally not as effective as external physical shading devices in achieving this balance. Consider low-emissivity, high-visible-transmittance glazings for effective control of radiant heat gains and losses. For shading control, judicious use of vegetation may block excess gain year-round or seasonally, depending on the plant species chosen. Lighting. Lighting is both a major energy end use in commercial buildings (especially office buildings) and a major contributor to internal loads by increasing cooling loads and decreasing heating loads. Design should both meet the lighting functional criteria of the space and minimize energy use. IES (2011) recommends illuminance levels for visual tasks and surrounding lighted areas. Principles of energy-conserving design within that context include the following: • Energy use is determined by the lighting load (demand power) and its duration of use (time). Minimize actual demand load rather than just apparent connected load. Control the load rather than just area switching, if switching may adversely affect the quality of the luminous environment. • Consider daylighting with proper controls to reduce costs of electric lighting. Design should be sensitive to window glare, sudden changes in luminances, and general user acceptance of daylighting controls. Carefully select window treatment (blinds, drapes, and shades) and glazing to control direct solar penetration and luminance extremes while maintaining the view and daylight penetration. • Design the lighting system so that illumination required for tasks is primarily limited to the location of the task and comes from a direction that minimizes direct glare and veiling reflections on the task. When the design is based on nonuniform illuminance, walls should be a light to medium color or illuminated to provide visual comfort. In densely occupied work spaces, uniform distribution of general lighting may be most appropriate. Where necessary, provide supplementary task illumination. General ambient illumination should not be lower than a third of the luminance required for the task, to help maintain visually comfortable luminance ratios. • Use local task lighting to accommodate needs for higher lighting levels because of task visual difficulty, glare, intermittently changing requirements, or individual visual differences (poor or aging eyesight).
2017 ASHRAE Handbook—Fundamentals (SI) • Group similar activities so that high illuminance or special lighting for particular tasks can be localized in certain rooms or areas, and so that less-efficient fixtures required for critical glare control do not have to be installed uniformly when they are only required sparsely. • Use lighting controls throughout so lighting is available when and where it is needed, but not wasted when tasks are less critical or spaces are not fully occupied. Also consider user acceptance of control strategies to maximize energy saving. • Only use lower-efficiency incandescent lamps in applications where their characteristics cannot be duplicated by other sources, because manufacturing of most incandescent lamps will be discontinued during the life of the building. • Carry lighting design through the rest of the building’s interior design. Reduced light absorption may be achieved by using lighter finishes, particularly on ceilings, walls, and partitions. Other Loads. • Minimize thermal impact of equipment and appliances on HVAC systems by using hoods, radiation shields, or other confining techniques, and by using controls to turn off equipment when not needed. Where practical, locate major heat-generating equipment where it can balance other heat losses. Computer centers or kitchen areas usually have separate, dedicated HVAC equipment. In addition, consider heat recovery for this equipment. • Use storage techniques to level or distribute loads that vary on a time or spatial basis to allow operation of a device at maximum (often full-load) efficiency. HVAC System Design. • Consider separate HVAC systems to serve areas expected to operate on widely differing operating schedules or design conditions. For instance, systems serving office areas should generally be separate from those serving retail areas. • Arrange systems so that spaces with relatively constant, weatherindependent loads are served by systems separate from those serving perimeter spaces. Areas with special temperature or humidity requirements (e.g., computer rooms) should be served by systems separate from those serving areas that require comfort heating and cooling only. Alternatively, provide these areas with supplementary or auxiliary systems. • Sequence the supply of zone cooling and heating to prevent simultaneous operation of heating and cooling systems for the same space, to the extent possible. Where this is not possible because of ventilation, humidity control, or air circulation requirements, reduce air quantities as much as possible before incorporating reheating, recooling, or mixing hot and cold airstreams. For example, if reheat is needed to dehumidify and prevent overcooling, only ventilation air needs to be treated, not the entire recirculated air quantity. Finally, reset supply air temperature up to the extent possible to reduce reheating, recooling, or mixing losses. • Provide controls to allow operation in occupied and unoccupied modes. In occupied mode, controls may provide for a gradually changing control point as system demands change from cooling to heating. In unoccupied mode, ventilation and exhaust systems should be shut off if possible, and comfort heating and cooling systems should be shut off except to maintain space conditions ready for the next occupancy cycle. • In geographical areas where diurnal temperature swings and humidity levels permit, consider judicious coupling of air distribution and building structural mass to allow nighttime cooling to reduce the requirement for daytime mechanical cooling. • High ventilation rates, where required for special applications, can impose enormous heating and cooling loads on HVAC equipment. In these cases, consider recirculating filtered and cleaned
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Sustainability air to the extent possible, rather than 100% outdoor air. Also, consider preheating outdoor air with reclaimed heat from other sources. HVAC Equipment Selection.
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• To allow HVAC equipment operation at the highest efficiencies, match conversion devices to load increments, and sequence the operation of modules. Oversized or large-scale systems should never serve small seasonal loads (e.g., a large heating boiler serving a summer-service water-heated load). Include specific low-load units and auxiliaries where prolonged use at minimal capacities is expected. • Select the most efficient (or highest-COP) equipment practical at both design and reduced capacity (part-load) operating conditions. • When selecting large-power devices such as chillers (including their auxiliary energy burdens), perform an economic analysis of the complete life-cycle costs. See Chapter 37 of the 2015 ASHRAE Handbook—HVAC Applications for more information on detailed economic analysis. • Keep fluid temperatures for heating equipment devices as low as practical and for cooling equipment as high as practical, while still meeting loads and minimizing flow quantities. Energy Transport Systems. Energy should be transported as efficiently as possible. The following options are listed in order of theoretical efficiency, from the lowest energy transport burden (most efficient) to the highest (least efficient): 1. 2. 3. 4.
Electric wire or fuel pipe Two-phase fluid pipe (steam or refrigerant) Single-phase liquid/fluid pipe (water, glycol, etc.) Air duct
Select a distribution system that complements other parameters such as control strategies, storage capabilities, conversion efficiency, and utilization efficiency. The following specific design techniques may be applied to thermal energy transport systems: Steam Systems. • Include provisions for seasonal or nonuse shutdown. • Minimize venting of steam and ingestion of air, with design directed toward full-vapor performance. • Avoid subcooling, if practical. • Return condensate to boilers or source devices at the highest possible temperature. Hydronic Systems. • Minimize flow quantity by designing for the maximum practical temperature range. • Vary flow quantity with load where possible. • Design for the lowest practical pressure rise (or drop). • Provide operating and idle control modes. • When locating equipment, identify the critical pressure path and size runs for the minimum reasonable pressure drop. Air Systems. • Minimize airflow by careful load analysis and an effective distribution system. If the application allows, supply air quantity should vary with sensible load (i.e., VAV systems). Hold the fan pressure requirement to the lowest practical value and avoid using fan pressure as a source for control power. • Provide normal and idle control modes for fan and psychrometric systems. • Keep duct runs as short as possible, and keep runs on the critical pressure path sized for minimum practical pressure drop.
35.11 Power Distribution. • Size transformers and generating units as closely as possible to the actual anticipated load (i.e., avoid oversizing to minimize fixed thermal losses). • Consider distribution of electric power at the highest practical voltage and load selection at the maximum power factor consistent with safety. • Consider tenant submetering in commercial and multifamily buildings as a cost-effective energy conservation measure. (A large portion of energy use in tenant facilities occurs simply because there is no economic incentive to conserve.) Domestic Hot-Water Systems. • Choose shower heads that provide and maintain user comfort and energy savings. They should not have removable flow-restricting inserts to meet flow limitation requirements. • Consider point-of-use water heaters where their use will reduce energy consumption and annual energy cost. • Consider using storage to facilitate heat recovery when the heat to be recovered is out of phase with the demand for hot water or when energy use for water heating can be shifted to take advantage of off-peak rates. Controls. Well-designed digital control provides information to managers and operators as well as to the data processor that serves as the intelligent controller. Include the energy-saving concepts discussed previously throughout the operating sequences and control logic. However, energy conservation should not be sought at the expense of adequate performance; in a well-designed system, these two parameters are compatible. See Chapter 7 of this volume and Chapter 47 of the 2015 ASHRAE Handbook—HVAC Applications for more information on controls.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. AIA. 2011. 2030 challenges. architecture2030.org/category/2030-challenges/. ASHRAE. 2006. ASHRAE’s sustainability roadmap—The approach to defining a leadership position in sustainability. Presidential Ad Hoc Committee. ASHRAE. 2013. ASHRAE greenguide: The design, construction, and operation of sustainable buildings, 4rd ed. A. Darwich, J. Means and T. Lawrence, eds. Athena. 2017. LCA data and software. Athena Sustainable Materials Institute, Ottawa, ON. www.athenasmi.org. Bruckner, T., I.A. Bashmakov, Y. Mulugetta, H. Chum, A. de la Vega Navarro, J. Edmonds, A. Faaij, B. Fungtammasan, A. Garg, E. Hertwich, D. Honnery, D. Infield, M. Kainuma, S. Khennas, S. Kim, H.B. Nimir, K. Riahi, N. Strachan, R. Wiser, and X. Zhang. 2014. Energy systems. Ch. 7 in Climate change 2014: Mitigation of climate change. Contribution of Working Group III to the Fifth Assessment Report of the Intergovernmental Panel on Climate Change, O. Edenhofer, R. Pichs-Madruga, Y. Sokona, E. Farahani, S. Kadner, K. Seyboth, A. Adler, I. Baum, S. Brunner, P. Eickemeier, B. Kriemann, J. Savolainen, S. Schlömer, C. von Stechow, T. Zwickel, and J.C. Minx, eds. Cambridge University Press, U.K. www.ipcc.ch/pdf/assessment-report/ar5/wg3/ipcc_wg3_ar5_chapter7 .pdf. CALGreen. 2013. California Green Building Standards Code, California Code of Regulations, Title 24, Part 11. California. 2006. Emission performance standards. Senate Bill SB 1368. California Energy Commission, Sacramento. www.energy.ca.gov/emission _standards/.
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CEN. 2013. Sustainability of construction works—Environmental product declarations—Core rules for the product category of construction products. European Standard EN 15804+A1. European Committee for Standardization, Brussels. CIBSE. 2005. Natural ventilation in non-domestic buildings. Applications Manual 10. Chartered Institution of Building Services Engineers, London. Crawford, R.H., G.J. Treloar, B.D. Ilozor and P.E.D. Love. 2003. Comparative greenhouse emissions analysis of domestic solar hot water systems. Building Research & Information 31(1):34-47. Dodd, N., S. Donatello, E. Garbarino and M. Gama-Caldas. 2015. Identifying macro-objectives for the life cycle environmental performance and resource efficiency of EU buildings. Joint Research Centre, European Commission. susproc.jrc.ec.europa.eu/Efficient_Buildings/docs/150528 %20Resource%20Efficient%20Buildings_Macro%20objectives%20WP _Draft%20final.pdf. Doris, E., J. Cochran, and M. Vorum. 2009. Energy efficiency policy of the United States: Overview of trends at different levels of government. Technical Report NREL/TP-6A2-46532. National Renewable Energy Laboratory, Golden, CO. www.nrel.gov/docs/fy10osti/46532.pdf. EIA. 2015. Annual energy outlook 2015. DOE/EIA-0383(2007). Energy Information Administration, U.S. Department of Energy, Washington, D.C. EPA. 2016. Tool for reduction and assessment of chemicals and other environmental impacts (TRACI), v. 2.1. U.S. Environmental Protection Agency, Washington, D.C. www.epa.gov/chemical-research/tool-reduction -and-assessment-chemicals-and-other-environmental-impacts-traci. E.U. 2008. Directive 2008/98/EC of the European Parliament and of the Council of 19 November 2008 on waste and repealing certain directives. E.U. Directive 2008/98/EC. eur-lex.europa.eu/legal-content/EN/TXT /?qid=1487004114713&uri=CELEX:32008L0098. E.U. 2009. Directive 2009/28/EC of the European Parliament and of the Council of 23 April 2009 on the promotion of the use of energy from renewable sources. E.U. Directive 2009/28/EC. eur-lex.europa.eu/legal -content/EN/TXT/?uri=celex:32009L0028. E.U. 2009. Directive 2009/125/EC of the European Parliament and of the Council of 21 October 2009 establishing a framework for the setting of ecodesign requirements for energy-related products. E.U. Directive 2009/125/EC. eur-lex.europa.eu/legal-content/EN/TXT/?qid=14870033 89208&uri=CELEX:32009L0125. E.U. 2010. Directive 2010/30/EU of the European Parliament and of the Council of 19 May 2010 on the indication by labelling and standard product information of the consumption of energy and other resources by energy-related products. E.U. Directive 2010/30/EU. eur-lex.europa.eu /legal-content/EN/TXT/?qid=1487003495623&uri=CELEX:32010 L0030. E.U. 2010. Directive 2010/31/EU of the European Parliament and of the Council of 19 May 2010 on the energy performance of buildings. E.U. Directive 2010/31/EU. eur-lex.europa.eu/legal-content/EN/TXT/?qid =1487003617334&uri=CELEX:32010L0031. E.U. 2012. Directive 2012/27/EU of the European Parliament and of the Council of 25 October 2012 on energy efficiency. E.U. Directive 2012/ 27/EU. eur-lex.europa.eu/legal-content/EN/TXT/?qid=1487003274569 &uri=CELEX:32012L0027. ICC. 2015. 2015 International Green Construction Code ® (IgCC ®). International Code Council, Washington, D.C. IES. 2011. The lighting handbook, 10th ed. Illuminating Engineering Society, New York. IPCC. 2014. Climate change 2014: Synthesis report, contribution of Working Groups I, II and III to the Fifth Assessment Report of the Intergovernmental Panel on Climate Change. R.K. Pachauri and L.A. Meyer, eds. Intergovernmental Panel on Climate Change, Geneva, Switzerland. www.ipcc.ch/report/ar5/syr/. ISO. 2006. Environmental management—Life cycle assessment—Principles and framework. Standard 14040-2006. International Organization for Standardization, Geneva. ISO. 2006. Environmental labels and declarations—Type III environmental declarations—Principles and procedures. Standard 14025. International Organization for Standardization, Geneva.
ISO. 2006. Environmental management—Life cycle assessment—Requirements and guidelines. Standard 14044. International Organization for Standardization, Geneva. Jones, D., S. Reilly, and M. Sadeghy. 2016. Actual energy performance of small office and K-12 school buildings designed to meet the 30% ASHRAE Advanced Energy Design Guides. ASHRAE Research Project RP-1627, Final Report. Kalogirou, S. 2004. Environmental benefits of domestic solar energy systems. Energy Conversion and Management 45:3075-3092. Lawrence, T., C.A. Balaras, and J.K. Means. 2016. A comparison of how sustainability and green building standards are being adopted into building construction codes within the United States and the EU. Proceedings of the SBE 16—Strategies—Stakeholders—Success Factors—International Conference on Sustainable Built Environment, Hamburg, Germany. NAS. 2010. Advancing the science of climate change. National Academy of Sciences, Washington, D.C. www.nap.edu/read/12782/chapter/1#v. NIST. 2017. BEES: Building for environmental and economic sustainability. National Institute for Standards and Technology. www.nist.gov/el /economics/BEESSoftware.cfm. Olivier, J.G.J., G. Janssens-Maenhout, M. Muntean, and J.A.H.W. Peters. 2015. Trends in global CO2 emissions—2015 report. PBL Netherlands Environmental Assessment Agency, The Hague, and European Commission, Joint Research Centre, Ispra, Italy. edgar.jrc.ec.europa.eu/news _docs/jrc-2015-trends-in-global-co2-emissions-2015-report-98184.pdf. Passer, A., S. Lasvaux, K. Allacker, D. De Lathauwer, C. Spirinckx, B. Wittstock, D. Kellenberger, F. Schösser, J. Wall, and H. Wallbaum. 2015. Environmental product declarations entering the building sector: Critical reflections based on 5 to 10 years experience in different European countries. International Journal of Life Cycle Assessment 20:1199-1212. link.springer.com/article/10.1007/s11367-015-0926-3. Srinivasan, R.S., W. Ingwersen, C. Trucco, R. Ries, and D. Campbell. 2014. Comparison of energy-based indicators used in life cycle assessment tools for buildings. Building & Environment 79:138-151. Townsend, T.E. 2006. The ASHRAE promise: A sustainable future. ASHRAE Journal 48(8):14-19. UK. 2015. HM Government, The Building Regulations 2010—Part G: Sanitation, hot water safety and water efficiency. Newcastle Upon Tyne: NBS. www.planningportal.gov.uk/buildingregulations/approveddocuments /partg/approved. UN. 1987. Our common future: Report of the world commission on environment and development. Annex to General Assembly document A/42/ 427, Development and International Co-operation: Environment. United Nations. www.un-documents.net/wced-ocf.htm. UNEP. 2003. Montreal Protocol handbook for the international treaties for the protection of the ozone layer, 6th ed., Annexes A, B, and C. Secretariat for the Vienna Convention for the Protection of the Ozone Layer and the Montreal Protocol on Substances that Deplete the Ozone Layer, United Nations Environment Programme, Nairobi. UNEP. 2014. Greening the supply chain. United Nations Environment Programme, Sustainable Buildings and Climate Initiative. www.unep.org /sbci/pdfs/greening_the_supply_chain_report.pdf. UNFCCC. 1998. Kyoto Protocol to the United Nations framework convention on climate change. United Nations Framework Convention on Climate Change, New York. unfccc.int/resource/docs/convkp/kpeng.pdf. UNFCCC. 2015. Adoption of the Paris Agreement. United Nations Framework Convention on Climate Change, New York. unfccc.int/resource /docs/2015/cop21/eng/l09r01.pdf.
BIBLIOGRAPHY ASHRAE. (Various.) Advanced energy design guides, 50% and 30% series. www.ashrae.org/standards-research--technology/advanced-energy -design-guides. ASHRAE. 2009. ASHRAE position document on climate change. ASHRAE. 2009/2011. Energy efficiency guides for existing commercial buildings, vol. 1: Technical implementation, and vol. 2: The business case for building owners and managers. ASHRAE. 2008. Vision 2020: Producing net zero energy buildings. www.ashrae.org/File Library/docLib/Public/20080226_ashraevision2020 .pdf.
Related Commercial Resources
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Related Commercial Resources CHAPTER 36
MOISTURE MANAGEMENT IN BUILDINGS Effects of Humidity and Dampness .......................................... Elements of Moisture Management ......................................... Envelope and HVAC Interactions............................................ Indoor Wetting and Drying...................................................... Vapor Release Related to Building Use...................................
36.1 36.1 36.2 36.2 36.4
Indoor/Outdoor Vapor Pressure Difference Analysis.............. 36.6 Avoiding Moisture Problems.................................................. 36.10 Climate-Specific Moisture Management................................ 36.11 Moisture Management in Other Handbook Chapters ............................................................................. 36.12
HE TERM moisture encompasses the gaseous, liquid, and solid states of water and any dissolved contaminants. Examples of liquid moisture include precipitation, wind-driven rain, construction moisture, rising damp, and water from incidental pipe drippings. Precipitation wets pitched roofs, low-slope roofs, and inclined facades, whereas wind-driven rain wets the enclosure as a whole. Buildings, including the envelope, start their service life containing significant quantities of construction moisture. This is particularly true for concrete, aerated concrete, mortar, and plaster. Groundwater and rain sinks may lead to rising damp, and pipe leakage reflects bad workmanship, lack of maintenance, failing fixtures, or pipe corrosion (Hens 2016; Mumovic and Santamouris 2009; Trechsel 1994). This chapter presents data on indoor vapor release and measured indoor/outdoor vapor pressure or vapor concentration differences not included elsewhere in the ASHRAE Handbook, and discusses moisture sources and sinks that can reduce materials’ durability, as well as the negative effects of insufficient or excessive indoor relative humidity. Gaseous water (vapor) comes from outdoor humidity and from interior vapor releases. Water vapor in the air, expressed as relative humidity, governs hygroscopic loading of materials. High relative humidity values at surfaces favor mold growth and, if reaching 100%, can lead to surface condensation. When vapor pressure and temperature gradients in and across assemblies point in the same direction, interstitial condensation is possible. Chapter 25 contains an in-depth analysis of heat, air, and moisture loads; Chapter 15 discusses surface condensation on windows; and Chapter 16 covers interstitial condensation caused by air leakage. Excessive indoor moisture and humidity interferes with the use and enjoyment of buildings and may shorten their useful life over the long term. Problems affecting owners and occupants include reduced comfort, poor indoor air quality, negative health effects, damage to the building’s materials and structural fasteners and wasted energy in HVAC operation. Consequently, moisture management demands attention from the architect, the builder, the mechanical system designer, and those charged with budgeting, management and maintenance of the building and its mechanical systems. All buildings experience occasional extremes in relative humidity and moisture. Short-term occurrences of these extremes can generally be accommodated by storage in the building materials, but when moisture and humidity accumulate for extended periods in vulnerable materials, major problems can and often do occur. Moisture problems are unfortunately quite common in buildings. It is the responsibility of those in a position of authority to reasonably reduce the risks associated with excessive moisture accumulation. Successful management of moisture and humidity requires understanding the complex and dynamic relationship between the building’s enclosure, its fabric, and the mechanical systems over the entire life of that building. This dynamic interaction holds the potential for either an excellent result
over decades, or for frequent, disruptive, and expensive problems. Experience suggests that human behavior can overcome virtually any building technology, so owner and occupant education are important elements in the successful design and usage of a building.
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T
The preparation of this chapter is assigned to TC 1.12, Moisture Manage-
ment in Buildings.
36.1 Copyright © 2017, ASHRAE
1.
EFFECTS OF HUMIDITY AND DAMPNESS
Moisture tolerance and appropriate indoor relative humidity levels must be considered requirements for a sustainable built environment: relative humidity affects comfort, indoor air quality, and health, and excessive wetness can shorten the service life of materials and assemblies. The preferred relative humidity range for human health and comfort is between 40 and 60%, although that interval is often broadened from 30 to 70%. High relative humidity degrades thermal comfort once the operative temperature passes 25 to 27°C, making the environment feel oppressive. It also facilitates release of volatile organic compounds (VOCs), especially of formaldehyde, thus degrading indoor air quality and triggering olfactory dissatisfaction. Finally, high relative humidity in specific environments (e.g., beds, on surfaces) activates dust mite reproduction and related allergy risks, and can activate mold germination and growth (ASHRAE 2012; IEAEBC 1990a, 1990b). Very low relative humidity activates electrostatic discharge and leads to complaints of dry mucous membranes (e.g., nose, lips, throat) and eyes, especially by people wearing contact lenses. On the other hand, some respiratory ailments can be relieved by dry environments. Excessively low or high relative humidity creates conditions favorable to bacterial and viral infections, allergic rhinitis, and asthma. Chapter 9 gives a more in-depth analysis of the impact of humidity on thermal comfort. Chapter 10 discusses the effects of relative humidity on indoor environmental quality, Chapter 11 covers mold, and Chapter 12 discusses the relationship between relative humidity and olfactory perception. Additional information is available in Holm (2008). Prolonged and excessive relative humidity and wetness can degrade materials physically, chemically, and biologically. Examples of physical degradation are frost damage and salt attack. Chemical degradation includes lime/gypsum reaction, carbonization of concrete, alkali/granulates reaction in concrete, and corrosion of ferrous and nonferrous metals, where moisture determines whether damage will occur in the presence of corrosive agents (e.g., sulfide, acetic acid). Wood rot by fungi and bacteria is an example of biological degradation. Very dry conditions also can damage wood, causing cracking and warping. Fluctuations between extremes of high and low relative humidity can induce cracking in hygroscopic materials.
2.
ELEMENTS OF MOISTURE MANAGEMENT
Designing for moisture and humidity management includes the choice of building materials and the layering of the envelope as well as the design and component selection of the HVAC system. For information on the building envelope, refer to Chapter 15; for details on building assemblies, see Chapters 25, 26, and 27.
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The largest contributors of liquid moisture are water from winddriven rain in building envelopes with insufficient drainage, leaks from roof or gutters, and leaks from internal plumbing. These sources must be addressed and resolved for the building envelope to succeed. Once liquid water is addressed, the next factors to consider are vapor pressure and relative humidity. The driving forces responsible for water vapor movement within buildings and across the envelope are air pressure differences that move air and vapor together, and vapor pressure differences that activate vapor diffusion. Temperature-, wind-, and fan-induced air-pressure differentials typically overwhelm diffusion in terms of total vapor flux displaced. However, in hot, humid climates or in buildings with high indoor/outdoor vapor pressure differences, pure diffusion may nevertheless cause problems. Chapters 16 and 20 contain information on air movement in buildings. Chapter 13 gives information on intrazone airflow, multizone network airflow, and contaminant transport, included vapor; and Chapter 24 discusses airflow around buildings. ANSI/ ASHRAE Standard 160-2009 provides criteria to evaluate the transient hygrothermal performance of envelopes. This analysis may be used to evaluate moisture tolerance in cases where, besides vapor, liquid water is a primary factor.
3.
ENVELOPE AND HVAC INTERACTIONS
As shown by Figure 1, heat, air, and moisture move continuously throughout a building, driven by differences in temperature, vapor pressure, and air pressure between the indoor and outdoor environment and by similar differences between adjacent indoor zones. Construction moisture contributes to the initial wetness of the whole building fabric, including the envelope. Over time, envelope assemblies exposed to rain and snow may become moist by seeping rain and melting snow or by capillary suction (rising damp) from moist below-grade earth, which may dampen walls just above grade. Leaking and dripping interior pipes can also be responsible for excessive wetness. The opaque and transparent portions of the building envelope must be designed so that their thermal transmittance approaches the energy related economic optimum. Together with the overall building fabric, the envelope should allow passive solar control, while all
assemblies proposed must prevent rain from wetting layers that must stay dry. The building details should withhold rainwater runoff and prevent sinking rainwater from wetting basements and floors on grade. Correct protective measures must exclude rising damp and promote effective construction moisture drying; water vapor moving across the envelope assemblies should not result in unacceptable interstitial moisture accumulation. Detailing and workmanship must minimize air leakage across the envelope as well as wind washing, indoor air washing, and air looping within envelope assemblies. Correct design, workmanship, and maintenance should prevent plumbing leakages or sweating of pipes running in or across walls and floors. The HVAC system must provide a thermally comfortable and healthy indoor environment. Ventilation is necessary for delivering fresh air to building occupants. To maintain a set-point temperature when heat losses by transmission, infiltration, and ventilation exceed solar gains through the transparent and opaque envelope parts and internal gains by lighting, appliances and occupancy, heating is required. If, instead, the gains exceed the losses, cooling is needed. In both cases, the temperature difference with outdoors creates a thermal stack effect, which together with wind and fan operation maintains air pressure differentials with the exterior. In heating mode, the indoor vapor concentration is typically left free floating, fluctuating with the vapor in the ventilation and infiltration air and the vapor released indoors. The net vapor concentration and the air temperature maintained by the HVAC system then determine the relative humidity indoors. Only when the ventilation and infiltrating air is too dry or too humid, when the vapor release indoors is very high, or when the building’s function requires relative humidity control, must the HVAC system actively intervene to control humidity by removing water vapor from or adding water vapor to the indoor air (dehumidification and humidification). The energy required to humidify or dehumidify is called the latent heat load. In cooling mode, the latent load derives from the supply air temperature required for cooling, the dew-point temperature of the outdoor air, and the vapor released indoors. Condensation on the coil removes some portion of the moisture in the incoming ventilation air.
4.
INDOOR WETTING AND DRYING
HVAC design or a durability assessment of the envelope and the whole building during design requires knowledge of the expected indoor humidity conditions. This is essential for making the right decisions to prevent mold, surface condensation, problematic interstitial condensation, and reduced drying. The function of a building and the conditions outdoors dictate whether the indoor relative humidity needs control. The vapor balance, which considers the various mechanisms involved in vapor release, vapor removal, and vapor storage, is fundamental for this decision. One of the unknowns in this balance is the vapor release related to building use. A first approach consists of quantifying these releases directly. A second approach to get information starts from extended indoor/outdoor climate measurements in large sets of buildings to deduce a statistically relevant indoor-to-outdoor timeaveraged difference in vapor pressure or vapor concentration in relation to the outdoor air temperature. Both methods are discussed in the following.
Understanding Vapor Balance The air humidity indoors is assumed to be free floating. Five vapor fluxes then contribute to maintaining the equilibrium in a room:
Fig. 1 Dynamic Interaction Between Air, Moisture, and Materials in HVAC Systems and Building Envelope
• Vapor carried and supplied by infiltrating outdoor air and ventilation • Vapor removed by exfiltrating indoor air and exhaust air • Vapor released by occupants and their activities, by plants, water surfaces, and drying fabric parts
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Moisture Management in Buildings
36.3
• Vapor adsorbed and desorbed by all hygroscopic surfaces present • Vapor condensing on indoor surfaces that are colder than the dewpoint temperature indoors or evaporating from indoor surfaces wetted by surface condensation These five fluxes load and unload the room air with vapor. There is even a sixth flux: vapor inflow and outflow by diffusion through the envelope. Its magnitude, however, is insignificant compared to the five fluxes mentioned, hence ignoring it does not introduce any significant error. Assuming ideal air mixing, the balance per room is expressed by
Ga ,in xv ,in + Ga ,out xv ,i + j Aj psat , A A – pi j
j=1
+
k Ak psat , A m
k=1
k
dx v ,i – p i + G v , P = V ----------dt
(1)
where
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= 0
(2)
If the only incoming airflow is from the outdoors (e.g., infiltration, natural ventilation, air-side economizer applications), equality between supply and exhaust yields the mean indoor relative humidity: R 273.15 + i G v , P 1 = ------------ o p sat , o + ----------------------------------------------· p sat ,i Va
(3)
where
n
j
Ga , in xv , in + Ga , out xv ,i + Gv , P
Ga,in = all airflows entering room, including infiltration from outdoors, ventilation supply air, and infiltration from adjacent rooms, kg/s xv,in = ratio of water vapor to dry air in entering air, kg/kg Ga,out = all airflows leaving room, including exfiltration to outdoors, exhaust air, and exfiltration to adjacent rooms, kg/s xv,i = ratio of water vapor to dry air in room, kg/kg j = surface film coefficient for diffusion at each hygroscopic surface, kg/(m2·s·Pa) Aj = area of each hygroscopic surface in room, m2 pi = water vapor pressure in room, Pa psat,Aj = vapor saturation pressure at different hygroscopic surfaces in room (fabric, furniture, and furnishings), Pa sat,Aj = relative humidity at different hygroscopic surfaces (fabric, furniture, and furnishings) in room on a scale from 0 to 1 k = surface film coefficient for diffusion at all room surfaces where vapor condenses or evaporates, kg/(m2·s·Pa) Aj = area of all room surfaces experiencing surface condensation or surface condensate drying, m2 psat,Aj = vapor saturation pressure for room’s various condensing and evaporating surfaces, Pa Gv,P = vapor release in room, kg/s V = air volume of room, m3 t = time, s
Solving Equation (1) requires knowledge of all in- and exfiltrating airflows between the room and all adjacent rooms and between the room and the outdoors, the ventilation supply airflow, exhaust airflow, water vapor ratio transported by all four airflows, surface temperatures and vapor balances at all hygroscopic surfaces, surface temperatures at all condensing and drying surfaces, and the surface film coefficients for diffusion at each of these surfaces (Hens 2012). Calculating requires numerical methods (preferably with dedicated software) to account for the transient nature of the variables involved (Woloszyn and Rode 2008). In reality, because of stack effects, supply air jets, and convective plumes around people, the air in a room is never fully mixed, and vapor pressure gradients exist inside every room. Wetter, warmer air is also lighter and creates a positive temperature and humidity gradient along a room’s height, even in stagnant conditions. Despite these internal gradients, whose quantification requires computational fluid dynamics (CFD) calculations, further discussion is based on the fully mixed assumption. When considering steady-state conditions where the room air maintains an average humidity that does not change with time (e.g., because sorption/desorption by the hygroscopic surfaces no longer intervenes and all airflows and vapor releases are time averaged), in the absence of surface condensation and surface drying, the humidity balance per room simplifies to
i psat,i o psat,o R i Gv,P · Va
= = = = = = = =
indoor relative humidity on scale from 0 to 1 vapor saturation pressure for indoor temperature, Pa outdoor relative humidity on scale from 0 to 1 vapor saturation pressure for outdoor temperature, Pa gas constant for water vapor, equal to 462 (Pa·m3)/(kg·K) indoor temperature, °C average indoor vapor release, kg/h outdoor air ventilation or/and infiltration flow, m3/h
Relative humidity over longer periods of time thus depends on mean outdoor temperature (through psat,e), mean outdoor relative humidity, mean indoor air temperature (through psat,i), mean indoor vapor release, and mean outdoor air ventilation and/or infiltration flows. In cold and temperate climates, the outdoor air ventilation required by building standards easily dilutes the vapor released in residences, offices, and schools. In temperate climates, heating tends to keep relative humidity levels below the upper acceptability threshold. In very cold climates, ambient temperature and outdoor relative humidity may be such that normal ventilation rates cannot raise the indoor relative humidity above 15 to 20% at room temperature. Humidification may then become necessary. In hot, humid climates, outdoor temperature and relative humidity can be so high that the combined action of ventilation, cooling, and vapor release raises the indoor relative humidity above the upper acceptability threshold. Supplemental air dehumidification then is required. Where outdoor air infiltrates into the room and the ventilation system delivers supply air at a fixed condition, the mean relative humidity indoors becomes =
· · 1 V a ,inf p sat , o o + V a , s p sat , s s R 273.15 + i G v , P ------------ ----------------------------------------------------------------------- + ----------------------------------------------· · · · p sat , i V a , inf + V a , s V a ,inf + V a , s
(4)
where o psat,o s psat,s · V a ,inf · V a ,s
= = = = = =
outdoor relative humidity on scale from 0 to 1 vapor saturation pressure for outdoor temperature, Pa relative humidity of supply air on scale from 0 to 1 vapor saturation pressure for supply air temperature, Pa infiltrating volumetric outdoor airflow, m3/h volumetric supply airflow, m3/h
Hygric Buffering In hygric buffering, vapor from the air is adsorbed into hygroscopic materials when the relative humidity rises, and released back into the air when the relative humidity lowers. On shorter time scales, hygric buffering by various elements such as indoor air and the building fabric (e.g., envelope, furniture, furnishings) dampens and shifts fluctuations in relative humidity and consequently in indoor vapor pressure compared to the fluctuations of the exterior vapor pressure and internal vapor release. Figure 2 compares measured indoor vapor pressures in an office building with the exterior vapor pressure over the course of a winter month when exterior conditions changed from cold and dry to warmer
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Fig. 4 Comparison of Daytime Relative Humidity for Summer and Winter Case (Antretter et al. 2012)
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Fig. 2
Measured Water Vapor Pressure Outdoors and Indoors for Office Building
Fig. 5 Annual Monthly Averaged Indoor/Outdoor Vapor Pressure Difference in Bedroom of Figure 3
Room is 16 m2 with air volume of 40 m3. Walls and ceiling finished with papered gypsum plaster, with (solid curve) and without (dashed curve) its moisture-buffering effects taken into account.
Fig. 3 Daily Vapor Pressure in Two-Person Bedroom and more humid (Hens 2009a). Indoor vapor pressure remained higher than the outdoors during the cold, dry spell but dropped lower than outdoors during the warmer and more humid spells. Figure 3 shows the calculated buffering effect on the daily indoor vapor pressure in a well-ventilated two-person bedroom during winter (Hens 2012a). The effects of hygric buffering in keeping the daily changes in vapor pressure much lower than if only the air acted as buffering volume are obvious. Antretter et al. (2012) performed a series of buffering experiments in a test room, simulating 4 to 6 day periods of summer and winter conditions. For the summer simulation, conditions were kept constant at 20°C, 90% rh, and 1 air change per hour (ach) for 1 to 3 days, then altered to an air change rate of zero and brought up to 35°C in 2 h. These conditions were held for 10 h, after which the temperature was returned to 20°C in 2 h and held there for the next 3 days. For the winter simulation, conditions were kept constant at 23°C, 50% rh, and 1 ach for 1 to 3 days, followed by 3 days with two vapor release peaks and a drop in the air change rate from 1 to 0.5 ach. The reference experiment was initially done without moisture buffering, and then repeated with 27 0.3 × 0.3 m sorptive tiles in the room. Figure 4 shows how the relative humidity changed the day after the initializing period. The effect of buffering is evident. Hygroscopic buffering can even induce dampening and phase shifting on an annual basis in buildings with massive construction,
as shown by Figure 5 for the same bedroom as in Figure 3 (Hens 2012a). For the same vapor release indoors and constant ventilation rate, the monthly mean indoor vapor pressures form an inclined ellipse, with the highest value in winter, lowest in summer, and values lower in springtime than in autumn.
5.
VAPOR RELEASE RELATED TO BUILDING USE
Water vapor indoors comes from several sources: building occupants and their activities (e.g., cleaning, cooking, showering, bathing), pets, plants, water surfaces, drying of wet fabric parts, etc. Releases can be divided into those dependent on indoor temperature and relative humidity, such as free water surfaces and wet fabrics, and those generally independent of these factors, such as human activities and plants (TenWolde and Pilon 2007). Accounting for these two types of sources, Equation (3) expands into 1 i = -----------p sat , i
(5) · o p sat ,o + R 273.15 + i G vp + A w psat , w V a ----------------------------------------------------------------------------------------------------------------------------· 1 + R 273.15 + i A w V a
where Aw is the free water surface, in m2. Solving Equations (3) and (4) presumes knowledge of the magnitudes of the water vapor sources.
Residential Buildings Estimating moisture loads requires knowledge of occupancy patterns and releases per individual, activity, plant, and water surface present. In most cases, the real occupancy is unknown, and the sources are even more difficult to quantify. Some activities in the same classification differ in the amount of vapor released (e.g.,
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36.5
Table 1 Vapor Released by Humans, Human Activities, and Plants Source
Units
Release
Humans
g/h g/h g/h g g g g g g g g g g g/h per pot g/h per pot
30-60 120-200 200-300 60 660 160 to 270 250 to 320 550 to 720 100 70 310 60 260 10 15
g/h g/h g/h
20-200 100-500 2130-2900
Light activity Medium activity Hard work Bathroom Bath (15 min) Shower (15 min) Breakfast preparation for 4 people Lunch preparation for 4 people Dinner preparation for 4 people Breakfast dish washing for 4 people Lunch dish washing for 4 people Dinner dish washing for 4 people Simmering pot (diameter 150 mm, 10 min) Boiling pot (diameter 150 mm, 10 min) Potted flowers Potted plants
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Laundry Already spin-dried, until dry Dripping wet, until dry Unvented drier, until dry
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
Table 2 Daily Vapor Release by Humans, Human Activities, and Plants: Data from Three Countries Release Source, units
U.K.
United Denmark States
Gv,P = 0.34 + 3.88n (r 2 = 0.999) in kg/day
(6)
The individual data, however, produce a different relationship with considerably more scatter: Gv,P = 3.23 + 2.82n (r 2 = 0.999) in kg/day
(7)
In both equations, n is the number of family members weighted by the ratio between the hours each is at home and 24 h a day. For example, if four people occupy the home for 16 h daily, n is
16 n = 4 ------ = 2.7 24 Equation (7) indicates that each person who is at home 24 h/day adds 2.82 kg of water vapor to the indoor ambient. Measurements in some German apartments, however, suggested that this release rate is too high: for a family of three, the rates documented ranged from 5.6 to 7.8 kg per day (Hartmann et al. 2001), giving a value between 0.8 and 2.48 kg per person per day. Data from Estonia suggest a person present 24 h/day adds on average 1 kg to the vapor released (Kumaran and Sanders 2008). Table 7 provides some country-related statistical information about the vapor release expected in different types of residential buildings.
Natatoriums Evaporation from the pool surface, wet swimmers, and the wet floor around the pool are the main sources of water vapor. The evaporation rate is calculated based on the pool surface, multiplied by a factor that accounts for the average number of pool users and the wet floor around the pool: Gv,P = f Apool ( psat, pool – pi)
(8)
Humans, kg/person per day
1.2
0.9
1.25
where
Cooking for 4 people using electricity, kg/day
2
0.9
1.2
Gv,P = evaporation rate, kg/s = surface film coefficient for diffusion at water surface in pool, kg/ (m2 ·s·Pa) f = factor that accounts for average number of pool users and wet floor around pool Apool = water surface area, m2 psat, pool = vapor saturation pressure at pool’s water temperature, Pa
Using gas, kg/day
3
Dishwashing, kg/day
0.4
0.4
2.5 0.5
Bathing/washing, kg/person per day
0.2
0.4
0.25
Washing clothes, kg/day
0.5
Drying clothes, kg/day
1.5
1.8
2.2
Mopping floors, kg/day
0.2
Plants, kg/plant per day
0.02
Refrigerator defrost, kg/day
0.05 0.5
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
Table 3 Fuel
Vapor Released by Fuel Burning Vapor Released, kg/kWh
Natural gas
0.15
Manufactured gas
0.10
Paraffin
0.10
Coke
0.03
Coal
0.01
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
cooking, which varies with the type of meals prepared). Literature sources only give global average numbers and should therefore be used with care and understanding of the implicit assumptions. Tables 1 to 5 list example amounts by type (IEA-ECB 1990; Kumaran and Sanders 2008; Sanders 1996; TenWolde and Walker 2001). The means in Table 6 yield, as least-square regressions,
In cases where the airflow comes from outdoors, Equation (9) can be combined with Equation (3) to yield the mean vapor pressure in a natatorium: · o p sat ,o + p sat , pool R 273.15 + i fA pool V a p i = ---------------------------------------------------------------------------------------------------------------------· 1 + R 273.15 + i fA pool V a
(9)
Measurement of the water surface, water temperature, air temperature, relative humidity, ventilation supply flow, and number of users in a university and a recreation natatorium allowed estimating a surface film coefficient for diffusion , with inclusion of the factor f (Table 8). For the university natatorium, the table includes all measured evaporation rates in g/(m2 ·h). Using the data from the university pool in Table 8, the unknown factor f could be derived as a function of the number of pool users (Hens 2009b): f = 1 + 8.46n (r 2 = 0.95)
(10)
where n is the number of users per square metre of pool surface. For zero users, f is 1. The least-square line is depicted in Figure 6. For example, given a water surface of 50 × 20 m, a water temperature of 28°C, an air temperature of 30°C, an indoor relative humidity of 60%, and 100 pool users, evaporation amounts to 80 kg/h.
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36.6
2017 ASHRAE Handbook—Fundamentals (SI) Table 4
Time, h
Number of Occupants
1-5 6 7 8-17 18 19-10 21-22 23 24
2 2 2 0 2 2 2 2 2
Vapor Release for Family of Two, Both Working, Weekday Schedule Vapor Released, g/h Occupants
Cooking
Hygiene
240 240
720
120 120 120 120 120 120 120 120
Sum/day
Washing
Sum, g 600 1080 360 0 120 1200 240 600 120
480 480
1680
1440
1200
4320
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
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Table 5 Vapor Release for Family of Four, One Parent and One Child at Home, Weekday Schedule Time, h
Number of Occupants
1-5 6-7 8 9 10 11-12 13-14 15 16-17 18 19 20 21-22 23 24
4 4 2 1 1 2 2 2 2 4 4 4 4 4 4
Vapor Released, g/h Occupants
Cooking
240 240 120 120 120 120 120 120 180 240 240 240 240 240 240
Sum/day
Hygiene
480
Washing
Sum, g 240 2880 240 300 1020 2880 1440 240 600 240 720 960 480 480 240
720 120 180 180
720 1200 480
120 120 120 120
480 480
240 240
4680
6000
2400
840
13 920
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
Table 6
Daily Vapor Release in Relation to Number of Family Members Family Members
Vapor release, kg/day
2 (no children)
3 (1 child)
8
12 10 20
7 13.2
Mean
4.3 8.2 8.14
19.9 11.5 5-12 6-10.5
14 + 1 kg/day per child
14.6 23.1
Canada Detached Denmark Flats/mechanical ventilation Detached/mechanical ventilation Detached/natural ventilation Sweden Detached Multifamily dwelling units
10%
50%
90%
3.0 2.5 3.5 3.0 4.6 2.4
8.7 5.8 8.0 6.5 8.7 5.6
21.0 10.0 13.5 13.0 15.7 11.0
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
12.1
13.7 14.1
14.4
11.9
15.9
—
Sources: IEA-ECB (1990), Kumaran and Sanders (2008), Sanders (1996), and TenWolde and Walker (2001).
6.
Vapor Release Rate (kg/day) Country Houses
4 4 (2 children) (2 children) 14
Table 7 Vapor Release Rates by Percentile
most experiments looking to quantify air humidity in buildings use the average indoor/outdoor vapor pressure difference pio or average indoor/outdoor vapor concentration difference io, represented by the second term in Equation (3) (i.e., the ratio between the average water vapor release and the average infiltration and ventilation flow):
INDOOR/OUTDOOR VAPOR PRESSURE DIFFERENCE ANALYSIS
Measuring vapor release rates directly is a difficult and timeconsuming task. The typical alternative is to monitor the indoor temperature and relative humidity and then compare the related mean indoor vapor pressure to the mean value outdoors. Therefore,
R 273.15 + i G v , p in Pa p io = ----------------------------------------------· Va
(11a)
Gv , p 3 io = ---------· - in g/m Va
(11b)
or
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36.7
Table 8 Measured Surface Film Coefficients for Diffusion, Related to Pool Surface Natatorium 1 (university pool, water surface 325 m2, water temperature between 27 and 28°C)
Pool Users
Evaporation g/(m2 ·h)
Water surface covered Uncovered
0 0
38 79
Water surface covered Uncovered
0 0
69 112
Pool in use, teaching revalidation free swim 1 50+ swim free swim 2 free swim 3
13 29 42 13 65 75
125 119 163 141 172 167
Natatorium 2 (recreation pool, 60 m long chute, waterfall, preschooler pool, massage pool, and whirlpool. Total water surface 483 m2, water temperature between 30 and 31°C)
kg/(m2 ·s·Pa) Measuring period 1 4.9E-9 9.7E-9 Measuring period 2 6.5E-9 13E-9 Measuring period 3 15E-9 14E-9 23E-9 19E-9 27E-9 31E-9
(1) = water uncovered (1) + chute (1) + waterfall (1) + preschooler pool (1) + massage pool (1) + whirlpool
Evaporation g/(m2 ·h)
kg/(m2 ·s·Pa)
171 — — — — —
15E-9 17E-9 18E-9 19E-9 15E-9 16E-9
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Source: Hens (2009b).
Fig. 6 Factor f as Function of Pool User Density That difference over a given time span is mostly linked to the average air temperature outdoors over the same time span.
Residential Buildings In North America and Europe, there have been many monitoring programs in residential buildings that gathered data on the indoor/ outdoor vapor pressure and vapor concentration difference. The least-square linear relationships found between these data, averaged over a given time span, and the outdoor air temperature averaged over the same time span typically show a decreasing trend at higher temperatures. Figure 7 shows such a measured relationship for a region where massive construction prevails and homes tend to be naturally ventilated (Hens 2016). There are two reasons for the decrease shown in Figure 7: (1) windows are more frequently opened in warmer weather, resulting in more outdoor air ventilation and thus lower indoor/outdoor vapor pressures and concentration differences at higher outdoor air temperatures; and (2) the massive building fabric that absorbs and releases significant amounts of vapor and consequently dampens
Fig. 7 Daytime Rooms in Dwellings and shifts the indoor vapor pressure compared to the outdoors even on an seasonal basis. Ridley et al. (2008) took long-term measurements in 1065 U.K. living rooms and 916 U.K. bedrooms. The least-square regressions, based on half-hour averages, are shown in Figures 8 and 9. They are given by Living rooms: pie = 400 – 19 e in Pa
(12)
pie = 425 – 23.6 e in Pa
(13)
Bedrooms:
Figure 10 and Table 9 are based on measurements in 101 dwellings in Finland and Estonia (Kalamees et al. 2006). The data gained were approximated by three distinct linear segments with the running weekly average indoor/outdoor vapor pressure differences constant below5°C and above 15°C outdoors and varying linearly in
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36.8
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Fig. 8
2017 ASHRAE Handbook—Fundamentals (SI)
Indoor/Outdoor Vapor Pressure Difference in 1065 U.K. Living Rooms
Fig. 10 Water Vapor Pressure Excess in Relation to the Running Weekly Mean Temperature for Northern Europe and Canada
Fig. 9
Indoor/Outdoor Vapor Pressure Difference in 916 U.K. Bedrooms
Table 9 Finland and Estonia, Indoor Climate, Boundaries (Weekly Means) Indoor/Outdoor Vapor Pressure Difference, Pa Humidity Load
e ,w 5°C
Low
540
Average
680
High
810
5°C e ,w 15°C 0.388 – 0.00556 e ,w 540 – 34 e ,w 0.473 – 0.00667 e ,w 680 – 41 e ,w 0.559 – 0.00778 e ,w 810 – 47 e ,w
e ,w 15°C 200 270 340
between. A distinction was also made between high, average, and low vapor load. The data collected in northern Canada (Kumaran and Sanders 2008) are represented by dots in Figure 10. The average indoor/outdoor vapor pressure the dot at top left represents a four-day period where the average outdoor temperature was –18°C and is substantially higher than the Finnish and Estonian high humidity results,
Fig. 11
Indoor/Outdoor Vapor Pressure Differences for 10 German Living Rooms
apparently because a boil-water safety advisory led to an extremely high vapor release. The other dot at 13°C outdoor temperature is close to the Finnish and Estonian low-humidity curve, even though the number of household members was higher than typical for southern Canada. The dwellings monitored were measured to have air change rates up to 12 ach at 75 Pa. Measurements in the living rooms of 10 German homes are shown in Figure 11. The main variable seems to be the number of household members and their living habits, including window operation (Antretter et al. 2010). Comparing the mean of the German averages with those from the U.K. in Figures 8 and 9 gives smaller differences at low outdoor temperatures, and higher differences at higher outdoor temperatures: p ie = 328 – 13.7 e in Pa
(14)
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36.9
Temperature and relative humidity data collected in 60 homes, equally spread over three climate zones in the United States (Pacific northwest, a cold location in the northeast, and a hot and humid location in the southeast) gave the values plotted in Figure 12 as the average relation between the moving monthly mean outdoor temperature and the moving monthly mean indoor/outdoor vapor pressure difference (Arena et al. 2010). Figure 12 clearly shows the dehumidifying effect of air conditioning. The least-square regression line for a running monthly mean outdoor temperature below 19.5°C equals pie = 397 – 16.3 e (r 2 = 0.86)
(15)
climate, and between April and November in the warmer, more humid Knoxville climate. Both situations are the result of dehumidification by air conditioning. Measurements taken in 71 homes in Rhode Island are shown in Figure 14, giving the distribution of the measured indoor/outdoor vapor pressure difference at 0°C outdoors (Francisco and Rose 2010). The mean was 396 Pa, close to what Equation (15) suggests, whereas the 95th percentile value equaled 772 Pa. This is higher than the data measured (Hens 2016) in the temperate climate of northwestern Europe, where for larger homes the 95th percentile on a weekly basis was p ie = 540 – 29 e (Pa)
Above 19.5°C, the line becomes pie = 2807 – 140 e (r 2 = 0.92)
(16)
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Indoor/outdoor vapor concentration difference measurements in 10 homes in Madison, WI, and 10 homes in Knoxville, TN, over a one-year period gave as monthly means the values shown in Figure 13 (Antretter et al. 2010). The data are consistent with what Figure 12 shows: vapor pressure deficits only occur during the summer months in the Madison
(17)
whereas for small homes the relation was p ie = 670 – 29 e (Pa)
(18)
To minimize risk, the 95th percentile is often used for design purposes. These results are based primarily on homes using natural ventilation. Many codes in the United States require homes to add mechanical ventilation to compensate for tighter envelopes. This significant difference must be addressed in evaluating residences with mechanical ventilation. Measurements in 30 homes by Francisco et al. (2009) found that unvented gas fireplaces added on average about 100 Pa to the indoor/outdoor vapor pressure difference.
Natatoriums Vapor pressure difference measurements in natatoriums are few; however, studies have been conducted in some countries. Figure 15 gives weekly mean indoor/outdoor vapor pressure differences in relation to the weekly mean outdoor air temperature, measured in 20 natatoriums in a temperate region (Hens 2016). There is a considerable amount of scatter. Least-square regressions for the mean and 95th percentile are Mean p ie = 1228 – 27.7 e ,w (r 2 = 0.13)
(19)
95th percentile Fig. 12 Monthly Mean Indoor/Outdoor Vapor Pressure Difference in Relation to Monthly Mean Outdoor Air Temperature in Three U.S. Climate Zones
Fig. 13 Measured Monthly Mean Indoor/Outdoor Vapor Concentration Difference in 10 Homes in Madison, WI, and 10 Homes in Knoxville, TN (Antretter et al. 2010)
p ie = 1841 – 27.7 e ,w (r 2 = 0.13)
Fig. 14
(20)
Indoor/Outdoor Vapor Pressure Difference with Intersect at 0°C for 71 Rhode Island Homes
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36.10
2017 ASHRAE Handbook—Fundamentals (SI)
The data clearly show that, although not highly correlated with the outdoor temperature, the vapor pressure differences in natatoriums are consistently high. These buildings therefore require specific measures to prevent moisture accumulation in the envelope, which can lead to degradation (Figure 16).
Student Residences and Schools The few measuring campaigns undertaken in student residences and schools in a temperate-climate region gave a very diffuse picture, with a large spread in weekly mean indoor/outdoor vapor pressure differences; see Figure 17 (Hens 2005) and Table 10 (Hens et al. 2007).
7.
AVOIDING MOISTURE PROBLEMS
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Avoiding mold, surface condensation, and interstitial condensation in cold and temperate climates requires the following measures:
• Building envelopes must be insulated such that the risk of surface mold and surface condensation at room temperature and typical levels of relative humidity is as close to zero as possible. Insulation requirements imposed by current building codes for energy efficiency are typically much higher than those needed to address risk of condensation. • Avoid structural thermal bridges. No point on the indoor surface of an opaque portion of the envelope should be colder than the indoor surface of the glazed portion • Minimize infiltration and exfiltration by making the envelope as airtight as possible. • Avoid air gaps in envelope construction that would lead to wind washing in and along the inner face, and indoor air washing along the outer face of the thermal insulation. • For low-mass construction in heating-dominated climates, if necessary, mount a continuous vapor retarding layer at the inner side of the insulation. In hot, humid climates, install the vapor control on the exterior or warm side of the insulation or building envelope. • Vapor control layers should be used with caution in stone and masonry construction. Take measures to avoid problems with solar-driven vapor flow when applying vapor retarders on the inner side of envelopes that have a rain-buffering outer finish, such as a brick veneer [see Derome and Saneinejad (2009)]. • Never sandwich layers between vaportight materials on both sides. This is especially detrimental when thermal insulation is one of the layers and the others contain construction moisture. Table 10 Indoor Air Temperature and Indoor/Outdoor Vapor Pressure Difference: Means and Extremes Measured in Five Temperate-Climate Schools Temperature, °C
Fig. 15 Weekly Mean Indoor/Outdoor Vapor Pressure Differences for 20 Natatoriums (Measured Data and Least-Square Straight Line)
Fig. 16 Natatoriums: (A) Low-Sloped Roof Damaged by Convection-Induced Interstitial Condensation; (B) Interstitial Condensation in Low-Sloped Roof Polyurethane Foam Insulation; (C) Timber Beam Collapse; (D) Abundant Surface Condensation on Window and Lintel
Indoor/Outdoor Vapor Pressure Difference, Pa
School
Mean
Min
Max
Mean
Min
Max
1
17.7
14.7
23.2
197
–138
722
2
15.0
12.0
26.8
215
–155
800
3
20.3
14.2
25.2
198
–266
1059
4
17.5
10.3
23.1
151
–169
609
5
17.9
14.1
23.4
36
–404
496
Fig. 17 Weekly Mean Indoor/Outdoor Vapor Pressure Differences in Four Student Residences
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36.11
HVAC Systems The mechanical system can affect the air pressure differences among building zones, and between building zones and the outdoors. HVAC operation influences the indoor relative humidity and, consequently, indoor/outdoor vapor pressure differences. To minimize humidity concerns, • Insulate chilled-water pipes and cool-air ducts to minimize heat gain. The insulation must have a moisture resistant vapor-control layer on the outer surface and be thick enough to prevent surface condensation. Warm-water pipes and warm-air ducts require insulation to minimize heat losses. For more details, see Chapter 23. • Natural or mechanical ventilation in temperate climates should keep the indoor relative humidity within the 30 to 70% range. • HVAC systems in hot and humid climates should ventilate the building while cooling and dehumidifying the ventilation air. The building should also be positively pressurized. In very cold climates, HVAC systems should ventilate and warm the building while humidifying the air and putting the building under negative pressure.
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Ground Pipes
Class I substrates are biologically recyclable building materials: wallpaper, plasterboard, permanent elastic joint materials, and other biodegradable materials.
Fig. 18
Sedlbauer’s Isopleth System for Class I Substrates: Time Until Germination (Sedlbauer 2001)
In temperate climates, ground pipes (ventilation air ducts buried in the soil near the building) are sometimes used to moderate the condition of incoming ventilation air. Rain and groundwater must be prevented from entering these pipes: water penetration turns the pipes into air humidifiers, with very negative consequences for the indoor air quality. In dry pipes, relative humidity during the warm season should not exceed the mold threshold: too high a relative humidity in the pipes can load the supply air with mold spores, which then germinate in the supply filter. These risks are high enough that ground pipes are not recommended as a safe choice (Hens 2012b).
Building Fabric If there is no watertight layer or capillary break inserted above grade, stone-based building fabrics may suffer from rising damp and salt transport. Building parts, wetted by rising damp, act as permanent moisture sources affecting indoor vapor release and the indoor/outdoor vapor pressure difference.
Building Envelope Even when raintightness is guaranteed, excessive indoor vapor pressures can induce mold and condensation on interior surfaces and favor interstitial condensation in assemblies. Chapters 25 and 27 deal in more detail with moisture transport in the building envelope. Mold Growth. Mold growth begins on interior opaque surfaces where the surface temperature and relative humidity pass the mold threshold for a long enough period of time. For example, if a material has absorbed an amount of water from a leak, drying may temporarily sustain a surface relative humidity high enough to be conductive for mold. The IEA Annex 14 reports (IEA-EBC 1990a, 1990b) contain two design formulas for estimating mold risk. The one calculates the four-week threshold, above which the likelihood that mold will develop is close to 100%. The other evaluates the surface relative humidity needed for mold development at shorter time intervals. Combining the two yields
– 0.072 ln T where crit = relative humidity at surface, on scale from 0 to 100
The threshold for a one-day period touches 99%. Since completion of Annex 14, two more comprehensive models have been published: Sedlbauer (2001) (Figure 18), and Viitanen and Ojanen (2007). More recent work, however, confirms that predictions with any model give varying results for growth probability and mold density (Vereecken and Roels 2012). Surface Condensation. Each time the indoor surface temperature somewhere on the envelope drops below the dew point of the indoor air, condensate forms. Glazed surfaces are especially prone to surface condensation. Hygroscopic materials (e.g., wood, masonry, gypsum, some plasters) may buffer the local surface wetness, preventing visible condensate but increasing the moisture content in the material. When short periods of condensation or high humidity alternate with periods of drying, there tend to be no negative consequences. Long-lasting surface condensation or long periods of high humidity, however, degrade timber and other materials, and condensate deposited on aluminum frames may wet the window reveals. Thermal bridges can experience prolonged periods of condensation that cause indoor-side finish damage, such as wallpaper glue giving away. Interstitial Condensation. Water vapor condensation within envelope parts is called interstitial condensation, and is caused by temperature and vapor pressure gradients pointing in the same direction. The driving forces are diffusion driven by vapor pressure differences, moist air leakage (in- or exfiltration), or absorption of wind-driven rain. In wood and wood-based materials, interstitial condensation may lift the surface relative humidity above the mold threshold and, worse, the moisture content above the rot threshold. When the relative humidity adjacent to nonporous, nonabsorbing, or capillary wet layers in an assembly becomes 100%, condensation with droplet formation occurs, followed by runoff and sometimes dripping, which is especially annoying for the occupants.
8.
2
crit min{100 ; 0.033 si – 1.5 si + 96 1.24
si = indoor surface temperature, °C (5°C si 25°C) T = time span, days
(21)
CLIMATE-SPECIFIC MOISTURE MANAGEMENT
Temperate and Mixed Climates In temperate and mixed climates, neither humidification nor dehumidification is explicitly required except when necessary for
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36.12 functional reasons, such as in museums and operating rooms. Some depressurization during winter is advisable to minimize interstitial condensation risk, such as using demand-controlled exhaust ventilation or balanced ventilation with heat recovery and a correctly chosen pressure balance between supply and exhaust. The envelope must be so airtight that depressurization will not result in ventilation airflow beyond that needed for health reasons. High air permeance (low airtightness) has additional drawbacks: sound insulation degrades, drafts can occur, heat loss increases, and indoor surface temperatures might lower to a point where mold develops or surface condensation appears. Ground floors above crawlspaces must be airtight enough to prevent soil moisture, crawlspace odors, and radon from infiltrating into the conditioned space. Acceptable envelope performance requires that the indoor surface temperature be high enough to prevent mold from occurring at the expected average vapor release indoors, the environmental conditions outdoors, the desired thermal comfort settings and the mean ventilation rate required for good IAQ and human health. Ventilation is also used to avoid acute surface condensation, which occurs each time the indoor dew-point temperature exceeds the envelope’s indoor surface temperature somewhere (a common site is glazed surfaces). For a detailed analysis of mold and surface condensation, see Hens (1999), Sedlbauer (2001), Vereecken and Roels (2012), and Viitanen and Ojanen (2007). With well-insulating, double- and triple-glazed, gas-filled, low-e windows in new construction and retrofits, water can also condense on the outdoor surface during cold, clear-sky nights, somewhat obscuring visibility.
Hot and Humid Climates In hot and humid climates, to avoid mold growth and surface condensation, the indoor relative humidity should be kept below 60% through dehumidification. Brief periods of elevated relative humidity, however, do not necessarily result in mold germination. Keeping the building under positive air pressure is required to minimize interstitial condensation in wall cavities. Dehumidification. Air-handling units that mix outdoor air with return air to provide constant airflows at variable supply temperatures do not provide continuous dehumidification. At higher supply air temperatures, the cooling coil does not remove moisture from the ventilation air, which can lead to elevated indoor relative humidity and a high potential for microbial growth. This becomes even more pronounced in energy-efficient buildings with reduced sensible cooling loads and long off times for the cooling coil, which significantly reduces the effectiveness of moisture removal (Rudd 2010; Rudd et al. 2005). For variable-supply-temperature systems, the ventilation air portion can be preconditioned using a dedicated outdoor air system (DOAS). These systems dehumidify only the ventilation air, and cooling loads are handled by a separate air-handling unit, so a DOAS is a good choice for providing continuously dehumidified ventilation air. If heat recovery is economically feasible without compromising pressurization, a DOAS with heat recovery is an energy-efficient alternative if the supply and return fans maintain correct pressurization, and the relative humidity of air passing through the supply filters remains low enough to avoid mold growth there. Because schools, office buildings, hospitals, and dwellings require continuous ventilation based on the number of occupants, demand-controlled DOAS-based outdoor ventilation air dehumidification can be achieved at a constant cooling-coil temperature. Reheat of the ventilation air may be required to avoid overcooling the supply air if ventilation demands are high. The resulting system provides more stable dehumidification and improved relative humidity control. Alternatives for residential construction include local stand-alone dehumidifiers, with or without central system mixing; continuous exhaust/supply indoor air dehumidification using either a central
2017 ASHRAE Handbook—Fundamentals (SI) HVAC system with subcooling followed by reheat, or ducted dehumidifiers; or ducted direct-expansion (DX) condenser-regenerated desiccant dehumidifiers (Harriman and Lstiburek 2009; Rudd 2013). Pressurization. Dehumidification reduces the indoor vapor pressure below the outdoor vapor pressure level. Cooling does the same for temperature. As a result, the vapor pressure and temperature gradients cause moisture migration by diffusion toward the interior. If the envelope is not airtight, building depressurization and related infiltration increase this moisture flow through the exterior wall and connected partition wall cavities. This may raise the relative humidity at or near surfaces in these cavities to levels that allow mold, or even condensation. Pressurization reverses the airflow, allowing dehumidified indoor air to exfiltrate and neutralizing inward vapor diffusion. In any case, the envelope should be constructed as airtight as possible so that pressurization does not require unnecessarily high supply ventilation.
Cold Climates In cold climates, the techniques for avoiding mold and surface condensation are the same as in temperate climates. During cold seasons, the indoor relative humidity can drop well below 30%. Maintaining healthy levels of relative humidity then requires humidification, which results in an increased indoor/outdoor vapor pressure difference for a temperature gradient pointing to the outdoors and vapor diffusion toward the exterior. Humidification. At part load, on/off air-handling units that heat and humidify a mixture of outdoor and return air operate so intermittently that indoor relative humidity levels may drop substantially during off periods. Moisture buffering by the building fabric, furniture, and furnishings can reduce that variation in relative humidity (Simonson et al. 2004a, 2004b). A better choice in larger buildings is a DOAS that restricts humidification to the ventilation air, handling heating loads by a separate system. Because the ventilation required mainly depends on the number of occupants, the space can be humidified by using steam injection modulated according to demand, or by preheating the ventilation air followed by adiabatic humidification and reheat to the indoor temperature set point. The result is more stable humidification and better relative humidity control. Alternatives in homes are stand-alone humidifiers, continuous-supply air humidification at constant dew-point temperature by the central HVAC system, etc. Pressurization/Depressurization. Pressurizing buildings in cold climates leads to air exfiltration, which provides an extra vehicle for vapor egress. Together with diffusion, this could lead to severe interstitial condensation in envelope assemblies, especially when the indoor vapor retarder is ineffective and the air barrier is not continuous. Depressurizing a building, which leads to infiltration, is as effective way to avoid interstitial condensation problems, as long as the envelope is sufficiently airtight that depressurization does not require ventilation rates beyond those needed by the occupant load. If the envelope is too air permeable, not only will large exhaust rates be required but all drawbacks mentioned in the section on Temperate and Mixed Climates will apply but be more severe.
9.
MOISTURE MANAGEMENT IN OTHER HANDBOOK CHAPTERS
Other chapters and volumes of the ASHRAE Handbook also provide useful information on moisture and humidity. In this volume, Ch. 1
Perfect gas relations and thermodynamic properties to analyze conditions and processes involving moist air 4 Heat transfer processes that affect temperature and moisture transport in buildings
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9
Effects of relative humidity, temperature, and vapor pressure on thermal comfort 10 Effects of relative humidity on indoor environmental quality (IEQ), mold, and humidifier fever 11 Controlling mold through control of relative humidity at material surfaces 12 Temperature and relative humidity’s effects on olfactory perception and perceived indoor air quality 13 Intra- and interzone airflows and pollutant transfer, including water vapor 14 Data on outdoor dry- and wet-bulb temperatures for design purposes 15 Relation between relative humidity and condensation potential on glazing 16 Control of interstitial condensation by controlling air leakage across the envelope, and moisture control by modulating air infiltration across the envelope 17 Formulas to calculate latent load for air humidification (often required to maintain comfortable relative humidity indoors) 18 Latent heat gain from moisture diffusion and various moisture sources, including natatoriums 19 Cooling and dehumidification coils, and cooling towers 23 Condensation control for below-ambient-temperature HVAC components 25/26/27 Detailed information on heat, air, and moisture response of building envelope assemblies; includes fundamentals of combined heat, air, and moisture movement (e.g., moisture content, buffering, flow), material properties (e.g., sorption isotherms, vapor permeability), and examples 2014 ASHRAE Handbook—Refrigeration Ch. 7 Role of moisture as a contaminant in refrigeration systems 10 How moisture diffusion through low-water-vapor-permeance insulation and water-permeable claddings can produce damaging condensation and moisture accumulation 23 Moisture aspects of refrigerated-facility envelope design 24 Sensible and latent heat loads in refrigerated facilities 44 Specific ceiling moisture problems in ice rinks as a consequence of long-wave radiation between ice surface and ceiling 2015 ASHRAE Handbook—Applications Ch. 1 Residential humidification and dehumidification 2 Humidity control in stores 3 Thermal comfort design criteria and load characteristics in commercial and public buildings 4 Stack effect and practical considerations for minimizing stack effects 5 Latent load in arenas, humidity problems in ice rinks, vapor release in natatoriums with some consequences for envelope design 6 Dehumidification as a load characteristic, considers moisture control using DOAS 7 Humidity control in classrooms, libraries, gymnasiums, showers, and natatoriums 8 Role of relative humidity in health care for different kinds of patients 9 Dry- and wet-bulb design criteria for general and specialized spaces in courthouses and detention/correction facilities 16 Relative humidity control in laboratories 18 Relative humidity control in clean spaces 19 Drawbacks for data processing and telecommunication facilities of high (anodic failure) and low relative humidity (electrostatic discharge); envelope considerations and humidity control
36.13 20 Paper moisture content control and controlling relative humidity in pressrooms for various printing processes 21 Relative humidity and textiles 22 Relative humidity and photographic materials 23 Importance and impact of relative humidity mean values and fluctuations on museum artifacts, library books, and other archival materials 24 Role of relative humidity in environmental control in animal and plant facilities 25 Role of relative humidity in crop drying 26 Moisture control in processing wood and paper 30 Industrial drying as a process 31 Ventilation for industrial processes 33 Ventilation to reduce relative humidity in kitchens (commercial and residential) 44 Control of liquid water and vapor in building envelopes 47 Design of system controls for humidity 52 Humidification and dehumidification in evaporative cooling 62 Practical applications of moisture management in buildings 2016 ASHRAE Handbook—Systems and Equipment Ch. 4 Humidity control and air-handling unit psychrometric processes 11 Condensate removal from temperature-regulated systems and steam traps 12 Condensate drainage and return in district systems 22 Optimum humidity range for human comfort and health; surface and interstitial condensation; some data on indoor vapor release 23 Dehumidification coils 24 Methods of dehumidification, with emphasis on sorption dehumidification 25 Mechanical dehumidification 26 Air-to-air energy recovery processes, from basic thermodynamics to types and applications of air-to-air heat exchangers and technical considerations; includes condensation and freeze-up in air-to-air heat exchangers, and gives an energy and moisture recovery procedure 40 Evaporation-based cooling towers 41 Evaporative humidification and dehumidification
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org /bookstore. Antretter, F., A. Holm, A. Karagiozis, and S. Glass. 2010. Interior temperature and relative humidity distributions in mixed-humid and cold climates as building simulation boundary conditions. Proceedings of the Buildings XI Conference, Clearwater Beach. Antretter, F., C. Mitterer, and S.-M. Young. 2012. Use of moisture-buffering tiles for indoor climate stability under different climatic requirements. HVAC&R Research (now Science and Technology for the Built Environment) 18(1-2):275-282. Arena, L., A. Karagiozis, and P. Mantha. 2010. Monitoring of internal moisture loads in residential buildings-research findings in three different climate zones. Proceedings of the Buildings XI Conference, Clearwater Beach. ASHRAE. 2009. Criteria for moisture-control design analysis in buildings. ANSI/ASHRAE Standard 160-2009. ASHRAE. 2012. Position document on limiting indoor mold and dampness in buildings. Derome, D., and S. Saneinejad. 2009. The nature, significance and control of solar-driven diffusion in wall systems. ASHRAE Research Project RP1235, Final Report.
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Francisco, P., and W. Rose. 2010. Temperature and humidity measurements in 71 homes. Proceedings of the Buildings XI Conference, Clearwater Beach. Francisco, P., J. Gordon, and W. Rose. 2009. Indoor moisture in 30 homes using unvented gas fireplaces. ASHRAE Transactions 115(2). Hartmann, T., D. Reichel, and W. Richter. 2001. Feuchteabgabe in Wohnungen—Alles gesagt? Gesundheits-Ingenieur 122(4):189-195. In German. Harriman, III, L., and J.W. Lstiburek. 2009. The ASHRAE guide for buildings in hot & humid climates. ASHRAE. Hens, H. 1999. Fungal defacement in buildings, a performance related approach. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 5(3):265-289. Hens, H. 2005. Energy consumption in student residences, impact of building performance, usage patterns and indoor climate. Proceedings Clima 2005, Lausanne. Hens, H. 2009a. IEA-ECBCS Annex 41 whole building heat, air and moisture response. ASHRAE Transactions 115(2). Hens, H. 2009b. Indoor climate and building envelope performance in indoor swimming pools. In Energy efficiency and new approaches, N. Bayazit, G. Manioglu, G. Oral, and Z. Yilmaz, eds. Istanbul Technical University. Hens, H. 2012a. Building physics, heat, air and moisture: Fundamentals and engineering methods with examples and exercises, 2nd ed. Ernst & Sohn, Berlin. Hens, H. 2012b. Passive houses: what may happen when energy efficiency becomes the only paradigm. ASHRAE Transactions 118(1). Hens, H. 2016. Applied building physics, ambient conditions, building performance and material properties, 2nd ed. Ernst & Sohn, Berlin. Hens, H., and V. De Meulenaer. 2007. Schools, all problem buildings? Proceedings Clima 2007, Helsinki. Holm, A. 2008. Whole building heat, air, moisture response: Applications: Indoor environment, energy, durability. Final Report, vol. 4. International Energy Agency Energy Conservation in Buildings and Community Systems (ECBCS), Annex 41: Whole Building Heat, Air and Moisture Response. www.ecbcs.org/annexes/annex41.htm. IEA-ECB. 1990. Sourcebook. International Energy Agency Energy and Community in Buildings Programme, Annex 14: Condensation and Energy, Leuven. IEA-ECB. 1990. Guidelines and practice. International Energy Agency Energy and Community in Buildings Programme, Annex 14: Condensation and Energy, Leuven. Kalamees, T., J. Vinha, and J. Kurnitski. 2006. Indoor humidity loads in light-weight timber-framed detached houses. Journal of Building Physics 29(3):219-246. jen.sagepub.com/content/29/3/219.full.pdf+html Kumaran, K., and C. Sanders. 2008. Whole building heat, air, moisture response: Boundary conditions and whole building HAM analysis. Final Report. International Energy Agency Energy Conservation in Buildings and Community Systems (ECBCS), Annex 41: Whole Building Heat, Air and Moisture Response. www.ecbcs.org/annexes/annex41.htm. Mumovic, D., and M. Santamouris, eds. 2009. A handbook on sustainable building design and engineering: An integrated approach to energy, health and operational performance. Earthscan, London. Ridley, I., M. Davies, S.H. Hong, and T. Oreszcyn. 2008. Vapour pressure excess in living rooms and bedrooms of English dwellings: Analysis of the warm front dataset. Appendix III of Boundary Conditions and Whole Building HAM Analysis. Final Report, International Energy Agency Energy and Community in Buildings Programme, Annex 41: Whole Building Heat, Air and Moisture Response (MOIST-EN), Leuven. Rudd, A.F. 2010. Enhanced dehumidification. Building America Report 1008. U.S. Department of Energy, Washington, D.C. buildingscience.com /documents/bareports/ba-1008-building-america-enhanced-dehumi dification/view.
Rudd, A.F. 2013. Supplemental dehumidification in warm-humid climates. Building America Report 1310. U.S. Department of Energy, Washington, D.C. buildingscience.com/documents/bareports/ba-1310-supplemental -dehumidification-warm-humid-climates/view. Rudd, A.F., J.W. Lstiburek, and K. Ueno. 2005. Residential dehumidification systems research for hot-humid climates, September 1, 2001-December 30, 2003. Report NREL/SR-550-36643. Sanders, C. 1996, Environmental conditions. Final Report, vol. 2, International Energy Agency Energy and Community in Buildings Programme, Annex 24: Heat, Air and Moisture Transport in Insulated Envelope Parts, Leuven. Sedlbauer, K. 2001. Vorhersage von Schimmelpilzbildung auf und in Bauteilen (Prediction of mould development on and in building assemblies). Ph.D. dissertation, University of Stuttgart (in German). Simonson, C., M. Salonvaara, and T. Ojanen. 2004a. Heat and mass transfer between indoor air and a permeable and hygroscopic building envelope: Part 1—Field measurements. Journal of Thermal Envelope and Building Science 28(1):63-101. Simonson, C., M. Salonvaara, and T. Ojanen. 2004b. Heat and mass transfer between indoor air and a permeable and hygroscopic building envelope: Part 2—Verification and numerical studies. Journal of Thermal Envelope and Building Science 28(2):161-185. TenWolde, A., and I. Walker. 2001. Interior moisture design loads for residences. Proceedings of the Buildings VIII Conference, Clearwater Beach. TenWolde, A., and C. Pilon. 2007. The effect of indoor humidity on water vapor release in homes. Proceedings of the Buildings X Conference, Clearwater Beach. Trechsel, H., ed. 1994. Moisture control in buildings. ASTM Manual Series MNL 18. American Society for Testing and Materials, West Conshohocken, PA. Vereecken, E., and S. Roels. 2012. Review of mould prediction models and their influence on mould risk evaluation. Building and Environment 52:296-310. Viitanen, J., and T. Ojanen. 2007. Improved model to predict mold growth in building materials. Proceedings of the Buildings X Conference, Clearwater Beach. Woloszyn, M., and C. Rode. 2008. Whole building heat, air, moisture response: Modeling principles and common exercises. Final Report, International Energy Agency Energy and Community in Buildings Programme, Annex 41: Whole Building Heat, Air and Moisture Response (MOIST-EN), Leuven.
BIBLIOGRAPHY BRE. 2016. BREEAM International New Construction. Building Research Establishment, Watford, U.K. Harriman, III, L., G. Brundrett, and R. Kittler. 2001. Humidity control design guide for commercial and institutional buildings. ASHRAE. Künzel, H.M. 2006. Indoor relative humidity in residential buildings—A necessary boundary condition to assess the moisture performance of building envelope systems. Zeitschrift für Wärmeschutz, Kälteschutz, Schallschutz, Brandschutz 51(56):31-41. Künzel, H.M., D. Zirkelbach, and K. Sedlbauer. 2003. Predicting indoor temperature and humidity conditions including hygrothermal interactions with the building envelope. Proceedings of 1st International Conference on Sustainable Energy and Green Architecture, Building Scientific Research Center (BSRC), Bangkok. USGBC. 2009. LEED 2009 for new construction and major renovation. U.S. Green Building Council, Washington, D.C.
Related Commercial Resources
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MEASUREMENT AND INSTRUMENTS Terminology ............................................................................. 37.1 Uncertainty Analysis................................................................ 37.3 Temperature Measurement ...................................................... 37.4 Humidity Measurement.......................................................... 37.10 Pressure Measurement........................................................... 37.13 Air Velocity Measurement...................................................... 37.14 Flow Rate Measurement ........................................................ 37.20 Air Infiltration, Airtightness, and Outdoor Air Ventilation Rate Measurement........................................... 37.24 Carbon Dioxide Measurement............................................... 37.25 Electric Measurement ............................................................ 37.27
Rotative Speed and Position Measurement............................ Sound and Vibration Measurement........................................ Lighting Measurement ........................................................... Thermal Comfort Measurement ............................................. Moisture Content and Transfer Measurement ...................................................................... Heat Transfer Through Building Materials ........................... Air Contaminant Measurement .............................................. Combustion Analysis.............................................................. Data Acquisition and Recording............................................ Mechanical Power Measurement...........................................
VAC engineers and technicians require instruments for both laboratory work and fieldwork. Precision is more essential in the laboratory, where research and development are undertaken, than in the field, where acceptance and adjustment tests are conducted. This chapter describes the characteristics and uses of some of these instruments.
Diameter, equivalent. The diameter of a circle having the same area as the rectangular flow channel cross section. Deviation, standard. Square root of the average of the squares of the deviations from the mean (root mean square deviation). A measure of dispersion of a population. Distortion. Unwanted change in wave form. Principal forms of distortion are inherent nonlinearity of the device, nonuniform response at different frequencies, and lack of constant proportionality between phase-shift and frequency. (A wanted or intentional change might be identical, but it is called modulation.) Drift. Gradual, undesired change in output over a period of time that is unrelated to input, environment, or load. Drift is gradual; if variation is rapid and recurrent, with elements of both increasing and decreasing output, the fluctuation is referred to as cycling. Dynamic error band. Spread or band of output-amplitude deviation incurred by a constant-amplitude sine wave as its frequency is varied over a specified portion of the frequency spectrum (see Static error band). Emissivity. Ratio of the amount of radiation emitted by a real surface to that of an ideal (blackbody) emitter at the same temperature. Error. Difference between the true or actual value to be measured (input signal) and the indicated value (output) from the measuring system. Errors can be systematic or random. Error, accuracy. See Error, systematic. Error, fixed. See Error, systematic. Error, instrument. Error of an instrument’s measured value that includes random or systematic errors. Error, precision. See Error, random. Error, probable. Error with a 50% or higher chance of occurrence. A statement of probable error is of little value. Error, random. Statistical error caused by chance and not recurring. This term is a general category for errors that can take values on either side of an average value. To describe a random error, its distribution must be known. Error, root mean square (RMS). Accuracy statement of a system comprising several items. For example, a laboratory potentiometer, volt box, null detector, and reference voltage source have individual accuracy statements assigned to them. These errors are generally independent of one another, so a system of these units displays an accuracy given by the square root of the sum of the squares of the individual limits of error. For example, four individual errors of 0.1% could yield a calibrated error of 0.4% but an RMS error of only 0.2%. Error, systematic. Persistent error not caused by chance; systematic errors are causal. It is likely to have the same magnitude and sign for every instrument constructed with the same components and
H
1.
TERMINOLOGY
The following definitions are generally accepted. Accuracy. Ability of an instrument to indicate the true value of measured quantity. This is often confused with inaccuracy, which is the departure from the true value to which all causes of error (e.g., hysteresis, nonlinearity, drift, temperature effect) contribute. Amplitude. Magnitude of variation from its equilibrium or average value in an alternating quantity. Average. Sum of a number of values divided by the number of values. Bandwidth. Range of frequencies over which a given device is designed to operate within specified limits. Bias. Tendency of an estimate to deviate in one direction from a true value (a systematic error). Calibration. (1) Process of comparing a set of discrete magnitudes or the characteristic curve of a continuously varying magnitude with another set or curve previously established as a standard. Deviation between indicated values and their corresponding standard values constitutes the correction (or calibration curve) for inferring true magnitude from indicated magnitude thereafter; (2) process of adjusting an instrument to fix, reduce, or eliminate the deviation defined in (1). Calibration reduces bias (systematic) errors. Calibration curve. (1) Path or locus of a point that moves so that its graphed coordinates correspond to values of input signals and output deflections; (2) plot of error versus input (or output). Confidence. Degree to which a statement (measurement) is believed to be true. Dead band. Range of values of the measured variable to which an instrument will not effectively respond. The effect of dead band is similar to hysteresis, as shown in Figure 1. Deviate. Any item of a statistical distribution that differs from the selected measure of control tendency (average, median, mode). Deviation. Difference between a single measured value and the mean (average) value of a population or sample. The preparation of this chapter is assigned to TC 1.2, Instruments and Measurements.
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37.28 37.29 37.31 37.31 37.32 37.34 37.35 37.35 37.35 37.37
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37.2
Fig. 1 Measurement and Instrument Terminology procedures. Errors in calibrating equipment cause systematic errors because all instruments calibrated are biased in the direction of the calibrating equipment error. Voltage and resistance drifts over time are generally in one direction and are classed as systematic errors. Frequency response (flat). Portion of the frequency spectrum over which the measuring system has a constant value of amplitude response and a constant value of time lag. Input signals that have frequency components within this range are indicated by the measuring system (without distortion). Hydraulic diameter Dh. Defined as 4Ac /Pwet , where Ac is flow cross-sectional area and Pwet is the wetted perimeter (perimeter in contact with the flowing fluid). For a rectangular duct with dimensions W × H, the hydraulic diameter is Dh = 2HW/(H + W ). The related quantity effective or equivalent diameter is defined as the diameter of a circular tube having the same cross-sectional area as the actual flow channel. For a rectangular flow channel, the effective diameter is Deff = 4HW . Hysteresis. Summation of all effects, under constant environmental conditions, that cause an instrument’s output to assume
different values at a given stimulus point when that point is approached with increasing or decreasing stimulus. Hysteresis includes backlash. It is usually measured as a percent of full scale when input varies over the full increasing and decreasing range. In instrumentation, hysteresis and dead band exhibit similar output error behavior in relation to input, as shown in Figure 1. Linearity. The degree of straightness of the transfer curve between an input and an output (e.g., the ideal line in Figure 1); that condition prevailing when output is directly proportional to input (see Nonlinearity). Note that the generic term linearity does not consider any parallel offset of the straight-line calibration curve. Loading error. Loss of output signal from a device caused by a current drawn from its output. It increases the voltage drop across the internal impedance, where no voltage drop is desired. Mean. See Average. Median. Middle value in a distribution, above and below which lie an equal number of values. Mode. Value in a distribution that occurs most frequently. Noise. Any unwanted disturbance or spurious signal that modifies the transmission, measurement, or recording of desired data. Nonlinearity. Prevailing condition (and the extent of its measurement) under which the input/output relationship (known as the input/output curve, transfer characteristic, calibration curve, or response curve) fails to be a straight line. Nonlinearity is measured and reported in several ways, and the way, along with the magnitude, must be stated in any specification. Minimum-deviation-based nonlinearity: maximum departure between the calibration curve and a straight line drawn to give the greatest accuracy; expressed as a percent of full-scale deflection. Slope-based nonlinearity: ratio of maximum slope error anywhere on the calibration curve to the slope of the nominal sensitivity line; usually expressed as a percent of nominal slope. Most other variations result from the many ways in which the straight line can be arbitrarily drawn. All are valid as long as construction of the straight line is explicit. Population. Group of individual persons, objects, or items from which samples may be taken for statistical measurement. Precision. Repeatability of measurements of the same quantity under the same conditions; not a measure of absolute accuracy. It describes the relative tightness of the distribution of measurements of a quantity about their mean value. Therefore, precision of a measurement is associated more with its repeatability than its accuracy. It combines uncertainty caused by random differences in a number of identical measurements and the smallest readable increment of the scale or chart. Precision is given in terms of deviation from a mean value. Primary calibration. Calibration procedure in which the instrument output is observed and recorded while the input stimulus is applied under precise conditions, usually from a primary external standard traceable directly to the National Institute of Standards and Technology (NIST) or to an equivalent international standards organization. Range. Statement of upper and lower limits between which an instrument’s input can be received and for which the instrument is calibrated. Reliability. Probability that an instrument’s precision and accuracy will continue to fall within specified limits. Repeatability. See Precision. Reproducibility. In instrumentation, the closeness of agreement among repeated measurements of the output for the same value of input made under the same operating conditions over a period of time, approaching from both directions; it is usually measured as a nonreproducibility and expressed as reproducibility in percent of span for a specified time period. Normally, this implies a long period of time, but under certain conditions, the period may be a short time so that drift is not included. Reproducibility includes
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Measurement and Instruments hysteresis, dead band, drift, and repeatability. Between repeated measurements, the input may vary over the range, and operating conditions may vary within normal limits. Resolution. Smallest change in input that produces a detectable change in instrument output. Resolution, unlike precision, is a psychophysical term referring to the smallest increment of humanly perceptible output (rated in terms of the corresponding increment of input). The precision, resolution, or both may be better than the accuracy. An ordinary six-digit instrument has a resolution of one part per million (ppm) of full scale; however, it is possible that the accuracy is no better than 25 ppm (0.0025%). Note that the practical resolution of an instrument cannot be any better than the resolution of the indicator or detector, whether internal or external. Sensitivity. Slope of a calibration curve relating input signal to output, as shown in Figure 1. For linear instruments, sensitivity represents the change in output for a unit change in the input. Sensitivity error. Maximum error in sensitivity displayed as a result of the changes in the calibration curve resulting from accumulated effects of systematic and random errors. Stability. (1) Independence or freedom from changes in one quantity as the result of a change in another; (2) absence of drift. Static error band. (1) Spread of error present if the indicator (pen, needle) stopped at some value (e.g., at one-half of full scale), normally reported as a percent of full scale; (2) specification or rating of maximum departure from the point where the indicator must be when an on-scale signal is stopped and held at a given signal level. This definition stipulates that the stopped position can be approached from either direction in following any random waveform. Therefore, it is a quantity that includes hysteresis and nonlinearity but excludes items such as chart paper accuracy or electrical drift (see Dynamic error band). Step-function response. Characteristic curve or output plotted against time resulting from the input application of a step function (a function that is zero for all values of time before a certain instant, and a constant for all values of time thereafter). Threshold. Smallest stimulus or signal that results in a detectable output. Resolution and threshold are sometimes used interchangeably. Time constant. Time required for an exponential quantity to change by an amount equal to 0.632 times the total change required to reach steady state for first-order systems. Transducer. Device for translating the changing magnitude of one kind of quantity into corresponding changes of another kind of quantity. The second quantity often has dimensions different from the first and serves as the source of a useful signal. The first quantity may be considered an input and the second an output. Significant energy may or may not transfer from the transducer’s input to output. Uncertainty. An estimated value for the bound on the error (i.e., what an error might be if it were measured by calibration). Although uncertainty may be the result of both systematic and precision errors, only precision error can be treated by statistical methods. Uncertainty may be either absolute (expressed in the units of the measured variable) or relative (absolute uncertainty divided by the measured value; commonly expressed in percent). Zero shift. Drift in the zero indication of an instrument without any static change in the measured variable.
2.
37.3 • Inaccuracy in the mathematical model that describes the physical quantity • Inherent stochastic variability of the measurement process • Uncertainties in measurement standards and calibrated instrumentation • Time-dependent instabilities caused by gradual changes in standards and instrumentation • Effects of environmental factors such as temperature, humidity, and pressure • Values of constants and other parameters obtained from outside sources • Uncertainties arising from interferences, impurities, inhomogeneity, inadequate resolution, and incomplete discrimination • Computational uncertainties and data analysis • Incorrect specifications and procedural errors • Laboratory practice, including handling techniques, cleanliness, and operator techniques, etc. • Uncertainty in corrections made for known effects, such as installation effect corrections
Uncertainty of a Measured Variable For a measured variable X, the total error is caused by both precision (random) and systematic (bias) errors. This relationship is shown in Figure 2. The possible measurement values of the variable are scattered in a distribution around the parent population mean (Figure 2A). The curve (normal or Gaussian distribution) is the theoretical distribution function for the infinite population of measurements that generated X. The parent population mean differs from (X)true by an amount called the systematic (or bias) error (Figure 2B). The quantity is the total fixed error that remains after all calibration corrections have been made. In general, there are several sources of bias error, such as errors in calibration standard, data acquisition, data reduction, and test technique. There is usually no direct way to measure these errors. These errors are unknown and are assumed to be zero; otherwise, an additional correction would
UNCERTAINTY ANALYSIS
Uncertainty Sources Measurement generally consists of a sequence of operations or steps. Virtually every step introduces a conceivable source of uncertainty, the effect of which must be assessed. The following list is representative of the most common, but not all, sources of uncertainty.
Fig. 2
Errors in Measurement of Variable X
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be applied to reduce them to as close to zero as possible. Figure 2B shows how the resulting deviation can be different for different random errors . The precision uncertainty for a variable, which is an estimate of the possible error associated with the repeatability of a particular measurement, is determined from the sample standard deviation, or the estimate of the error associated with the repeatability of a particular measurement. Unlike systematic error, precision error varies from reading to reading. As the number of readings of a particular variable tends to infinity, the distribution of these possible errors becomes Gaussian. For each bias error source, the experimenter must estimate a systematic uncertainty. Systematic uncertainties are usually estimated from previous experience, calibration data, analytical models, and engineering judgment. The resultant uncertainty is the square root of the sum of the squares of the bias and precision uncertainties; see Coleman and Steele (2009). For further information on measurement uncertainty, see Abernethy et al. (1985), ASME Standards MFC-2M and PTC 19.1, Brown et al. (1998), and Coleman and Steele (1995).
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3.
TEMPERATURE MEASUREMENT
Instruments for measuring temperature are listed in Table 1. Temperature sensor output must be related to an accepted temperature
scale by manufacturing the instrument according to certain specifications or by calibrating it against a temperature standard. To help users conform to standard temperatures and temperature measurements, the International Committee of Weights and Measures (CIPM) adopted the International Temperature Scale of 1990 (ITS-90). The unit of temperature of the ITS-90 is the kelvin (K) and has a size equal to the fraction 1/273.16 of the thermodynamic temperature of the triple point of water. In the United States, ITS-90 is maintained by the National Institute of Standards and Technology (NIST), which provides calibrations based on this scale for laboratories. Benedict (1984), Considine (1985), DeWitt and Nutter (1988), Holman (2001), Quinn (1990), and Schooley (1986, 1992) cover temperature measurement in more detail.
Sampling and Averaging Although temperature is usually measured within, and is associated with, a relatively small volume (depending on the size of the thermometer), it can also be associated with an area (e.g., on a surface or in a flowing stream). To determine average stream temperature, the cross section must be divided into smaller areas and the temperature of each area measured. The temperatures measured are then combined into a weighted mass flow average by using either (1) equal areas and multiplying each temperature by
Table 1 Common Temperature Measurement Techniques Approximate Range, °C
Uncertainty, K Limitations
Measurement Means
Application
Liquid-in-glass thermometers Mercury-in-glass
Temperature of gases and liquids by contact
–38/550
0.03 to 2
Temperature of gases and liquids by contact
–200/200
0.03 to 2
Precision; remote readings; temperature of fluids or solids by contact Transfer standard for cryogenic applications Remote readings; temperature by contact Remote readings; temperature by contact Remote readings; temperature by contact
–259/1000
Standard for thermocouples on IPTS-68, not on ITS-90 Highly accurate reference thermometer for laboratory applications General testing of high temperature; remote rapid readings by direct contact Same as above Same as above; especially suited for low temperature Same as above; especially suited for low temperature For approximate temperature
0/1450
0.1 to 3
High cost
–50/1000
0.05 to 1
High cost
–200/1250
0.1 to 10
0 to 760 –200/370
0.1 to 6 0.1 to 3
Less accurate than Pt-Rh/Pt or Au/Pt thermocouples Subject to oxidation
–200/900
0.1 to 7
–20/660
1, usually much more
–75/660 –5/250 –50/1150 800 and up
2 2 2 15
Organic fluid Resistance thermometers Platinum Rhodium/iron Nickel Germanium Thermistors Thermocouples Pt-Rh/Pt (type S) Au/Pt Types K and N Iron/Constantan (type J) Copper/Constantan (type T) Ni-Cr/Constantan (type E) Bimetallic thermometers Pressure-bulb thermometers Gas-filled bulb Vapor-filled bulb Liquid-filled bulb Optical pyrometers Infrared (IR) radiometers IR thermography
Remote reading Remote testing Remote testing For intensity of narrow spectral band of hightemperature radiation (remote) For intensity of total high-temperature radiation (remote) Infrared imaging
Seger cones (fusion pyrometers) Approximate temperature (within temperature source)
–273/–243 –250/200 –273/–243 –90/200
In gases, accuracy affected by radiation (unless adequately shielded) In gases, accuracy affected by radiation
Less than High cost; accuracy affected by radiation in 0.0001 to 0.1 gases 0.0001 to 0.1 High cost 0.01 to 1 Accuracy affected by radiation in gases 0.0001 to 0.1 0.0001 to 0.1
Time lag; unsuitable for remote use
Use caution to ensure installation is correct Use caution to ensure installation is correct Use caution to ensure installation is correct Generally requires knowledge of surface emissivity
Any range Any range 660/2000
Generally requires knowledge of surface emissivity 50
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37.5
the fraction of total mass flow in its area or (2) areas of size inversely proportional to mass flow and taking a simple arithmetic average of the temperatures in each. Mixing or selective sampling may be preferable to these cumbersome procedures. Although mixing can occur from turbulence alone, transposition is much more effective. In transposition, the stream is divided into parts determined by the type of stratification, and alternate parts pass through one another.
necessary to determine the proper corrections to be applied to the scale readings. For application and calibration of liquid-in-glass thermometers, refer to NIST (1976, 1986). Liquid-in-glass thermometers are calibrated by the manufacturer for total or partial stem immersion. If a thermometer calibrated for total immersion is used at partial immersion (i.e., with part of the liquid column at a temperature different from that of the bath), an emergent stem correction must be made, as follows: Stem correction = Kn(tb – ts)
Static Temperature Versus Total Temperature When a fluid stream impinges on a temperature-sensing element such as a thermometer or thermocouple, the element is at a temperature greater than the true stream temperature. The difference is a fraction of the temperature equivalent of the stream velocity te . V2 te = ----------2Jc p
(1)
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where te V J cp
= = = =
temperature equivalent of stream velocity, °C stream velocity, m/s mechanical equivalent of heat = 1000 (N·m)/kJ specific heat of stream at constant pressure, kJ/(kg·K)
This fraction of the temperature equivalent of the velocity is the recovery factor, which varies from 0.3 to 0.4 K for bare thermometers to 0.5 K for aerodynamically shielded thermocouples. For precise temperature measurement, each temperature sensor must be calibrated to determine its recovery factor. However, for most applications with air velocities below 10 m/s (or Mach number M below approximately 0.1), the recovery factor can be omitted. Various sensors are available for temperature measurement in fluid streams. The principal ones are the static temperature thermometer, which indicates true stream temperature but is cumbersome, and the thermistor, used for accurate temperature measurement within a limited range.
3.1
LIQUID-IN-GLASS THERMOMETERS
Any device that changes monotonically with temperature is a thermometer; however, the term usually signifies an ordinary liquidin-glass temperature-indicating device. Mercury-filled thermometers have a useful range from –38.8°C, the freezing point of mercury, to about 550°C, near which the glass usually softens. Lower temperatures can be measured with organic-liquid-filled thermometers (e.g., alcohol-filled), with ranges of –200 to 200°C. During manufacture, thermometers are roughly calibrated for at least two temperatures, often the freezing and boiling points of water; space between the calibration points is divided into desired scale divisions. Thermometers that are intended for precise measurement applications have scales etched into the glass that forms their stems. The probable error for as-manufactured, etched-stem thermometers is ±1 scale division. The highest-quality mercury thermometers may have uncertainties of ±0.03 to 2 K if they have been calibrated by comparison against primary reference standards. Liquid-in-glass thermometers are used for many HVAC applications, including local temperature indication of process fluids (e.g., cooling and heating fluids and air). Mercury-in-glass thermometers are fairly common as temperature measurement standards because of their relatively high accuracy and low cost. If used as references, they must be calibrated on the ITS-90 by comparison in a uniform bath with a standard platinum resistance thermometer that has been calibrated either by the appropriate standards agency or by a laboratory that has direct traceability to the standards agency and the ITS-90. This calibration is
(2)
where K = differential expansion coefficient of mercury or other liquid in glass. K is 0.00016 for Celsius mercury thermometers. For K values for other liquids and specific glasses, refer to Schooley (1992). n = number of degrees that liquid column emerges from bath tb = temperature of bath, °C ts = average temperature of emergent liquid column of n degrees, °C
Because the true temperature of the bath is not known, this stem correction is only approximate.
Sources of Thermometer Errors A thermometer measuring gas temperatures can be affected by radiation from surrounding surfaces. If the gas temperature is approximately the same as that of the surrounding surfaces, radiation effects can be ignored. If the temperature differs considerably from that of the surroundings, radiation effects should be minimized by shielding or aspiration (ASME Standard PTC 19.3). Shielding may be provided by highly reflective surfaces placed between the thermometer bulb and the surrounding surfaces such that air movement around the bulb is not appreciably restricted (Parmelee and Huebscher 1946). Improper shielding can increase errors. Aspiration involves passing a high-velocity stream of air or gas over the thermometer bulb. When a thermometer well within a container or pipe under pressure is required, the thermometer should fit snugly and be surrounded with a high-thermal-conductivity material (oil, water, or mercury, if suitable). Liquid in a long, thin-walled well is advantageous for rapid response to temperature changes. The surface of the pipe or container around the well should be insulated to eliminate heat transfer to or from the well. Industrial thermometers are available for permanent installation in pipes or ducts. These instruments are fitted with metal guards to prevent breakage. However, the considerable heat capacity and conductance of the guards or shields can cause errors. Allowing ample time for the thermometer to attain temperature equilibrium with the surrounding fluid prevents excessive errors in temperature measurements. When reading a liquid-in-glass thermometer, keep the eye at the same level as the top of the liquid column to avoid parallax.
3.2
RESISTANCE THERMOMETERS
Resistance thermometers depend on a change of the electrical resistance of a sensing element (usually metal) with a change in temperature; resistance increases with increasing temperature. Use of resistance thermometers largely parallels that of thermocouples, although readings are usually unstable above about 550°C. Twolead temperature elements are not recommended because they do not allow correction for lead resistance. Three leads to each resistor are necessary for consistent readings, and four leads are preferred. Wheatstone bridge circuits or 6-1/2-digit multimeters can be used for measurements. A typical circuit used by several manufacturers is shown in Figure 3. This design uses a differential galvanometer in which coils L and H exert opposing forces on the indicating needle. Coil L is in
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2017 ASHRAE Handbook—Fundamentals (SI)
series with the thermometer resistance AB, and coil H is in series with the constant resistance R. As the temperature falls, the resistance of AB decreases, allowing more current to flow through coil L than through coil H. This increases the force exerted by coil L, pulling the needle down to a lower reading. Likewise, as the temperature rises, the resistance of AB increases, causing less current to flow through coil L than through coil H and forcing the indicating needle to a higher reading. Rheostat S must be adjusted occasionally to maintain constant current. The resistance thermometer is more costly to make and likely to have considerably longer response times than thermocouples. It gives best results when used to measure steady or slowly changing temperature.
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Resistance Temperature Devices Resistance temperature devices (RTDs) are typically constructed from platinum, rhodium/iron, nickel, nickel/iron, tungsten, or copper. These devices are further characterized by their simple circuit designs, high degree of linearity, good sensitivity, and excellent stability. The choice of materials for an RTD usually depends on the intended application; selection criteria include temperature range, corrosion protection, mechanical stability, and cost. Presently, for HVAC applications, RTDs constructed of platinum are the most widely used. Platinum is extremely stable and resistant to corrosion. Platinum RTDs are highly malleable and can thus be drawn into fine wires; they can also be manufactured inexpensively as thin films. They have a high melting point and can be refined to high purity, thus attaining highly reproducible results. Because of these properties, platinum RTDs are used to define the ITS-90 for the range of 13.8033 K (triple point of equilibrium hydrogen) to 1234.93 K (freezing point of silver). Platinum resistance temperature devices can measure the widest range of temperatures and are the most accurate and stable temperature sensors. Their resistance/temperature relationship is one of the most linear. The higher the purity of the platinum, the more stable and accurate the sensor. With high-purity platinum, primarygrade platinum RTDs can achieve reproducibility of ±0.00001 K, whereas the minimum uncertainty of a recently calibrated thermocouple is ±0.2 K. The most widely used RTD is designed with a resistance of 100 at 0°C (R0 = 100 ). Other RTDs are available that use lower
resistances at temperatures above 600°C. The lower the resistance value, the faster the response time for sensors of the same size. Thin-Film RTDs. Thin-film 1000 platinum RTDs are readily available. They have the excellent linear properties of lowerresistance platinum RTDs and are more cost-effective because they are mass produced and have lower platinum purity. However, many platinum RTDs with R0 values of greater than 100 are difficult to provide with transmitters or electronic interface boards from sources other than the RTD manufacturer. In addition to a nonstandard interface, higher-R0-value platinum RTDs may have higher self-heating losses if the excitation current is not controlled properly. Thin-film RTDs have the advantages of lower cost and smaller sensor size. They are specifically adapted to surface mounting. Thin-film sensors tend to have an accuracy limitation of ±0.1% or ±0.1 K. This may be adequate for most HVAC applications; only in tightly controlled facilities may users wish to install the standard wire-wound platinum RTDs with accuracies of 0.01% or ±0.01 K (available on special request for certain temperature ranges). Assembly and Construction. Regardless of the R0 value, RTD assembly and construction are relatively simple. Electrical connections come in three basic types, depending on the number of wires to be connected to the resistance measurement circuitry. Two, three, or four wires are used for electrical connection using a Wheatstone bridge or a variation (Figure 4). In the basic two-wire configuration, the RTD’s resistance is measured through the two connecting wires. Because the connecting wires extend from the site of the temperature measurement, any additional changes in resistivity caused by a change in temperature may affect the measured resistance. Three- and four-wire assemblies are built to compensate for the connecting lead resistance values. The original three-wire circuit improved resistance measurement by adding a compensating wire to the voltage side of the circuit. This helps reduce part of the connecting wire resistance. When more accurate measurements (better than ±0.1 K) are required, the four-wire bridge, which eliminates all connecting wire resistance errors, is recommended. All bridges discussed here are direct current (DC) circuits and were used extensively until the advent of precision alternating current (AC) circuits using microprocessor-controlled ratio transformers, dedicated analog-to-digital converters, and other solid-state devices that measure resistance with uncertainties of less than 1 ppm. Resistance measurement technology now allows more portable thermometers, lower cost, ease of use, and high-precision temperature measurement in industrial uses.
Thermistors Some semiconductor compounds (usually sintered metallic oxides) exhibit large changes in resistance with temperature, usually decreasing as the temperature increases. For use, the thermistor element may be connected by lead wires into a galvanometer bridge circuit and calibrated. Alternatively, a 6-1/2-digit multimeter and a constant-current source with a means for reversing the current to eliminate thermal electromotive force (emf) effects may also be used. This method is easier and faster, and may be more precise and accurate. Thermistors are usually applied to electronic temperature compensation circuits, such as thermocouple reference junction compensation, or to other applications requiring high resolution and having limited operating-temperature ranges. Figure 5 shows a typical thermistor circuit.
Semiconductor Devices
Fig. 3 Typical Resistance Thermometer Circuit
In addition to positive-resistance-coefficient RTDs and negativeresistance-coefficient thermistors, there are two other types of devices that vary resistance or impedance with temperature. Although the principle of their operation has long been known, their reliability was questioned because of imprecise manufacturing techniques.
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Fig. 4 Typical Resistance Temperature Device (RTD) Bridge Circuits
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compensation to achieve an accuracy as high as ±0.01 K, but with a much higher cost. Lower-cost devices achieve accuracies of ±0.1 K using mass-manufacturing techniques and single-point calibration. A mass-produced silicon temperature sensor can be interchanged easily. If one device fails, only the sensor element need be changed. Electronic circuitry can be used to recalibrate the new device. Winding Temperature. The winding temperature of electrical operating equipment is usually determined from the resistance change of these windings in operation. The relationship between the winding resistance R1 at a known temperature t1 and the winding resistance at any other temperature t2 is given by Equation (3) (IEEE Standard 112-2004). 1 --- – t 0 + t 2 R2 k + t2 ------ = ------------------------------ = ------------R1 k + t1 1 --- – t 0 + t 1
(3)
where
Fig. 5 Basic Thermistor Circuit Improved silicon microelectronics manufacturing techniques have brought semiconductors to the point where low-cost, precise temperature sensors are commercially available. Elemental Semiconductors. Because of controlled doping of impurities into elemental germanium, a germanium semiconductor is a reliable temperature sensor for cryogenic temperature measurement in the range of 1 to 84 K. Junction Semiconductors. The first simple junction semiconductor device consisted of a single diode or transistor, in which the forward-connected base emitter voltage was very sensitive to temperature. Today, the more common form is a pair of diodeconnected transistors, which make the device suitable for ambient temperature measurement. Applications include thermocouple reference junction compensation. The primary advantages of silicon transistor temperature sensors are their extreme linearity and exact R0 value, as well as the incorporation of signal conditioning circuitry into the same device as the sensor element. As with thermocouples, these semiconductors require highly precise manufacturing techniques, extremely precise voltage measurements, multiple-point calibration, and temperature
R1 R2 t1, t2 t0 k
= winding resistance at temperature t1, = winding resistance at temperature t2, = winding temperatures, °C = wire resistivity coefficient, °C–1 = reference temperature for resistivity, 25°C = 1/ – t0 = winding material constant, 390.1°C for 100% IACS conductivity copper
The classical method of determining winding temperature is to measure the equipment when it is inoperative and temperaturestabilized at room temperature. After the equipment has operated sufficiently to stabilize temperature under load conditions, the winding resistance should be measured again by taking resistance measurements at known, short time intervals after shutdown. These values may be extrapolated to zero time to indicate the winding resistance at the time of shutdown. The obvious disadvantage of this method is that the device must be shut down to determine winding temperature. A circuit described by Seely (1955), however, makes it possible to measure resistances while the device is operating.
3.3
THERMOCOUPLES
When two wires of dissimilar metals are joined by soldering, welding, or twisting, they form a thermocouple junction or thermojunction. An emf that depends on the wire materials and the junction temperature exists between the wires. This is known as the Seebeck voltage.
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37.8
2017 ASHRAE Handbook—Fundamentals (SI) Table 2
Thermocouple Tolerances on Initial Values of Electromotive Force Versus Temperature Reference Junction Tolerance at 0°Ca
Thermocouple Type
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T J E K N
Temperature Range, °C
Material Identification
R S B
Copper versus constantan Iron versus constantan Nickel/10% chromium versus constantan Nickel/10% chromium versus 5% aluminum, silicon Nickel/14% chromium, 1.5% silicon versus nickel/4.5% silicon, 0.1% magnesium Platinum/13% rhodium versus platinum Platinum/10% rhodium versus platinum Platinum/30% rhodium versus platinum/6% rhodium
Tb Eb Kb
Copper versus constantan Nickel/10% chromium versus constantan Nickel/10% chromium versus 5% aluminum, silicon
Source: ASTM Standard E230. aTolerances in this table apply to new thermocouple wire, normally in the size range of 0.25 to 3 mm diameter and used at temperatures not exceeding recommended limits. Thermocouple wire is available in two grades: standard and special. bThermocouples and thermocouple materials are normally supplied to meet the tolerance specified in the table for temperatures above 0°C. The same materials, however, may not fall within the tolerances given in the second section of the table when operated below freezing (0°C). If materials are required to meet tolerances at subfreezing temperatures, the purchase order must state so.
Thermocouples for temperature measurement yield less precise results than platinum resistance thermometers, but, except for glass thermometers, thermocouples are the most common instruments of temperature measurement for the range of 0 to 1000°C. Because of their low cost, moderate reliability, and ease of use, thermocouples are widely accepted. The most commonly used thermocouples in industrial applications are assigned letter designations. Tolerances of such commercially available thermocouples are given in Table 2. Because the measured emf is a function of the difference in temperature and the type of dissimilar metals used, a known temperature at one junction is required; the remaining junction temperature may be calculated. It is common to call the one with known temperature the (cold) reference junction and the one with unknown temperature the (hot) measured junction. The reference junction is typically kept at a reproducible temperature, such as the ice point of water. Various systems are used to maintain the reference junction temperature (e.g., mixed ice and water in an insulated flask, commercially available thermoelectric coolers to maintain the ice-point temperature automatically in a reference chamber). When these systems cannot be used in an application, measuring instruments with automatic reference junction temperature compensation may be used. These types of instruments typically use a thermistor or RTD to measure the reference junction temperature to provide reference junction temperature compensation. As previously described, the principle for measuring temperature with a thermocouple is based on accurate measurement of the Seebeck voltage. Acceptable DC voltage measurement methods are (1) millivoltmeter, (2) millivolt potentiometer, and (3) high-input impedance digital voltmeter. Many digital voltmeters include builtin software routines for direct calculation and display of temperature. Regardless of the method selected, there are many ways to simplify measurement. Solid-state digital readout devices in combination with a milli- or microvoltmeter, as well as packaged thermocouple readouts with built-in cold junction and linearization circuits, are available. The latter requires a proper thermocouple to provide direct meter reading of temperature. Accuracy approaching or surpassing that of potentiometers can be attained, depending on the instrument quality. This method is popular because it eliminates the null balancing requirement and reads temperature directly in a digital readout.
Standard Tolerance Special Tolerance (Whichever Is Greater) (Whichever Is Greater)
0 to 350 0 to 750 0 to 900 0 to 1250 0 to 1250
±1 K or ±0.75% ±2.2 K or ±0.75% ±1.7 K or ±0.5% ±2.2 K or ±0.75% ±2.2 K or ±0.75%
±0.5 K or ±0.4% ±1.1 K or ±0.4% ±1 K or ±0.4% ±1.1 K or ±0.4% ±1.1 K or ±0.4%
0 to 1450 0 to 1450 870 to 1700
±1.5 K or ±0.25% ±1.5 K or ±0.25% ±0.5%
±0.6 K or ±0.1% ±0.6 K or ±0.1% ±0.25%
–200 to 0 –200 to 0 –200 to 0
±1 K or ±1.5% ±1.7 K or ±1% ±2.2 K or ±2%
c
c c c
cLittle
information is available to justify establishing special tolerances for belowfreezing temperatures. Limited experience suggests the following special tolerances for types E and T thermocouples: Type E Type T
–200 to 0°C; ±1 K or ±0.5% (whichever is greater) –200 to 0°C; ±0.5 K or ±0.8% (whichever is greater)
These tolerances are given only as a guide for discussion between purchaser and supplier.
Wire Diameter and Composition Thermocouple wire is selected by considering the temperature to be measured, the corrosion protection afforded to the thermocouple, and the precision and service life required. Type T thermocouples are suitable for temperatures up to 350°C; type J, up to 750°C; and types K and N, up to 1250°C. Higher temperatures require noble metal thermocouples (type S, R, or B), which have a higher initial cost and do not develop as high an emf as the base metal thermocouples. Thermocouple wires of the same type have small compositional variation from lot to lot from the same manufacturer, and especially among different manufacturers. Consequently, calibrating samples from each wire spool is essential for precision. Calibration data on wire may be obtained from the manufacturer. Computer-friendly reference functions are available for relating temperature and emf of letter-designated thermocouple types. The functions depend on thermocouple type and temperature range; they are used to generate reference tables of emf as a function of temperature, but are not well suited for calculating temperatures directly from values of emf. Approximate inverse functions are available, however, for calculating temperature and are of the form t=
ai E i n
(4)
i=0
where t = temperature, ai = thermocouple constant coefficients, and E = voltage. Burns et al. (1992) give reference functions and approximate inverses for all letter-designated thermocouples. The emf of a thermocouple, as measured with a high-input impedance device, is independent of the diameters of its constituent wires. Thermocouples with small-diameter wires respond faster to temperature changes and are less affected by radiation than larger ones. Large-diameter wire thermocouples, however, are necessary for high-temperature work when wire corrosion is a problem. For use in heated air or gases, thermocouples are often shielded and sometimes aspirated. One way to avoid error caused by radiation is using several thermocouples of different wire sizes and estimating the true temperature by extrapolating readings to zero diameter. With thermocouples, temperatures can be indicated or recorded remotely on conveniently located instruments. Because thermocouples can be made of small-diameter wire, they can be used to
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Measurement and Instruments measure temperatures within thin materials, within narrow spaces, or in otherwise inaccessible locations.
Multiple Thermocouples Thermocouples in series, with alternate junctions maintained at a common temperature, produce an emf that, when divided by the number of thermocouples, gives the average emf corresponding to the temperature difference between two sets of junctions. This series arrangement of thermocouples, often called a thermopile, is used to increase sensitivity and is often used for measuring small temperature changes and differences. Connecting several thermocouples of the same type in parallel with a common reference junction is useful for obtaining an average temperature of an object or volume. In such measurements, however, it is important that the electrical resistances of the individual thermocouples be the same. Use of thermocouples in series and parallel arrangements is discussed in ASTM Manual 12.
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Surface Temperature Measurement The thermocouple is useful in determining surface temperature. It can be attached to a metal surface in several ways. For permanent installations, soldering, brazing, or peening (i.e., driving the thermocouple measuring junction into a small drilled hole) is suggested. For temporary arrangements, thermocouples can be attached by tape, adhesive, or putty-like material. For boiler or furnace surfaces, use furnace cement. To minimize the possibility of error caused by heat conduction along wires, a surface thermocouple should be made of fine wires placed in close contact with the surface being measured for about 25 mm from the junction to ensure good thermal contact. Wires must be insulated electrically from each other and from the metal surface (except at the junction).
Thermocouple Construction Thermocouple (TC) wires are insulated with various materials, including fibrous glass, fluorocarbon resin, and ceramic insulators. Perhaps the most common insulators are polyimides or, in hightemperature applications, braided glass. At high temperatures, insulation can break down, inadvertently forming an unanticipated TC junction. In another form of thermocouple, the wires are insulated with compacted ceramic insulation inside a metal sheath, providing both mechanical protection and protection from stray electromagnetic fields. The measuring junction may be exposed or enclosed within the metal sheath. An enclosed junction may be either grounded or ungrounded to the metal sheath. An exposed junction is in direct contact with the process stream; it is therefore subject to corrosion or contamination, but provides a fast temperature response. A grounded enclosed junction, in which the wires are welded to the metal sheath, provides electrical grounding, as well as mechanical and corrosion protection, but has a slower response time. Response time is even slower for ungrounded enclosed junctions, but the thermocouple wires are isolated electrically and are less susceptible to some forms of mechanical strain than those with grounded construction.
3.4
OPTICAL PYROMETRY
Optical pyrometry determines a surface’s temperature from the color of the radiation it emits. As the temperature of a surface increases, it becomes deep red in color, then orange, and eventually white. This behavior follows from Wein’s law, which indicates that the wavelength corresponding to the maximum intensity of emitted radiation is inversely proportional to the absolute temperature of the emitting surface. Thus, as temperature increases, the wavelength decreases. To determine the unknown surface temperature, the color of radiation from the surface is optically compared to the color of a heated filament. By adjusting the current in the filament, the color of the
37.9 filament is made to match the color of radiation from the source surface. When in balance, the filament virtually disappears into the background image of the surface color. Filament calibration is required to relate the filament current to the unknown surface temperature. For further information, see Holman (2001).
3.5
INFRARED RADIATION THERMOMETERS
Infrared radiation (IR) thermometers, also known as remote temperature sensors (Hudson 1969) or pyrometers, allow noncontact measurement of surface temperature over a wide range. In these instruments, radiant flux from the observed object is focused by an optical system onto an infrared detector that generates an output signal proportional to the incident radiation that can be read from a meter or display unit. Both point and scanning radiometers are available; the latter can display the temperature variation in the field of view. IR thermometers are usually classified according to the detector used: either thermal or photon. In thermal detectors, a change in electrical property is caused by the heating effect of the incident radiation. Examples of thermal detectors are the thermocouple, thermopile, and metallic and semiconductor bolometers. Typical response times are one-quarter to one-half second. In photon detectors, a change in electrical property is caused by the surface absorption of incident photons. Because these detectors do not require an increase in temperature for activation, their response time is much shorter than that of thermal detectors. Scanning radiometers usually use photon detectors. An IR thermometer only measures the power level of radiation incident on the detector, a combination of thermal radiation emitted by the object and surrounding background radiation reflected from the object’s surface. Very accurate measurement of temperature, therefore, requires knowledge of the long-wavelength emissivity of the object as well as the effective temperature of the thermal radiation field surrounding the object. Calibration against an internal or external source of known temperature and emissivity may be needed to obtain true surface temperature from the radiation measurements. In other cases, using published emissivity factors for common materials may suffice. Many IR thermometers have an emissivity adjustment feature that automatically calculates the effect of emissivity on temperature once the emissivity factor is entered. Thermometers that do not have an emissivity adjustment are usually preset to calculate emissivity at 0.95, a good estimate of the emissivity of most organic substances, including paint. Moreover, IR thermometers are frequently used for relative, rather than absolute, measurement; in these cases, adjustment for emissivity may be unnecessary. The most significant practical problem is measuring shiny, polished objects. Placing electrical tape or painting the measurement area with flat black paint and allowing the temperature of the tape or paint to equilibrate can mitigate this problem. A key factor in measurement quality can be the optical resolution or spot size of the IR thermometer, because this specification determines the instrument’s measurement area from a particular distance and, thus, whether a user is actually measuring the desired area. Optical resolution is expressed as distance to spot size (D:S) at the focal. Part of the D:S specification is a description of the amount of target infrared energy encircled by the spot; typically it is 95%, but may be 90%. Temperature resolution of an IR thermometer decreases as object temperature decreases. For example, a radiometer that can resolve a temperature difference of 0.3 K on an object near 20°C may only resolve a difference of 1 K on an object at 0°C.
3.6
INFRARED THERMOGRAPHY
Infrared thermography acquires and analyzes thermal information using images from an infrared imaging system. An infrared
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37.10
2017 ASHRAE Handbook—Fundamentals (SI)
imaging system consists of (1) an infrared video camera and (2) a display unit. The infrared camera scans a surface and senses the selfemitted and reflected radiation viewed from the surface. The display unit contains either a cathode-ray tube (CRT) that displays a graytone or color-coded thermal image of the surface or a color liquid crystal display (LCD) screen. A photograph of the image on the CRT is called a thermogram. Introductions to infrared thermography are given by Madding (1989) and Paljak and Pettersson (1972). Thermography has been used to detect missing insulation and air infiltration paths in building envelopes (Burch and Hunt 1978). Standard practices for conducting thermographic inspections of buildings are given in ASTM Standard C1060. A technique for quantitatively mapping heat loss in building envelopes is given by Mack (1986). Aerial infrared thermography of buildings is effective in identifying regions of an individual built-up roof that have wet insulation (Tobiasson and Korhonen 1985), but it is ineffective in ranking a group of roofs according to their thermal resistance (Burch 1980; Goldstein 1978). In this latter application, the emittances of the separate roofs and outdoor climate (i.e., temperature and wind speed) throughout the microclimate often produce changes in the thermal image that may be incorrectly attributed to differences in thermal resistance. Industrial applications include locating defective or missing pipe insulation in buried heat distribution systems, surveys of manufacturing plants to quantify energy loss from equipment, and locating
defects in coatings (Bentz and Martin 1987). Madding (1989) discusses applications to electrical power systems and electronics.
4.
HUMIDITY MEASUREMENT
Any instrument that can measure the humidity or psychrometric state of air is a hygrometer, and many are available. The indication sensors used on the instruments respond to different moisture property contents. These responses are related to factors such as wetbulb temperature, relative humidity, humidity (mixing) ratio, dew point, and frost point. Table 3 lists instruments for measuring humidity. Each is capable of accurate measurement under certain conditions and within specific limitations. The following sections describe the various instruments in more detail.
4.1
PSYCHROMETERS
A typical industrial psychrometer consists of a pair of matched electrical or mechanical temperature sensors, one of which is kept wet with a moistened wick. A blower aspirates the sensor, which lowers the temperature at the moistened temperature sensor. The lowest temperature depression occurs when the evaporation rate required to saturate the moist air adjacent to the wick is constant. This is a steady-state, open-loop, nonequilibrium process, which depends on the purity of the water, cleanliness of the wick, ventilation rate, radiation effects, size and accuracy of the temperature sensors, and transport properties of the gas.
Table 3 Humidity Sensor Properties Type of Sensor
Sensor Category
Psychrometer
Evaporative cooling
Adiabatic saturation psychrometer Chilled mirror Heated saturated salt solution Hair Nylon Dacron thread Goldbeater’s skin Cellulosic materials Carbon Dunmore type
Method of Operation
Temperature measurement of wet bulb Evaporative cooling Temperature measurement of thermodynamic wet bulb Dew point Optical determination of moisture formation Water vapor pressure Vapor pressure depression in salt solution Mechanical Dimensional change Mechanical Dimensional change Mechanical Dimensional change Mechanical Dimensional change Mechanical Dimensional change Mechanical Dimensional change Electrical Impedance
Polymer film electronic Electrical hygrometer Ion exchange resin Electrical
Impedance or capacitance
Porous ceramic Aluminum oxide
Impedance or capacitance Capacitance Capacitance Electrolyzes due to adsorbed moisture Optical diodes SAW attenuation Mass changes due to adsorbed moisture Moisture absorption of UV or IR radiation Comparison of sample gas with dry airstream Color changes
Electrical Electrical Electrical Electrolytic cell
Electrolytic hygrometer Infrared laser diode Surface acoustic wave Piezoelectric
Electrical Electrical Mass sensitive
Radiation absorption
Moisture absorption
Gravimetric
Direct measurement of mixing ratio Physical
Color change
Impedance or capacitance
Approximate Range Some Uses
Approximate Accuracy
0 to 80°C
Measurement, standard
±3 to 7% rh
5 to 30°C
Measurement, standard
±0.2 to 2% rh
–75 to 95°C dpt
Measurement, control, meteorology
±0.2 to 2 K
–30 to 70°C dpt
Measurement, control, meteorology
±1.5 K
5 to 100% rh 5 to 100% rh 5 to 100% rh 5 to 100% rh 5 to 100% rh 5 to 100% rh 7 to 98% rh at 5 to 60°C 10 to 100% rh
Measurement, control Measurement, control Measurement Measurement Measurement, control Measurement Measurement, control
±5% rh ±5% rh ±7% rh ±7% rh ±5% rh ±5% rh ±1.5% rh
10 to 100% rh at –40 to 90°C Up to 200°C 5 to 100% rh –80 to 60°C dpt 1 to 1000 ppm
Measurement, control
Measurement, control ±1 to 1.5% rh Measurement, control ±3% rh Trace moisture measurement, control ±1 K dpt Measurement
0.1 to 100 ppm 85 to 98% rh –75 to –20°C
Trace moisture measurement ±0.1 ppm Measurement, control ±1% rh Trace moisture measurement, control ±1 to 5 K dpt
–20 to 80°C dpt
Measurement, control, meteorology
120 to 20 000 ppm mixing ratio 10 to 80% rh
Primary standard, research and laboratory Warning device
±2 to 3% rh ±5% rh
±2 K dpt, ±5% rh ±0.13% of reading ±10% rh
Notes: dpt = dew-point temperature 3. Approximate accuracy is based on manufacturers’ data. 1. This table does not encompass all of available technology for measurement of humidity. 4. Presently, NIST only certifies instruments with operating ranges within –75 to 2. Approximate range for device types listed is based on surveys of device manufacturers. 100°C dpt.
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Measurement and Instruments ASHRAE Standard 41.6 recommends an airflow over both the wet and dry bulbs of 3 to 5 m/s for transverse ventilation and 1.5 to 2.5 m/s for axial ventilation. The sling psychrometer consists of two thermometers mounted side by side in a frame fitted with a handle for whirling the device through the air. The thermometers are spun until their readings become steady. In the ventilated or aspirated psychrometer, the thermometers remain stationary, and a small fan, blower, or syringe moves air across the thermometer bulbs. Various designs are used in the laboratory, and commercial models are available. Other temperature sensors, such as thermocouples and thermistors, are also used and can be adapted for recording temperatures or for use where a small instrument is required. Small-diameter wetbulb sensors operate with low ventilation rates. Charts and tables showing the relationship between the temperatures and humidity are available. Data are usually based on a barometric pressure equal to one standard atmosphere. To meet special needs, charts can be produced that apply to nonstandard pressure (e.g., the ASHRAE 2250 m psychrometric chart). Alternatively, mathematical calculations can be made (Kusuda 1965). Uncertainties of 3 to 7% rh are typical for psychrometer-based derivation. The degree of uncertainty is a function of the accuracy of temperature measurements (wet- and dry-bulb), knowledge of the barometric pressure, and conformance to accepted operational procedures such as those outlined in ASHRAE Standard 41.6. In air temperatures below 0°C, water on the wick may either freeze or supercool. Because the wet-bulb temperature is different for ice and water, the state must be known and the proper chart or table used. Some operators remove the wick from the wet bulb for freezing conditions and dip the bulb in water a few times; this allows water to freeze on the bulb between dips, forming a film of ice. Because the wet-bulb depression is slight at low temperatures, precise temperature readings are essential. A psychrometer can be used at high temperatures, but if the wet-bulb depression is large, the wick must remain wet and water supplied to the wick must be cooled so as not to influence the wet-bulb temperature by carrying sensible heat to it (Richardson 1965; Worrall 1965). Greenspan and Wexler (1968) and Wentzel (1961) developed devices to measure adiabatic saturation temperature.
4.2
DEW-POINT HYGROMETERS
Condensation Dew-Point Hygrometers The condensation (chilled-mirror) dew-point hygrometer is an accurate and reliable instrument with a wide humidity range. However, these features are gained at increased complexity and cost compared to the psychrometer. In the condensation hygrometer, a surface is cooled (thermoelectrically, mechanically, or chemically) until dew or frost begins to condense out. The condensate surface is maintained electronically in vapor-pressure equilibrium with the surrounding gas, while surface condensation is detected by optical, electrical, or nuclear techniques. The measured surface temperature is then the dew-point temperature. The largest source of error stems from the difficulty in measuring condensate surface temperature accurately. Typical industrial versions of the instrument are accurate to ±0.5 K over wide temperature spans. With proper attention to the condensate surface temperature measuring system, errors can be reduced to about ±0.2 K. Condensation hygrometers can be made surprisingly compact using solidstate optics and thermoelectric cooling. Wide span and minimal errors are two of the main features of this instrument. A properly designed condensation hygrometer can measure dew points from 95°C down to frost points of –75°C. Typical condensation hygrometers can cool to 80 K below ambient temperature, establishing lower limits of the instrument to dew points corresponding to approximately 0.5% rh. Accuracies for
37.11 measurements above –40°C can be ±1 K or better, deteriorating to ±2 K at lower temperatures. The response time of a condensation dew-point hygrometer is usually specified in terms of its cooling/heating rate, typically 2 K/s for thermoelectric cooled mirrors. This makes it somewhat faster than a heated salt hygrometer. Perhaps the most significant feature of the condensation hygrometer is its fundamental measuring technique, which essentially renders the instrument self-calibrating. For calibration, it is necessary only to manually override the surface cooling control loop, causing the surface to heat, and confirm that the instrument recools to the same dew point when the loop is closed. Assuming that the surface temperature measuring system is correct, this is a reasonable check on the instrument’s performance. Although condensation hygrometers can become contaminated, they can easily be cleaned and returned to service with no impairment to performance.
Salt-Phase Heated Hygrometers Another instrument in which the temperature varies with ambient dew-point temperature is called a self-heating salt-phase transition hygrometer or a heated electrical hygrometer. This device usually consists of a tubular substrate covered by glass fiber fabric, with a spiral bifilar winding for electrodes. The surface is covered with a salt solution, usually lithium chloride. The sensor is connected in series with a ballast and a 24 V (AC) supply. When the instrument is operating, electrical current flowing through the salt film heats the sensor. The salt’s electrical resistance characteristics are such that a balance is reached with the salt at a critical moisture content corresponding to a saturated solution. The sensor temperature adjusts automatically so that the water vapor pressures of the salt film and ambient atmosphere are equal. With lithium chloride, this sensor cannot be used to measure relative humidity below approximately 12% (the equilibrium relative humidity of this salt), and it has an upper dew-point limit of about 70°C. The regions of highest precision are between –23 and 34°C, and above 40°C dew point. Another problem is that the lithium chloride solution can be washed off when exposed to water. In addition, this type of sensor is subject to contamination problems, which limits its accuracy. Its response time is also very slow: it takes approximately 2 min for a 67% step change.
4.3
MECHANICAL HYGROMETERS
Many organic materials change in dimension with changes in humidity; this action is used in a number of simple and effective humidity indicators, recorders, and controllers (see Chapter 7). They are coupled to pneumatic leak ports, mechanical linkages, or electrical transduction elements to form hygrometers. Commonly used organic materials are human hair, nylon, Dacron, animal membrane, animal horn, wood, and paper. Their inherent nonlinearity and hysteresis must be compensated for within the hygrometer. These devices are generally unreliable below 0°C. The response is generally inadequate for monitoring a changing process, and can be affected significantly by exposure to extremes of humidity. Mechanical hygrometers require initial calibration and frequent recalibration; however, they are useful because they can be arranged to read relative humidity directly, and they are simpler and less expensive than most other types.
4.4 ELECTRICAL IMPEDANCE AND CAPACITANCE HYGROMETERS Many substances adsorb or lose moisture with changing relative humidity and exhibit corresponding changes in electrical impedance or capacitance.
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37.12 Dunmore Hygrometers This sensor consists of dual electrodes on a tubular or flat substrate; it is coated with a film containing salt, such as lithium chloride, in a binder to form an electrical connection between windings. The relation of sensor resistance to humidity is usually represented by graphs. Because the sensor is highly sensitive, the graphs are a series of curves, each for a given temperature, with intermediate values found by interpolation. Several resistance elements (Dunmore elements) cover a standard range. Systematic calibration is essential because the resistance grid varies with time and contamination as well as with exposure to temperature and humidity extremes.
Polymer Film Electronic Hygrometers
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These devices consist of a hygroscopic organic polymer deposited by means of thin or thick film processing technology on a waterpermeable substrate. Both capacitance and impedance sensors are available. The impedance devices may be either ionic or electronic conduction types. These hygrometers typically have integrated circuits that provide temperature correction and signal conditioning. The primary advantages of this sensor technology are small size; low cost; fast response times (on the order of 1 to 120 s for 64% change in relative humidity); and good accuracy over the full range, including the low end, where most other devices are less accurate.
Ion Exchange Resin Electric Hygrometers A conventional ion exchange resin consists of a polymer with a high relative molecular mass and polar groups of positive or negative charge in cross-link structure. Associated with these polar groups are ions of opposite charge that are held by electrostatic forces to the fixed polar groups. In the presence of water or water vapor, the electrostatically held ions become mobile; thus, when a voltage is impressed across the resin, the ions are capable of electrolytic conduction. The Pope cell is one example of an ion exchange element. It is a wide-range sensor, typically covering 15 to 95% rh; therefore, one sensor can be used where several Dunmore elements would be required. The Pope cell, however, has a nonlinear characteristic from approximately 1000 at 100% rh to several megohms at 10% rh.
Impedance-Based Porous Ceramic Electronic Hygrometers Using oxides’ adsorption characteristics, humidity-sensitive ceramic oxide devices use either ionic or electronic measurement techniques to relate adsorbed water to relative humidity. Ionic conduction is produced by dissociation of water molecules, forming surface hydroxyls. The dissociation causes proton migration, so the device’s impedance decreases with increasing water content. The ceramic oxide is sandwiched between porous metal electrodes that connect the device to an impedance-measuring circuit for linearizing and signal conditioning. These sensors have excellent sensitivity, are resistant to contamination and high temperature (up to 200°C), and may get fully wet without sensor degradation. These sensors are accurate to about ±1.5% rh (±1% rh when temperature compensated) and have a moderate cost.
Aluminum Oxide Capacitive Sensor This sensor consists of an aluminum strip that is anodized by a process that forms a porous oxide layer. A very thin coating of cracked chromium or gold is then evaporated over this structure. The aluminum base and cracked chromium or gold layer form the two electrodes of what is essentially an aluminum oxide capacitor. Water vapor is rapidly transported through the cracked chromium or gold layer and equilibrates on the walls of the oxide pores in a manner functionally related to the vapor pressure of water in the atmosphere surrounding the sensor. The number of water molecules
2017 ASHRAE Handbook—Fundamentals (SI) adsorbed on the oxide structure determines the capacitance between the two electrodes.
4.5 ELECTROLYTIC HYGROMETERS In electrolytic hygrometers, air is passed through a tube, where moisture is adsorbed by a highly effective desiccant (usually phosphorous pentoxide) and electrolyzed. The airflow is regulated to 1.65 mL/s at a standard temperature and pressure. As the incoming water vapor is absorbed by the desiccant and electrolyzed into hydrogen and oxygen, the current of electrolysis determines the mass of water vapor entering the sensor. The flow rate of the entering gas is controlled precisely to maintain a standard sample mass flow rate into the sensor. The instrument is usually designed for use with moisture/air ratios in the range of less than 1 ppm to 1000 ppm, but can be used with higher humidities.
4.6
PIEZOELECTRIC SORPTION
This hygrometer compares the changes in frequency of two hygroscopically coated quartz crystal oscillators. As the crystal’s mass changes because of absorption of water vapor, the frequency changes. The amount of water sorbed on the sensor is a function of relative humidity (i.e., partial pressure of water as well as ambient temperature). A commercial version uses a hygroscopic polymer coating on the crystal. Humidity is measured by monitoring the change in the vibration frequency of the quartz crystal when the crystal is alternately exposed to wet and dry gas.
4.7 SPECTROSCOPIC (RADIATION ABSORPTION) HYGROMETERS Radiation absorption devices operate on the principle that selective absorption of radiation is a function of frequency for different media. Water vapor absorbs infrared radiation at 2 to 3 m wavelengths and ultraviolet radiation centered about the Lyman-alpha line at 0.122 m. The amount of absorbed radiation is directly related to the absolute humidity or water vapor content in the gas mixture, according to Beer’s law. The basic unit consists of an energy source and optical system for isolating wavelengths in the spectral region of interest, and a measurement system for determining the attenuation of radiant energy caused by water vapor in the optical path. Absorbed radiation is measured extremely quickly and independent of the degree of saturation of the gas mixture. Response times of 0.1 to 1 s for 90% change in moisture content are common. Spectroscopic hygrometers are primarily used where a noncontact application is required; this may include atmospheric studies, industrial drying ovens, and harsh environments. The primary disadvantages of this device are its high cost and relatively large size.
4.8
GRAVIMETRIC HYGROMETERS
Humidity levels can be measured by extracting and finding the mass of water vapor in a known quantity or atmosphere. For precise laboratory work, powerful desiccants (e.g., phosphorous pentoxide, magnesium perchlorate) are used for extraction; for other purposes, calcium chloride or silica gel is satisfactory. When the highest level of accuracy is required, the NIST gravimetric hygrometer is recommended. The gravimetric hygrometer gives the absolute water vapor content, where the mass of absorbed water and precise measurement of the gas volume associated with the water vapor determine the mixing ratio or absolute humidity of the sample. This system is the primary standard because the required measurements of mass, temperature, pressure, and volume can be made with extreme precision. However, its complexity and required attention to detail limit its usefulness.
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Measurement and Instruments 4.9
CALIBRATION
For many hygrometers, the need for recalibration depends on the accuracy required, the sensor’s stability, and the conditions to which the sensor is subjected. Many hygrometers should be calibrated regularly by exposure to an atmosphere maintained at a known humidity and temperature, or by comparison with a transfer standard hygrometer. Complete calibration usually requires observation of a series of temperatures and humidities. Methods for producing known humidities include saturated salt solutions (Greenspan 1977); sulfuric acid solutions; and mechanical systems, such as the divided flow, two-pressure (Amdur 1965); twotemperature (Till and Handegord 1960); and NIST two-pressure humidity generator (Hasegawa 1976). All these systems rely on precise methods of temperature and pressure control in a controlled environment to produce a known humidity, usually with accuracies of 0.5 to 1.0%. The operating range for the precision generator is typically 5 to 95% rh.
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5.
PRESSURE MEASUREMENT
Pressure is the force exerted per unit area by a medium, generally a liquid or gas. Pressure so defined is sometimes called absolute pressure. Thermodynamic and material properties are expressed in terms of absolute pressures; thus, the properties of a refrigerant are given in terms of absolute pressures. Vacuum refers to pressures below atmospheric. Differential pressure is the difference between two absolute pressures, or the difference between two relative pressures measured with respect to the same reference pressure. Often, it can be very small compared to either of the absolute pressures (these are often referred to as low-range, high-line differential pressures). A common example of differential pressure is the pressure drop, or difference between inlet and outlet pressures, across a filter or flow element. Gage pressure is a special case of differential pressure where the reference pressure is atmospheric pressure. Many pressure gages, including most refrigeration test sets, are designed to make gage pressure measurements, and there are probably more gage pressure measurements made than any other. Gage pressure measurements are often used as surrogates for absolute pressures. However, because of variations in atmospheric pressure caused by elevation (e.g., atmospheric pressure in Denver, Colorado, is about 81% of sea-level pressure) and weather changes, using gage pressures to determine absolute pressures can significantly restrict the accuracy of the measured pressure, unless corrections are made for the local atmospheric pressure at the time of measurement. Pressures can be further classified as static or dynamic. Static pressures have a small or undetectable change with time; dynamic pressures include a significant pulsed, oscillatory, or other timedependent component. Static pressure measurements are the most common, but equipment such as blowers and compressors can generate significant oscillatory pressures at discrete frequencies. Flow in pipes and ducts can generate resonant pressure changes, as well as turbulent “noise” that can span a wide range of frequencies.
Units A plethora of pressure units, many of them poorly defined, are in common use. The international (SI) unit is the newton per square metre, called the pascal (Pa). Although the bar and standard atmosphere are used, they should not be introduced where they are not used at present.
5.1
INSTRUMENTS
Broadly speaking, pressure instruments can be divided into three different categories: standards, mechanical gages, and electromechanical transducers. Standards instruments are used for the most
37.13 accurate calibrations. The liquid-column manometer, which is the most common and potentially the most accurate standard, is used for a variety of applications, including field applications. Mechanical pressure gages are generally the least expensive and the most common. However, electromechanical transducers have become much less expensive and are easier to use, so they are being used more often.
Pressure Standards Liquid-column manometers measure pressure by determining the vertical displacement of a liquid of known density in a known gravitational field. Typically, they are constructed as a U-tube of transparent material (glass or plastic). The pressure to be measured is applied to one side of the U-tube. If the other (reference) side is evacuated (zero pressure), the manometer measures absolute pressure; if the reference side is open to the atmosphere, it measures gage pressure; if the reference side is connected to some other pressure, the manometer measures the differential between the two pressures. Manometers filled with water and different oils are often used to measure low-range differential pressures. In some low-range instruments, one tube of the manometer is inclined to enhance readability. Mercury-filled manometers are used for higher-range differential and absolute pressure measurements. In the latter case, the reference side is evacuated, generally with a mechanical vacuum pump. Typical full-scale ranges for manometers vary from 25 Pa to 300 kPa. For pressures above the range of manometers, standards are generally of the piston-gage, pressure-balance, or deadweight-tester type. These instruments apply pressure to the bottom of a vertical piston, which is surrounded by a close-fitting cylinder (typical clearances are micrometres). The pressure generates a force approximately equal to the pressure times the area of the piston. This force is balanced by weights stacked on the top of the piston. If the mass of the weights, local acceleration of gravity, and area of the piston (or more properly, the “effective area” of the piston and cylinder assembly) are known, the applied pressure can be calculated. Piston gages usually generate gage pressures with respect to the atmospheric pressure above the piston. They can be used to measure absolute pressures either indirectly, by separately measuring the atmospheric pressure and adding it to the gage pressure determined by the piston gage, or directly, by surrounding the top of the piston and weights with an evacuated bell jar. Piston gage full-scale ranges vary from 35 kPa to 1.4 GPa. At very low absolute pressures (below about 100 Pa), a number of different types of standards are used. These tend to be specialized and expensive instruments found only in major standards laboratories. However, one low-pressure standard, the McLeod gage, has been used for field applications. Unfortunately, although its theory is simple and straightforward, it is difficult to use accurately, and major errors can occur when it is used to measure gases that condense or are adsorbed (e.g., water). In general, other gages should be used for most low-pressure or vacuum applications.
Mechanical Pressure Gages Mechanical pressure gages couple a pressure sensor to a mechanical readout, typically a pointer and dial. The most common type uses a Bourdon tube sensor, which is essentially a coiled metal tube of circular or elliptical cross section. Increasing pressure applied to the inside of the tube causes it to uncoil. A mechanical linkage translates the motion of the end of the tube to the rotation of a pointer. In most cases, the Bourdon tube is surrounded by atmospheric pressure, so that the gages measure gage pressure. A few instruments surround the Bourdon tube with a sealed enclosure that can be evacuated for absolute measurements or connected to another pressure for differential measurements. Available instruments vary widely in cost, size, pressure range, and accuracy. Full-scale ranges can vary
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37.14 from 35 kPa to 700 MPa. Accuracy of properly calibrated and used instruments can vary from 0.1 to 10% of full scale. Generally there is a strong correlation between size, accuracy, and price; larger instruments are more accurate and expensive. For better sensitivity, some low-range mechanical gages (sometimes called aneroid gages) use corrugated diaphragms or capsules as sensors. The capsule is basically a short bellows sealed with end caps. These sensors are more compliant than a Bourdon tube, and a given applied pressure causes a larger deflection of the sensor. The inside of a capsule can be evacuated and sealed to measure absolute pressures or connected to an external fitting to allow differential pressures to be measured. Typically, these gages are used for lowrange measurements of 100 kPa or less. In better-quality instruments, accuracies can be 0.1% of reading or better.
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Electromechanical Transducers Mechanical pressure gages are generally limited by inelastic behavior of the sensing element, friction in the readout mechanism, and limited resolution of the pointer and dial. These effects can be eliminated or reduced by using electronic techniques to sense the distortion or stress of a mechanical sensing element and electronically convert that stress or distortion to a pressure reading. Various sensors are used, including Bourdon tubes, capsules, diaphragms, and different resonant structures whose vibration frequency varies with the applied pressure. Capacitive, inductive, and optical lever sensors are used to measure the sensor element’s displacement. In some cases, feedback techniques may be used to constrain the sensor in a null position, minimizing distortion and hysteresis of the sensing element. Temperature control or compensation is often included. Readout may be in the form of a digital display, analog voltage or current, or a digital code. Size varies, but for transducers using a diaphragm fabricated as part of a silicon chip, the sensor and signal-conditioning electronics can be contained in a small transistor package, and the largest part of the device is the pressure fitting. The best of these instruments achieve long-term instabilities of 0.01% or less of full scale, and corresponding accuracies when properly calibrated. Performance of less-expensive instruments can be more on the order of several percent. Although the dynamic response of most mechanical gages is limited by the sensor and readout, the response of some electromechanical transducers can be much faster, allowing measurements of dynamic pressures at frequencies up to 1 kHz and beyond in the case of transducers specifically designed for dynamic measurements. Consult manufacturers’ literature as a guide to the dynamic response of specific instruments. As the measured pressure drops below about 10 kPa, it becomes increasingly difficult to sense mechanically. Various gages have been developed that measure some other property of the gas that is related to the pressure. In particular, thermal conductivity gages, also known as thermocouple, thermistor, Pirani, and convection gages, are used for pressures down to about 0.1 Pa. These gages have a sensor tube with a small heated element and a temperature sensor; the temperature of the heated element is determined by the thermal conductivity of the gas, and the output of the temperature sensor is displayed on an analog or digital electrical meter contained in an attached electronics unit. The accuracy of thermal conductivity gages is limited by their nonlinearity, dependence on gas species, and tendency to read high when contaminated. Oil contamination is a particular problem. However, these gages are small, reasonably rugged, and relatively inexpensive; in the hands of a typical user, they give far more reliable results than a McLeod gage. They can be used to check the base pressure in a system that is being evacuated before being filled with refrigerant. They should be checked periodically for contamination by comparing the reading with that from a new, clean sensor tube.
2017 ASHRAE Handbook—Fundamentals (SI) General Considerations Accurate values of atmospheric or barometric pressure are required for weather prediction and aircraft altimetry. In the United States, the National Weather Service, Federal Aviation Administration, and local airport operating authorities maintain a network of calibrated instruments, generally accurate to within 0.1% of reading and located at airports. These agencies are usually cooperative in providing current values of atmospheric pressure that can be used to check the calibration of absolute pressure gages or to correct gage pressure readings to absolute pressures. However, pressure readings generally reported for weather and altimetry purposes are not the true atmospheric pressure, but rather a value adjusted to an equivalent sea-level pressure. Therefore, unless the location is near sea level, it is important to ask for the station or true atmospheric pressure rather than using the adjusted values broadcast by radio stations. Further, atmospheric pressure decreases with increasing elevation at a rate (near sea level) of about 10 Pa/m, and corresponding corrections should be made to account for the difference in elevation between the instruments being compared. Gage-pressure instruments are sometimes used to measure absolute pressures, but their accuracy can be compromised by uncertainties in atmospheric pressure. This error can be particularly serious when gage-pressure instruments are used to measure vacuum (negative gage pressures). For all but the crudest measurements, absolute-pressure gages should be used for vacuum measurements; for pressures below about 100 Pa, a thermal conductivity gage should be used. All pressure gages are susceptible to temperature errors. Several techniques are used to minimize these errors: sensor materials are generally chosen to minimize temperature effects, mechanical readouts can include temperature compensation elements, electromechanical transducers may include a temperature sensor and compensation circuit, and some transducers operate at a controlled temperature. Clearly, temperature effects are of greater concern for field applications, and it is prudent to check the manufacturers’ literature for the temperature range over which the specified accuracy can be maintained. Abrupt temperature changes can also cause large transient errors that may take some time to decay. Readings of some electromechanical transducers with a resonant or vibrating sensor can depend on the gas species. Although some of these units can achieve calibrated accuracies of the order of 0.01% of reading, they are typically calibrated with dry air or nitrogen, and readings for other gases can be in error by several percent, possibly much more for refrigerants and other high-density gases. Highaccuracy readings can be maintained by calibrating these devices with the gas to be measured. Consult manufacturers’ literature. Measuring dynamic pressures is limited not just by the frequency response of the pressure gage, but also by the hydraulic or pneumatic time constant of the connection between the gage and the system to be monitored. Generally, the longer the connecting lines and the smaller their diameter, the lower the system’s frequency response. Further, even if only the static component of the pressure is of interest, and a gage with a low-frequency response is used, a significant pulsating or oscillating pressure component can cause significant errors in pressure gage readings and, in some cases, can damage the gage, particularly one with a mechanical readout mechanism. In these cases, a filter or snubber should be used to reduce the higherfrequency components.
6.
AIR VELOCITY MEASUREMENT
HVAC engineers measure the flow of air more often than any other gas, and usually at or near atmospheric pressure. Under this condition, air can be treated as an incompressible (i.e., constantdensity) fluid, and simple formulas give sufficient precision to solve
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Measurement and Instruments many problems. Instruments that measure fluid velocity and their application range and precision are listed in Table 4.
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6.1
AIRBORNE TRACER TECHNIQUES
Tracer techniques are suitable for measuring velocity in an open space. Typical tracers include smoke, feathers, pieces of lint, and radioactive or nonradioactive gases. Measurements are made by timing the rate of movement of solid tracers or by monitoring the change in concentration level of gas tracers. Smoke is a useful qualitative tool in studying air movements. Smoke can be obtained from titanium tetrachloride (irritating to nasal membranes) or by mixing potassium chlorate and powdered sugar (nonirritating) and firing the mixture with a match. The latter process produces considerable heat and should be confined to a pan away from flammable materials. Titanium tetrachloride smoke works well for spot tests, particularly for leakage through casings and ducts, because it can be handled easily in a small, pistol-like ejector. Another alternative is theatrical smoke, which is nontoxic, but requires proper illumination. Fumes of ammonia water and sulfuric acid, if allowed to mix, form a white precipitate. Two bottles, one containing ammonia water and the other containing acid, are connected to a common nozzle by rubber tubing. A syringe forces air over the liquid surfaces in the bottles; the two streams mix at the nozzle and form a white cloud. A satisfactory test smoke also can be made by bubbling an airstream through ammonium hydroxide and then hydrochloric acid (Nottage et al. 1952). Smoke tubes, smoke candles, and smoke bombs are available for studying airflow patterns.
6.2 ANEMOMETERS Deflecting Vane Anemometers The deflecting vane anemometer consists of a pivoted vane enclosed in a case. Air exerts pressure on the vane as it passes through the instrument from an upstream to a downstream opening. A hair spring and a damping magnet resist vane movement. The instrument gives instantaneous readings of directional velocities on an indicating scale. With fluctuating velocities, needle swings must be visually averaged. This instrument is useful for studying air motion in a room, locating objectionable drafts, measuring air velocities at supply and return diffusers and grilles, and measuring laboratory hood face velocities.
Propeller or Revolving (Rotating) Vane Anemometers The propeller anemometer consists of a light, revolving, winddriven wheel connected through a gear train to a set of recording dials that read linear metres of air passing in a measured length of time. It is made in various sizes, though 75, 100, and 150 mm are the most common. Each instrument requires individual calibration. At low velocities, the mechanism’s friction drag is considerable, and is usually compensated for by a gear train that overspeeds. For this reason, the correction is often additive at the lower range and subtractive at the upper range, with the least correction in the middle range. The best instruments have starting speeds of 0.25 m/s or higher; therefore, they cannot be used below that air speed. Electronic revolving vane anemometers, with optical or magnetic pickups to sense the rotation of the vane, are available in vane sizes as small as 13 mm.
Cup Anemometers The cup anemometer is primarily used to measure outdoor, meteorological wind speeds. It consists of three or four hemispherical cups mounted radially from a vertical shaft. Wind from any direction with a vector component in the plane of cup rotation causes the cups and shaft to rotate. Because it is primarily used to measure
37.15 meteorological wind speeds, the instrument is usually constructed so that wind speeds can be recorded or indicated electrically at a remote point.
Thermal Anemometers The thermal (or hot-wire, or hot-film) anemometer consists of a heated RTD, thermocouple junction, or thermistor sensor constructed at the end of a probe; it is designed to provide a direct, simple method of determining air velocity at a point in the flow field. The probe is placed into an airstream, and air movement past the electrically heated velocity sensor tends to cool the sensor in proportion to the speed of the airflow. The electronics and sensor are commonly combined into a portable, hand-held device that interprets the sensor signal and provides a direct reading of air velocity in either analog or digital display format. Often, the sensor probe also incorporates an ambient temperature-sensing RTD or thermistor, in which case the indicated air velocity is temperature compensated to standard air density conditions (typically 1.20 kg/m3). Thermal anemometers have long been used in fluid flow research. Research anemometer sensors have been constructed using very fine wires in configurations that allow characterization of fluid flows in one, two, and three dimensions, with sensor/electronics response rates up to several hundred kilohertz. This technology has been incorporated into more ruggedized sensors suitable for measurements in the HVAC field, primarily for unidirectional airflow measurement. Omnidirectional sensing instruments suitable for thermal comfort studies are also available. The principal advantages of thermal anemometers are their wide dynamic range and their ability to sense extremely low velocities. Commercially available portable instruments often have a typical accuracy (including repeatability) of 2 to 5% of reading over the entire velocity range. Accuracies of ±2% of reading or better are obtainable from microcontroller (microprocessor)-based thermistor and RTD sensor assemblies, some of which can be factorycalibrated to known reference standards (e.g., NIST air speed tunnels). An integrated microcontroller also allows an array of sensor assemblies to be combined in one duct or opening, providing independently derived velocity and temperature measurements at each point. Limitations of thermistor-based velocity measuring devices depend on sensor configuration, specific thermistor type used, and the application. At low velocities, thermal anemometers can be significantly affected by their own thermal plumes (from self heating). Products using this technology can be classified as handheld instruments or permanently mounted probes and arrays, and as those with analog electronic transmitters and those that are microcontroller-based. Limitations of hand-held and analog electronic thermal anemometers include the following: (1) the unidirectional sensor must be carefully aligned in the airstream (typically to within ±20° rotation) to achieve accurate results; (2) the velocity sensor must be kept clean because contaminant build-up can change the calibration (which may change accuracy performance); and (3) because of the inherent high speed of response of thermal anemometers, measurements in turbulent flows can yield fluctuating velocity measurements. Electronically controlled time-integrated functions are now available in many digital air velocity meters to help smooth these turbulent flow measurements. Microcontroller-based thermal dispersion devices are typically configured as unidirectional instruments, but may have multiple velocity-sensing elements capable of detecting flow direction. These devices can be used to measure a “bleed” air velocity between two spaces or across a fixed orifice. With mathematical conversion, these measured velocities can closely approximate equivalents in differential pressure down to two decimal places (Pa). They can be used for space pressure control, to identify minute changes in flow
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37.16
2017 ASHRAE Handbook—Fundamentals (SI) Table 4
Licensed for single user. © 2017 ASHRAE, Inc.
Measurement Means
Application
Air Velocity Measurement
Range, m/s
Precision
Smoke puff or airborne Low air velocities in rooms; 0.025 to 0.25 10 to 20% solid tracer highly directional Deflecting vane aneAir velocities in rooms, at out0.15 to 120 5% mometer lets, etc.; directional Revolving (rotating) vane Moderate air velocities in ducts 0.5 to 15 2 to 5% anemometer and rooms; somewhat directional Thermal (hot-wire or a. Low air velocities; directional 0.05 to 50 2 to 10% hot-film) anemometer and omnidirectional available b. Transient velocity and turbulence Pitot-static tube Standard (typically hand-held) 0.9 to 50 with 2 to 5% instrument for measuring micromanometer; 3 single-point duct velocities to 50 with draft gages; 50 up with manometer Impact tube and sidewall High velocities, small tubes, and 0.6 to 50 with 2 to 5% or other static tap where air direction may be micromanometer; 3 variable to 50 with draft gages; 50 up with manometer Cup anemometer Meteorological Up to 60 2 to 5% Ultrasonic Large instruments: 0.005 to 30 1 to 2% meteorological Small instruments: in-duct and room air velocities Laser Doppler velocime- Calibration of air velocity instru0.005 to 30 1 to 3% ter (LDV) ments Particle image velocime- Full-field (2D, 3D) velocity meatry (PIV) surements in rooms, outlets
0.005 to 30
Pitot array, self-averaging In duct assemblies, ducted or fan differential pressure, inlet probes typically using equalizing manifolds
3 to 50
Piezometer and piezoCentrifugal fan inlet cone ring variations, selfaveraging differential pressure using equalizing manifolds
3 to 50
Vortex shedding
In-duct assemblies, ducted or fan inlet probes
2 to 30
Thermal (analog electronic) using thermistors
In-duct assemblies or ducted probes
0.25 to 25
Drag force
In-duct flow
0.1 to 50
Thermal dispersion Ducted or fan inlet probes, bleed (microcontroller-based) velocity sensors using thermistors to independently determine temperatures and velocities Thermal (analog elecIn-duct assemblies or ducted tronic) using RTDs probes; stainless steel and platinum RTDs have industrial environment capabilities
0.1 to 50
0.5 to 90
Limitations Awkward to use but valuable in tracing air movement. Requires periodic calibration check. Subject to significant errors when variations in velocities with space or time are present. Easily damaged. Affected by turbulence intensity. Requires periodic calibration. Requires accurate calibration at frequent intervals. Some are relatively costly. Affected by thermal plume because of selfheating. Accuracy falls off at low end of range because of square-root relationship between velocity and dynamic pressure. Also affected by alignment with flow direction. Accuracy depends on constancy of static pressure across stream section.
Poor accuracy at low air velocity (<2.5 m/s). High cost.
High cost and complexity limit LDVs to laboratory applications. Requires seeding of flow with particles, and transparent optical access (window). 10% High cost and complexity limits measurements to laboratory applications. Requires seeding of flow with particles, and transparent optical access (window). ±2 to >40% Performance depends heavily on quality and range of associof reading ated differential pressure transmitter. Very susceptible to measurement errors caused by duct placement and temperature changes. Nonlinear output (square-root function). Mathematical averaging errors likely because of sampling method. Must be kept clean to function properly. Must be set up and field calibrated to hand-held reference, or calibrated against nozzle standard. ±5 to >40% Performance depends heavily on quality and range of of reading required differential pressure transmitter. Very susceptible to measurement errors caused by inlet cone placement, inlet obstructions, and temperature changes. Nonlinear output (square-root function). Must be kept clean. Must be field calibrated to hand-held reference. ±2.5 to 10% Highest cost per sensing point. Largest physical size. Lowof reading temperature accuracy questionable. Must be set up and field calibrated to hand-held reference. ±2 to 40% Mathematical averaging errors may be caused by analog elecof reading tronic circuitry when averaging nonlinear signals. Sensing points may not be independent. May not be able to compensate for temperatures beyond a narrow range. Must be set up and field calibrated to hand-held reference. Must be recalibrated regularly to counteract drift. ±2% Piezoelectric or strain-gage methods are used to sense dynamic drag-force variations. ±2 to 10% Cost increases with number of sensor assemblies in array. of reading Honeycomb air straighteners are recommended by some manufacturers. Accuracy verified only to –29°C. Not suitable for abrasive or high-temperature environments. ±1 to 20% of reading
Requires long duct/pipe runs. Sensitive to placement conditions. Mathematical averaging errors may be caused by analog electronic circuitry when averaging nonlinear signals. Must be recalibrated regularly to counteract drift. Fairly expensive.
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Measurement and Instruments
37.17
direction, or for estimating volumetric flow rates across a fixed orifice by equating to velocity pressure. Thermal anemometers are suitable for a variety of HVAC applications. They are particularly well suited to the low velocities associated with outdoor air intake measurement and control, return or relief fan tracking for pressurization in variable-air-volume (VAV) systems, VAV terminal box measurement, unit ventilator and packaged equipment intake measurement, space pressurization for medical isolation, and laboratory fume hood face velocity measurements (typically in the 0.25 to 1 m/s range). Thermal anemometers can also take multipoint traverse measurements in ventilation ductwork.
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Laser Doppler Velocimeters (or Anemometers) The laser Doppler velocimeter (LDV) or laser Doppler anemometer (LDA) is an extremely complex system that collects scattered light produced by particles (i.e., seed) passing through the intersection volume of two intersecting laser beams of the same light frequency, which produces a regularly spaced fringe pattern (Mease et al. 1992). The scattered light consists of bursts containing regularly spaced oscillations whose frequency is linearly proportional to the speed of the particle. Because of their cost and complexity, they are usually not suitable for in situ field measurements. Rather, the primary HVAC application of LDV systems is calibrating systems used to calibrate other air velocity instruments. The greatest advantage of an LDV is its performance at low air speeds: as low as 0.075 m/s with uncertainty levels of 1% or less (Mease et al. 1992). In addition, it is nonintrusive in the flow; only optical access is required. It can be used to measure fluctuating components as well as mean speeds and is available in one-, two-, and even three-dimensional configurations. Its biggest disadvantages are its high cost and extreme technological complexity, which requires highly skilled operators. Modern fiber-optic systems require less operator skill but at a considerable increase in cost.
Particle Image Velocimetry (PIV) Particle image velocimetry (PIV) is an optical method that measures fluid velocity by determining the displacement of approximately neutrally buoyant seed particles introduced in the flow. Particle displacements are determined from images of particle positions at two instants of time. Usually, statistical (correlation) methods are used to identify the displacement field. The greatest advantage of PIV is its ability to examine two- and three-dimensional velocity fields over a region of flow. The method usually requires laser light (sheet) illumination, and is typically limited to a field area of less than 1 m2. Accuracy is usually limited to about ±10% by the resolution of particle displacements, which must be small enough to remain in the field of view during the selected displacement time interval. For more comprehensive information on PIV, including estimates of uncertainty, see Raffel et al. (1998).
= density of air, kg/m3
The type of manometer or differential pressure transducer used with a pitot-static tube depends on the magnitude of velocity pressure being measured and on the desired accuracy. Over 7.5 m/s, a draft gage of appropriate range is usually satisfactory. If the pitotstatic tube is used to measure air velocities lower than 7.5 m/s, a precision manometer or comparable pressure differential transducer is essential. Example 1. Step 1. Numerical evaluation. Let pw = 93.16 ± 0.95 Pa and = 1.185 ± 0.020 kg/m3. Then, 2p w 2 93.16 --------- = --------------------- = 12.54 m/s 1.185 Step 2. Uncertainty estimate. Let the typical bias (i.e., calibration) uncertainty of the pitot tube be uV,bias = ±1% of reading. The uncertainty in the velocity measurement is thus estimated to be
V=
uV =
u V , bias + u V , prec
2
2
=
1 2 u V , bias + --- u pw 2
2
=
1 0.95 2 0.01 + --- ------------- 2 93.16
1 + --- u 2 2
1 0.020 + --- ------------- 2 1.185
Therefore, UV = ±uVV = ±(0.014)(12.54 m/s) = ±0.18 m/s In summary, V = 12.54 ± 0.18 m/s
Other pitot-static tubes have been used and calibrated. To meet special conditions, various sizes of pitot-static tubes geometrically similar to the standard tube can be used. For relatively high velocities in ducts of small cross-sectional area, total pressure readings can be obtained with an impact (pitot) tube. Where static pressure across the stream is relatively constant, as in turbulent flow in a straight duct, a sidewall tap to obtain static pressure can be used with the impact tube to obtain the velocity pressure. One form of impact tube is a small streamlined tube with a fine hole in its upstream end and its axis parallel to the stream.
The pitot-static tube, in conjunction with a suitable manometer or differential pressure transducer, provides a simple method of determining air velocity at a point in a flow field. Figure 6 shows the construction of a standard pitot tube (ASHRAE Standard 51) and the method of connecting it with inclined manometers to display both static pressure and velocity pressure. The equation for determining air velocity from measured velocity pressure is 2p w --------
(5)
where V = velocity, m/s pw = velocity pressure (pitot-tube manometer reading), Pa
2
= 0.014 = 1.4%
6.3 PITOT-STATIC TUBES
V=
2
Fig. 6
Standard Pitot Tube
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37.18
2017 ASHRAE Handbook—Fundamentals (SI)
Fig. 7
Pitot-Static Probe Pressure Coefficient Yaw Angular Dependence
If the Mach number of the flow is greater than about 0.3, the effects of compressibility should be included in the computation of the air speed from pitot-static and impact (stagnation or pitot) tube measurements (Mease et al. 1992). It is extremely important to recognize that the pitot-static probe is designed to make measurements when aligned with the flow. Misalignment in yaw angle of up to about 15 to 20° generally do not result in large errors; however, for greater angles, errors can be very large. For large misalignment with flow, the total pressure port of a pitot-static probe does not measure the true total (or stagnation) pressure, and the static pressure ports likewise do not measure the true static pressure of the flow stream. The error in the probe can be represented as a function of tilt (yaw or pitch) angle in terms of a pressure coefficient defined as follows: p total – p static Cp() --------------------------------------------------------p total 0 – p static 0
(6)
where ptotal () is the pressure registered at the total pressure port (see Figures 6 and 7A), and pstatic() is the pressure registered at the static pressure port (see Figure 6) at tilt angle . Note that, at a tilt angle of 0°, the probe is correctly aligned with flow and the total pressure and static pressure are correctly registered at each of the corresponding ports. Figure 7B shows the typical yaw (or pitch) angle dependence of a pitot-static probe subjected to a uniform velocity field U in a wind tunnel, as shown in Figure 7A. The polar plot shows the variation of pressure coefficient Cp() with yaw (or pitch) angle over the entire 360° range (essentially a symmetrical ±180 degrees). A pressure coefficient of Cp = 1 corresponds to a situation of good alignment with the flow and thus negligible error. Note that the pressure coefficient varies from +1 to –1 over the entire range of yaw (or pitch) angles. In reverse flows ( near 180°), output hovers around zero flow coefficient. Because the pitot-static probe does not provide a flow direction indication, it is not possible to determine the particular region of yaw (or pitch) angle operation. Therefore, the correct output for assessing volumetric flow rate (where it is desired to measure the axial flow component) cannot be determined with confidence (Hickman et al. 2012, 2015a, 2015b).
6.4
MEASURING FLOW IN DUCTS
Because velocity in a duct is seldom uniform across any section, and a pitot tube reading or thermal anemometer indicates velocity at only one location, a traverse is usually made to determine average velocity. Generally, velocity is lowest near the side-wall edges or corners and greatest at or near the center of a duct. To determine velocity in a traverse plane, a straight average of individual point velocities gives satisfactory results when point velocities are determined by the log-Tchebycheff (log-T) rule or, if care is taken, by the equal-area method. Figure 8 shows suggested sensor locations for traversing round and rectangular ducts. The logTchebycheff rule provides the greatest accuracy because its location of traverse points accounts for the effect of wall friction and the falloff of velocity near wall ducts. For single-path disturbances (straight ducts, transitions, and elbow fittings), the equal-area method has been shown to give a consistent 3 to 4% positive bias, regardless of probe type (pitot-static, hot-wire anemometer), volumetric flow rate, or traverse location within 7.5 equivalent diameters downstream of a fitting disturbance (Hickman et al. 2012, 2015a, 2015b). The log-T method is now recommended for rectangular ducts with H and W > 460 mm. For circular ducts, the log-T and log-linear methods are similar. Log-T minimizes the positive error (measured greater than actual) caused by the failure to account for losses at the duct wall. This error can occur when using the older method of equal subareas to traverse rectangular ducts. The equal-area method is perhaps easier to implement, because it does not require nonuniform measurement grid spacing and generally specifies fewer measurement locations. Therefore, it seems reasonable to first assess the volumetric flow rate with equal area method and subsequently reduce the indicated result by approximately 3 to 4%, thereby achieving a good approximation of the log-T traverse measurement. This may be a reasonable compromise for those who do not wish to use the log-T method (Hickman et al. 2012). When using the log-T method for a rectangular duct traverse, measure a minimum of 25 points. For a circular duct traverse, the log-linear rule and three symmetrically disposed diameters may be used (Figure 8). Points on two perpendicular diameters may be used where access is limited. If possible, measuring points should be located at least 7.5 hydraulic diameters downstream and 3 hydraulic diameters upstream
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Fig. 8 Measuring Points for Rectangular and Round Duct Traverse from a disturbance (e.g., caused by a turn). However, for common single-path rectangular duct fitting disturbances (60° and 90° transitions, 90° elbows), measurements can be made to uncertainties within about ±3 to 4% for traverses even as close as 1 to 2 equivalent diameters downstream of the disturbance using the log-T traverse method. Furthermore, similar results can be obtained in single-path rectangular ducts with these types disturbances using either a pitot-static probe or a hot-wire anemometer (Hickman et al. 2012). Because field-measured airflows are rarely steady and uniform, particularly near disturbances, accuracy can be improved by increasing the number of measuring points. Straightening vanes (ASHRAE Standard 51) located 1.5 duct diameters ahead of the traverse plane improve measurement precision. When velocities at a traverse plane fluctuate, the readings should be averaged on a time-weighted basis. Two traverse readings in short succession also help to average out velocity variations that occur with time. If negative velocity pressure readings are encountered, this is an indication that highly nonuniform flows are present. From the characteristics of the pitot-static probe yaw variation shown in Figure 7, it is not possible to draw meaningful and reliable conclusions from the measurements, particularly downstream of tee fitting disturbances, where boundary layer separation occurs (which causes flow reversal) in the branch region downstream of the tee. Even if no actual reversal occurs, the flow may also be highly nonaxial, and the flow directional limitations of the pitot-static probe, as shown in Figure 7B, may still result in meaningless results. Also, it is important to note that, although the pitot-static probe can produce a negative and potentially meaningless output, it does indicate obvious flow uniformity problems. Important Note: negative velocity pressures measured by a pitot-static tube indicate an unacceptable traverse location. To achieve meaningful volumetric flow rate measurements,
traverses must be performed where no negative velocity pressure values occur. The presence of negative velocity pressures (even when those values are considered to be zero-velocity values when summing and averaging) results in a completely meaningless duct flow calculation. A hot-wire anemometer cannot indicate flow direction. Consequently, it always indicates a positive velocity, even under reverse flow. Hence, use of a thermal anemometer probe wherever flow reversals occur, such as those encountered in the downstream branch of tee fittings, can result in large errors. For traverse measurements in the downstream branch of tees, regions of apparent negative velocity pressure have been encountered as far as 7.5 equivalent diameters downstream of the tee, and are more likely to occur at high flow rates with relatively low relative branch flows (Hickman et al. 2012). These regions should be avoided when using thermal anemometers. Example 2. Step 1. Numerical Evaluation of Duct Average Velocity. Velocity measurements for a 610 × 610 mm square duct traverse using the log-T method are given in the following tables. Air density is = 1.185 kg/m3. Air temperature and absolute pressure conditions in the duct are 30.5°C and 103.3 kPa. The top row shows the horizontal traverse point position in the duct cross section, and the left column gives the vertical traverse point position (mm) in the duct cross-section. Note that, for this duct, there are 25 measurement positions across the duct, with 5 in each direction (see Figure 8 for how these positions are determined). Velocity Measurements, m/s 46 mm 175 mm
46 mm
175 mm
305 mm
434 mm
564 mm
5.644 6.142
5.878 6.330
5.578 5.949
6.253 6.452
5.761 6.325
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37.20
2017 ASHRAE Handbook—Fundamentals (SI) 6.182 6.147 5.796
305 mm 434 mm 564 mm
6.980 6.711 5.781
6.655 6.843 5.690
6.965 6.985 5.832
6.360 5.959 5.928
Velocity Pressure Measurements, Pa 46 mm
175 mm
305 mm
434 mm
564 mm
18.87 22.35 22.65 22.39 19.91
20.47 23.74 28.87 26.68 19.80
18.43 20.97 26.24 27.74 19.18
23.17 24.66 28.74 28.91 20.15
19.66 23.70 23.97 21.04 20.82
46 mm 175 mm 305 mm 434 mm 564 mm
The average air velocity is then V1 + V2 + V3 + V4 + + VN 1 N Vave = ---- V i = -------------------------------------------------------------------------- = 6.203 m/s N N i=1
Alternatively, if local measurements are made in terms of velocity pressure pw, the average velocity is
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1 N 2 1 N V ave = ---- V i = C --- ---- p w ,i N N i=1 i=1 =
2 1 25 ------------- ------ 18.87 + 20.47 1.185 25
+ 18.43 + = 6.203 m/s i=1
where the term in brackets represents the average of the square root of the individual velocity pressure measurements. Step 2: Numerical Evaluation of Duct Volumetric Flow Rate. For the given duct cross-sectional area, the volumetric flow rate of air in cubic metres per second is then QActual = Vave A =(6.203 m/s)(0.610 0.610 m) = 2.308 m3/s where A = 0.610 × 0.610 m is the duct cross-sectional area. The preceding actual volumetric flow rates can be converted to standard volumetric flow rates by referencing the flow rates to standard air density conditions for the same mass flow rate. The standard volumetric flow rate is the flow rate that would exist of the air were at standard air density conditions. Thus, P Actual T S tan dard - ---------------------- QStandard = QActual -------------------- P S tan dard T Actual
If standard conditions are defined as 21.1°C and 101.4 kPa, then the standard volumetric flow rate is 103.3 kPa 21.1 + 273.16 - --------------------------------- = 2.278 m/s QStandard = (2.308 m/s) ---------------------- 101.4 kPa 30.5 + 273.16
Note: Different manufacturers can use different values for standard air density and standard conditions. It is very important to use a consistent set of standard conditions when comparing flow rates.
6.5
AIRFLOW-MEASURING HOODS
Flow-measuring hoods are portable instruments designed to measure supply or exhaust airflow through diffusers and grilles in HVAC systems. The assembly typically consists of a fabric hood section, a plastic or metal base, an airflow-measuring manifold, a meter, and handles for carrying and holding the hood in place. For volumetric airflow measurements, the flow-measuring hood is placed over a diffuser or grille. The fabric hood captures and directs airflow from the outlet or inlet across the flow-sensing manifold in the base of the instrument. The manifold consists of a number of tubes containing upstream and downstream holes in a grid, designed to simultaneously sense and average multiple
velocity points across the base of the hood. Air from the upstream holes flows through the tubes past a sensor and then exits through the downstream holes. Sensors used by different manufacturers include swinging vane anemometers, electronic micromanometers, and thermal anemometers. In electronic micromanometers, air does not actually flow through the manifold, but the airtight sensor senses the pressure differential from the upstream to downstream series of holes. The meter on the base of the hood interprets the signal from the sensor and provides a direct reading of volumetric flow in either an analog or digital display format. As a performance check in the field, the indicated flow of a measuring hood can be compared to a duct traverse flow measurement (using a pitot-tube or thermal anemometer). All flow-measuring hoods induce some back pressure on the air-handling system because the hood restricts flow out of the diffuser. This added resistance alters the true amount of air coming out of the diffuser. In most cases, this error is negligible and is less than the accuracy of the instrument. For proportional balancing, this error need not be taken into account because all similar diffusers have about the same amount of back pressure. To determine whether back pressure is significant, a velocity traverse can be made in the duct ahead of the diffuser with and without the hood in place. The difference in average velocity of the traverse indicates the degree of back-pressure compensation required on similar diffusers in the system. For example, if the average velocity is 4.0 m/s with the hood in place and 4.1 m/s without the hood, the indicated flow reading can be multiplied by 1.025 on similar diffusers in the system (4.1/4.0 = 1.025). As an alternative, the designer of the air-handling system can predict the head-induced airflow reduction by using a curve supplied by the hood manufacturer. This curve indicates the pressure drop through the hood for different flow rates.
7.
FLOW RATE MEASUREMENT
Various means of measuring fluid flow rate are listed in Table 5. Values for volumetric or mass flow rate measurement (ASME Standard PTC 19.5; Benedict 1984) are often determined by measuring pressure difference across an orifice, nozzle, or venturi tube. The various meters have different advantages and disadvantages. For example, the orifice plate is more easily changed than the complete nozzle or venturi tube assembly. However, the nozzle is often preferred to the orifice because its discharge coefficient is more precise. The venturi tube is a nozzle followed by an expanding recovery section to reduce net pressure loss. Differential pressure flow measurement has benefited through workshops addressing fundamental issues, textbooks, research, and improved standards (ASME Standards B40.100, MFC-1M, MFC-9M, MFC-10M; DeCarlo 1984; ISO Standards 5167:2003, 5801:2007; Mattingly 1984; Miller 1983). Fluid meters use a wide variety of physical techniques to measure flow (ASME Standard PTC 19.5; DeCarlo 1984; Miller 1983); more common ones are described in this section. To validate accuracy of flow rate measurement instruments, calibration procedures should include documentation of traceability to the calibration facility. The calibration facility should, in turn, provide documentation of traceability to national standards.
Flow Measurement Methods Direct. Both gas and liquid flow can be measured accurately by timing a collected amount of fluid that is measured gravimetrically or volumetrically. This method is common for calibrating other metering devices, but it is particularly useful where flow rate is low or intermittent and where a high degree of accuracy is required. These systems are generally large and slow, but in their simplicity, they can be considered primary devices.
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Table 5 Volumetric or Mass Flow Rate Measurement Measurement Means
Application
Nominal Range
Orifice and differential pressure measurement system
Flow through pipes, ducts, and plenums for all fluids
Above Reynolds number of 5000
1 to 5%
Nozzle and differential pressure measurement system
Flow through pipes, ducts, and plenums for all fluids
Above Reynolds number of 5000
0.5 to 2.0%
Venturi tube and differential pressure measurement system
Flow through pipes, ducts, and plenums for all fluids
Above Reynolds number of 5000
0.5 to 2.0%
Timing given mass or volumetric flow Rotameters
Liquids or gases; used to calibrate other flowmeters Liquids or gases
Any
0.1 to 0.5%
Any
0.5 to 5.0%
Coriolis
Mass or volume; variable-density liquids or gases Relatively small volumetric flow with high pressure loss
As high as 907 kg/s
0.05 to 1.5%
As high as 500 L/s, depending on type
0.1 to 2.0% depending on type 0.5 to 1.0%
Most types require calibration with fluid being metered.
1%
Uniform velocity; usually used with gases.
Displacement meter
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Gasometer or volume displacement Short-duration tests; used to calibrate other flowmeters Thomas meter (temperature rise of stream caused by electrical heating) Element of resistance to flow and differential pressure measurement system Turbine flowmeters Single- or multipoint instrument for measuring velocity at specific point in flow Heat input and temperature changes with steam and water coil Laminar flow element and differential pressure measurement system Magnetohydrodynamic flowmeter (electromagnetic) Swirl flowmeter and vortex shedding meter
Elaborate setup justified by need for good accuracy Used for check where system has calibrated resistance element
Precision
Total flow limited by available volume of containers Any Lower limit set by readable pressure drop
1 to 5%
Liquids or gases Any 0.25 to 2.0% Primarily for installed air-handling Lower limit set by accuracy 2 to 10% systems with no special proviof velocity measurement sion for flow measurement instrumentation Check value in heater or cooler tests
Any
1 to 3%
Measure liquid or gas volumetric 50 mm3/s to 1 m3/s flow rate; nearly linear relationship with pressure drop; simple and easy to use Measures electrically conductive 0.006 to 600 L/s fluids, slurries; meter does not obstruct flow; no moving parts Measure liquid or gas flow in pipe; Above Reynolds number no moving parts of 104
The variable-area meter or rotameter is a convenient directreading flowmeter for liquids and gases. This is a vertical, tapered tube in which the flow rate is indicated by the position of a float suspended in the upward flow. The float’s position is determined by its buoyancy and the upward fluid drag. Displacement meters measure total liquid or gas flow over time. The two major types of displacement meters used for gases are the conventional gas meter, which uses a set of bellows, and the wet test meter, which uses a water displacement principle. Indirect. The Thomas meter is used in laboratories to measure high gas flow rates with low pressure losses. Gas is heated by electric heaters, and the temperature rise is measured by two resistance thermometer grids. When heat input and temperature rise are known, the mass flow of gas is calculated as the quantity of gas that removes the equivalent heat at the same temperature rise. A velocity traverse (made using a pitot tube or other velocitymeasuring instrument) measures airflow rates in the field or calibrates large nozzles. This method can be imprecise at low velocities and impracticable where many test runs are in progress. Another field-estimating method measures pressure drop across elements with known pressure drop characteristics, such as heating
Limitations Discharge coefficient and accuracy influenced by installation conditions. Discharge coefficient and accuracy influenced by installation conditions. Discharge coefficient and accuracy influenced by installation conditions. System is bulky and slow. Should be calibrated for fluid being metered.
—
Secondary reading depends on accuracy of calibration. Uses electronic readout. Accuracy depends on uniformity of flow and completeness of traverse. May be affected by disturbances near point of measurement. —
1%
Fluid must be free of dirt, oil, and other impurities that could plug meter or affect its calibration.
1%
At present state of the art, conductivity of fluid must be greater than 5 mho/cm. —
1%
and cooling coils or fans. If the pressure drop/flow rate relationship has been calibrated against a known reference (typically, at least four points in the operating range), the results can be precise. If the method depends on rating data, it should be used for check purposes only.
7.1
VENTURI, NOZZLE, AND ORIFICE FLOWMETERS
Flow in a pipeline can be measured by a venturi meter (Figure 9), flow nozzle (Figure 10), or orifice plate (Figure 11). American Society of Mechanical Engineers (ASME) Standard MFC-3M describes measurement of fluid flow in pipes using the orifice, nozzle, and venturi; ASME Standard PTC 19.5 specifies their construction. Assuming an incompressible fluid (liquid or slow-moving gas), uniform velocity profile, frictionless flow, and no gravitational effects, the principle of conservation of mass and energy can be applied to the venturi and nozzle geometries to give 2 p 1 – p 2 w = V1A1 = V2A2 = A2 -----------------------------1 – 4
(7)
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where w V A p
= = = = = =
mass flow rate, kg/s velocity of stream, m/s flow area, m2 density of fluid, kg/m3 absolute pressure, Pa ratio of diameters D2/D1 for venturi and sharp-edge orifice and d/ D for flow nozzle, where D = pipe diameter and d = throat diameter
Note: Subscript 1 refers to entering conditions; subscript 2 refers to throat conditions.
Because flow through the meter is not frictionless, a correction factor C is defined to account for friction losses. If the fluid is at a high temperature, an additional correction factor Fa should be included to account for thermal expansion of the primary element. Because this amounts to less than 1% at 260°C, it can usually be omitted. Equation (7) then becomes
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2 p 1 – p 2 w = CA2 -----------------------------1 – 4
(8)
where C is the friction loss correction factor. The factor C is a function of geometry and Reynolds number. Values of C are given in ASME Standard PTC 19.5. The jet passing through an orifice plate contracts to a minimum area at the vena
contracta located a short distance downstream from the orifice plate. The contraction coefficient, friction loss coefficient C, and approach factor 1/(1 4) 0.5 can be combined into a single constant K, which is a function of geometry and Reynolds number. The orifice flow rate equations then become Q = KA2
Fig. 10
(9)
where Q = discharge flow rate, m3/s A2 = orifice area, m2 p1 p2 = pressure drop as obtained by pressure taps, Pa
Values of K are shown in ASME Standard PTC 19.5. Valves, bends, and fittings upstream from the flowmeter can cause errors. Long, straight pipes should be installed upstream and downstream from flow devices to ensure fully developed flow for proper measurement. ASHRAE Standard 41.8 specifies upstream and downstream pipe lengths for measuring flow of liquids with an orifice plate. ASME Standard PTC 19.5 gives piping requirements between various fittings and valves and the venturi, nozzle, and orifice. If these conditions cannot be met, flow conditioners or straightening vanes can be used (ASME Standards PTC 19.5, MFC-10M; Mattingly 1984; Miller 1983). Compressibility effects must be considered for gas flow if pressure drop across the measuring device is more than a few percent of the initial pressure. Nozzles are sometimes arranged in parallel pipes from a common manifold; thus, the capacity of the testing equipment can be changed by shutting off the flow through one or more nozzles. An apparatus designed for testing airflow and capacity of airconditioning equipment is described by Wile (1947), who also presents pertinent information on nozzle discharge coefficients, Reynolds numbers, and resistance of perforated plates. Some laboratories refer to this apparatus as a code tester.
7.2
Fig. 9 Typical Herschel-Type Venturi Meter
2 p1 – p2 -------------------------
VARIABLE-AREA FLOWMETERS (ROTAMETERS)
In permanent installations where high precision, ruggedness, and operational ease are important, the variable-area flowmeter is satisfactory. It is frequently used to measure liquids or gases in smalldiameter pipes. For ducts or pipes over 150 mm in diameter, the expense of this meter may not be warranted. In larger systems,
Dimensions of ASME Long-Radius Flow Nozzles
From ASME PTC 19.5. Reprinted with permission of ASME.
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37.23
Fig. 11 Sharp-Edge Orifice with Pressure Tap Locations From ASME PTC 19.5. Reprinted with permission of ASME.
Fig. 12 Variable-Area Flowmeter however, the meter can be placed in a bypass line and used with an orifice. The variable-area meter (Figure 12) commonly consists of a float that is free to move vertically in a transparent tapered tube. The fluid to be metered enters at the narrow bottom end of the tube and moves upward, passing at some point through the annulus formed between the float and the inside wall of the tube. At any particular flow rate, the float assumes a definite position in the tube; a calibrated scale on the tube shows the float’s location and the fluid flow rate. The float’s position is established by a balance between the fluid pressure forces across the annulus and gravity on the float. The buoyant force Vf (f –)g supporting the float is balanced by the pressure difference acting on the cross-sectional area of the float Af p, where f , Af , and Vf are, respectively, the float density, float cross-sectional area, and float volume. The pressure difference across the annulus is V f f – g p = -----------------------------Af
(10)
7.4 POSITIVE-DISPLACEMENT METERS
The mass flow follows from Equation (9) as w = KA2
2V f f – g ------------------------------------Af
(11)
Flow for any fluid is nearly proportional to the area, so that calibration of the tube is convenient. To use the meter for different fluids, the flow coefficient variation for any float must be known. Float design can reduce variation of the flow coefficient with Reynolds number; float materials can reduce the dependence of mass flow calibration on fluid density.
7.3
sensor tube vibrates the tube in cantilever motion at a known frequency. The liquid enters the vibrating tube and is given the vertical momentum of the tube. The liquid in the entry portion of the sensor tube resists in the downward direction when the tube is moving upward. Conversely, when the tube is moving downward, the liquid in the exit portion of the sensor tube resists in the upward direction. Combined, these effects create a symmetrical twist angle. According to Newton’s second law of motion, the amount of sensor tube twist angle is directly proportional to the mass flow rate of liquid flowing through the tube. Electromagnetic velocity sensors on opposing sides of the sensor tube measure the velocity of the vibrating tube. Mass flow rate is determined by measuring the time difference in the velocity measurements: the greater the time difference, the greater the mass flow rate. The measuring tubes are vibrated at their natural frequency; a change of the fluid mass inside the tubes causes a corresponding change to the tube’s natural frequency. The frequency change of the tube is used to calculate the fluid density.
CORIOLIS PRINCIPLE FLOWMETERS
Coriolis liquid flowmeters directly measure liquid mass flow rates. In a Coriolis flowmeter, the liquid flows through a vibrating sensor tube within the meter. An electromagnetic coil located on the
Many positive-displacement meters are available for measuring total liquid or gas volumetric flow rates. The measured fluid flows progressively into compartments of definite size. As the compartments fill, they rotate so that the fluid discharges from the meter. The flow rate through the meter equals the product of the compartment volume, number of compartments, and rotation rate of the rotor. Most of these meters have a mechanical register calibrated to show total flow.
7.5
TURBINE FLOWMETERS
Turbine flowmeters are volumetric flow-rate-sensing meters with a magnetic stainless steel turbine rotor suspended in the flow stream of a nonmagnetic meter body. The fluid stream exerts a force on the blades of the turbine rotor, setting it in motion and converting the fluid’s linear velocity to an angular velocity. Design motivation for turbine meters is to have the rotational speed of the turbine
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proportional to the average fluid velocity and thus to the volume rate of fluid flow (DeCarlo 1984; Mattingly 1992; Miller 1983). The rotor’s rotational speed is monitored by an externally mounted pickoff assembly. The magnetic pickoff contains a permanent magnet and coil. As the turbine rotor blades pass through the field produced by the permanent magnet, a shunting action induces AC voltage in the winding of the coil wrapped around the magnet. A sine wave with a frequency proportional to the flow rate develops. With the radio frequency pickoff, an oscillator applies a highfrequency carrier signal to a coil in the pickoff assembly. The rotor blades pass through the field generated by the coil and modulate the carrier signal by shunting action on the field shape. The carrier signal is modulated at a rate corresponding to the rotor speed, which is proportional to the flow rate. With both pickoffs, pulse frequency is a measure of flow rate, and the total number of pulses measures total volume (Mattingly 1992; Shafer 1961; Woodring 1969). Because output frequency of the turbine flowmeter is proportional to flow rate, every pulse from the turbine meter is equivalent to a known volume of fluid that has passed through the meter; the sum of these pulses yields total volumetric flow. Summation is done by electronic counters designed for use with turbine flowmeters; they combine a mechanical or electronic register with the basic electronic counter. Turbine flowmeters should be installed with straight lengths of pipe upstream and downstream from the meter. The length of the inlet and outlet pipes should be according to manufacturers’ recommendations or pertinent standards. Where recommendations of standards cannot be accommodated, the meter installation should be calibrated. Some turbine flowmeters can be used in bidirectional flow applications. A fluid strainer, used with liquids of poor or marginal lubricity, minimizes bearing wear. The lubricity of the process fluid and the type and quality of rotor bearings determine whether the meter is satisfactory for the particular application. When choosing turbine flowmeters for use with fluorocarbon refrigerants, pay attention to the type of bearings used in the meter and to the refrigerant’s oil content. For these applications, sleeve-type rather than standard ball bearings are recommended. The amount of oil in the refrigerant can severely affect calibration and bearing life. In metering liquid fluorocarbon refrigerants, the liquid must not flash to a vapor (cavitate), which tremendously increases flow volume. Flashing results in erroneous measurements and rotor speeds that can damage bearings or cause a failure. Flashing can be avoided by maintaining adequate back pressure on the downstream side of the meter (Liptak 1972).
7.6
ELECTROMAGNETIC (MAG) FLOWMETERS
Magnetic flowmeters operate on the principle of Faraday’s law of induction that states that the electromotive force induced in a circuit equals the negative of the time rate of change of the magnetic flux through the circuit. In a magnetic flowmeter, a magnetic field is electrically generated and channeled into the liquid flowing through the pipe. Faraday’s law states that the voltage generated is proportional to the movement of the flowing liquid. Electronics in a magnetic flowmeter sense voltage and determine the volumetric flow rate. Magnetic tube flowmeters have no flow obstructions, so the pressure loss in these flowmeters is less than for many other types of flowmeters. Magnetic flowmeters require a minimum electrical conductivity. Most HVAC&R liquids have enough electrical conductivity to be used with these flowmeters.
7.7
VORTEX-SHEDDING FLOWMETERS
Vortex-shedding flowmeters are used to determine liquid velocities. Piezoelectric methods, strain-gage methods, or hot-film methods are used to sense dynamic pressure variations created by vortex
shedding. The operating principle for these flowmeters is based on vortex shedding that occurs downstream of an immersed bluntshaped solid body. As the liquid stream passes a blunt-shaped body, the liquid separates and generates small vortices that are shed alternately along and downstream of each side of the blunt-shaped body. Each vortex-shedding meter is designed to have a constant Strouhal number so that the vortex shedding frequency is proportional to the liquid flow velocity over a specified flow velocity range. The Strouhal number for each vortex shedding meter is experimentally determined by the flowmeter manufacturer and is provided with each vortex-shedding meter.
8. AIR INFILTRATION, AIRTIGHTNESS, AND OUTDOOR AIR VENTILATION RATE MEASUREMENT Air infiltration is the flow of outdoor air into a building through unintentional openings. Airtightness refers to the building envelope’s ability to withstand flow when subjected to a pressure differential. The outdoor air ventilation rate is the rate of outdoor airflow intentionally introduced to the building for dilution of occupant- and building-generated contaminants. Measurement approaches to determine these factors are described briefly here, and in greater detail in Chapter 16. Air infiltration depends on the building envelope’s airtightness and the pressure differentials across the envelope. These differentials are induced by wind, stack effect, and operation of building mechanical equipment. For meaningful results, the air infiltration rate should be measured under typical conditions. Airtightness of a residential building’s envelope can be measured relatively quickly using building pressurization tests. In this technique, a large fan or blower mounted in a door or window induces a large and roughly uniform pressure difference across the building shell. The airflow required to maintain this pressure difference is then measured. The more leakage in the building, the more airflow is required to induce a specific indoor/outdoor pressure difference. Building airtightness is characterized by the airflow rate at a reference pressure, normalized by the building volume or surface area. Under proper test conditions, results of a pressurization test are independent of weather conditions. Instrumentation requirements for pressurization testing include air-moving equipment, a device to measure airflow, and a differential pressure gage. Commercial building envelope leakage can also be measured using building pressurization tests. Bahnfleth et al. (1999) describe a protocol for testing envelope leakage of tall buildings using the building’s air-handling equipment. Outdoor airflow can be measured directly using the flow rate measurement techniques described in this chapter. The flow measurement equipment must be appropriate for the required accuracy, operating conditions, range of airflows, and temperatures expected. The outdoor airflow rate is normally measured during testing and balancing, during commissioning, or for continuous ventilation flow rate control using permanently mounted flow sensors. An additional factor that may be of interest is the building’s air exchange rate, which compares airflow into the building with the building’s volume. Typically, this includes both mechanical ventilation and infiltration. Building air exchange rates can be measured by injecting a tracer gas (ideally, a chemically stable, nontoxic gas not normally present in buildings) into a building and monitoring and analyzing the tracer gas concentration response. Equipment required for tracer testing includes (1) a means of injecting the tracer gas and (2) a tracer gas concentration measurement device, such as a gas chromatograph. Various tracer gas techniques are used, distinguished by their injection strategy and analysis approach. These techniques include constant concentration (equilibrium tracer),
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Carbon Dioxide Carbon dioxide is often used as a tracer gas because CO2 gas monitors are relatively inexpensive and easy to use, and occupantgenerated CO2 can be used for most tracer gas techniques. Bottled CO2 or CO2 fire extinguishers are also readily available for tracer gas injection. Carbon dioxide may be used as a tracer gas to measure ventilation rates under the conditions and methods described in ASTM Standard D6245, for diagnostic purposes and point-intime snapshots of the system’s ventilation capabilities. CO2 sensors are also used in building controls strategies to avoid overventilation by approximating the level of occupancy in a space; this is one method of demand-controlled ventilation. The concentration output may be used in a mathematical formula that allows the system to modulate ventilation rates when spaces with high density have highly variable or intermittent occupancy (e.g., churches, theaters, gymnasiums). This method of control is less effective in lowerdensity occupancies and spaces with more stable populations (Persily and Emmerich 2001). Carbon dioxide may also be used together with outdoor air intake rate data to estimate the current population of a space. CO2 input for ventilation control does not address contaminants generated by the building itself, and therefore cannot be used without providing a base level of ventilation for non-occupant-generated contaminants that have been shown to total a significant fraction if not a majority of those found in the space.
9. CARBON DIOXIDE MEASUREMENT Carbon dioxide has become an important measurement parameter for HVAC&R engineers, particularly in indoor air quality (IAQ) applications. Although CO2 is generally not of concern as a specific toxin in indoor air, it is used as a surrogate indicator of odor related to human occupancy. ANSI/ASHRAE Standard 62.1 recommends specific minimum outdoor air ventilation rates to ensure adequate indoor air quality.
37.25 calibrated between 0 and 5000 ppm, are typically accurate within 150 ppm, but the accuracy of some sensors can be improved to within 50 ppm if the instrument is calibrated for a narrower range. Portable NDIR meters are available with direct-reading digital displays; however, response time varies significantly among different instruments. Most NDIR cell designs facilitate very rapid CO2 sample diffusion, although some instruments now in widespread use respond more slowly, resulting in stabilization times greater than 5 min (up to 15 min), which may complicate walk-through inspections. CO2 instruments can be classified as either single or dual wavelength. Single-wavelength instruments contain a single spectral filter at the absorption band of CO2. This configuration is prone to drift as the infrared source ages and loses intensity. These instruments often use a self-adjustment procedure to compensate and require periodic exposure to outdoor CO2 levels of approximately 400 ppm during periods when a building is unoccupied. This makes it impractical for buildings such as hospitals that are always occupied. Dual-wavelength instruments contain two spectral filters: one at the absorption band of CO2 and one at a wavelength unaffected by any gases. This second filter allows the instrument to monitor the intensity of the infrared source to directly compensate for source aging.
Calibration In a clean, stable environment, NDIR sensors can hold calibration for months, but condensation, dust, dirt, and mechanical shock may offset calibration. As with all other CO2 sensor technologies, NDIR sensor readings are proportional to pressure, because the density of gas molecules changes when the sample pressure changes. This leads to errors in CO2 readings when the barometric pressure changes from the calibration pressure. Weather-induced errors are small but should be considered when using CO2 instruments at an altitude that is significantly different form the calibration altitude. Some instruments can accept a user adjustment to compensate for altitude. Instruments that do not have this ability should be recalibrated. Some NDIR sensors are sensitive to cooling effects when placed in an airstream. This is an important consideration when locating a fixed sensor or when using a portable system to evaluate airhandling system performance, because airflow in supply and return ducts may significantly shift readings.
Applications 9.1
NONDISPERSIVE INFRARED CO2 DETECTORS
The most widespread technology for IAQ applications is the nondispersive infrared (NDIR) sensor (Figure 13). This device uses the strong absorption band that CO2 produces at 4.2 m when excited by an infrared light source. IAQ-specific NDIR instruments,
Nondispersive infrared sensors are well suited for equilibrium tracer and tracer decay ventilation studies, and faster-response models are ideal for a quick, basic evaluation of human-generated pollution and ventilation adequacy. When properly located, these sensors are also appropriate for continuous monitoring and for control strategies using equilibrium tracer and air fraction tracer calculations.
9.2
Fig. 13
Nondispersive Infrared Carbon Dioxide Sensor
AMPEROMETRIC ELECTROCHEMICAL CO2 DETECTORS
Amperometric electrochemical CO2 sensors (Figure 14) use a measured current driven between two electrodes by the reduction of CO2 that diffuses across a porous membrane. Unlike NDIR sensors, which normally last the lifetime of the instrument, electrochemical CO2 sensors may change in electrolyte chemistry over time (typically 12 to 18 months) and should be replaced periodically. These sensors typically hold their calibration for several weeks, but they may drift more if exposed to low humidity; this drift makes them less suitable for continuous monitoring applications. At low humidity (below 30% rh), the sensors must be kept moist to maintain specified accuracy. Amperometric electrochemical sensors require less power than NDIR sensors, usually operating continuously for weeks where NDIR
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instruments typically operate for 6 h (older models) to 150 h (newer models). The longer battery life can be advantageous for spot checks and walk-throughs, and for measuring CO2 distribution throughout a building and within a zone. Unlike most NDIR sensors, amperometric electrochemical sensors are not affected by high humidity, although readings may be affected if condensate is allowed to form on the sensor.
9.3
PHOTOACOUSTIC CO2 DETECTORS
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Open-Cell Sensors Open-cell photoacoustic CO2 sensors (Figure 15) operate as air diffuses through a permeable membrane into a chamber that is pulsed with filtered light at the characteristic CO2 absorption frequency of 4.2 m. The light energy absorbed by the CO2 heats the sample chamber, causing a pressure pulse, which is sensed by a piezoresistor. Open-cell photoacoustic CO2 sensors are presently unavailable in portable instruments, in part because any vibration during transportation would affect calibration and might affect the signal obtained for a given concentration of CO2. Ambient acoustical noise may also influence readings. For continuous monitoring, vibration is a concern, as are temperature and airflow cooling effects. However, if a sensor is located properly and the optical filter is kept relatively clean, photoacoustic CO2 sensors may be very stable. Commercially available open-cell photoacoustic transmitters do not allow recalibration to adjust for pressure differences, so an offset should be incorporated in any control system using these sensors at an altitude or duct pressure other than calibration conditions.
Closed-Cell Sensors Closed-cell photoacoustic sensors (Figure 16) operate under the same principle as the open-cell version, except that samples are pumped into a sample chamber that is sealed and environmentally stabilized. Two acoustic sensors are sometimes used in the chamber to minimize vibration effects. Closed-cell units, available as portable or fixed monitors, come with particle filters that are easily replaced (typically at 3- to 6-month intervals) if dirt or dust accumulates on
Fig. 14
Fig. 15
them. Closed-cell photoacoustic monitors allow recalibration to correct for drift, pressure effects, or other environmental factors that might influence accuracy.
9.4
POTENTIOMETRIC ELECTROCHEMICAL CO2 DETECTORS
Potentiometric electrochemical CO2 sensors use a porous fluorocarbon membrane that is permeable to CO2, which diffuses into a carbonic acid electrolyte, changing the electrolyte’s pH. This change is monitored by a pH electrode inside the cell. The pH electrode isopotential drift prohibits long-term monitoring to the accuracy and resolution required for continuous measurement or control or for detailed IAQ evaluations, although accuracy within 100 ppm, achievable short-term over the 2000 ppm range, may be adequate for basic ventilation and odor evaluations. In addition, this type of sensor has a slow response, which increases the operator time necessary for field applications or for performing a walk-through of a building.
9.5
COLORIMETRIC DETECTOR TUBES
Colorimetric detector tubes contain a chemical compound that discolors in the presence of CO2 gas, with the amount of discoloration related to the CO2 concentration. These detector tubes are often used to spot-check CO2 levels; when used properly, they are accurate to within 25%. If numerous samples are taken (i.e., six or more), uncertainty may be reduced. However, CO2 detector tubes are generally not appropriate for specific ventilation assessment because of their inaccuracy and inability to record concentration changes over time.
9.6
LABORATORY MEASUREMENTS
Laboratory techniques for measuring CO2 concentration include mass spectroscopy, thermal conductivity, infrared spectroscopy, and gas chromatography. These techniques typically require taking onsite grab samples for laboratory analysis. Capital costs for each piece of equipment are high, and significant training is required. A considerable drawback to grab sampling is that CO2 levels change significantly during the day and over the course of a week, making it sensible to place sensors on site with an instrument capable of recording or data logging measurements continuously over the course of a workweek. An automated grab sampling system capturing many samples of data would be quite cumbersome and expensive if designed to provide CO2 trend information over time. However, an advantage to laboratory techniques is that they can be highly accurate. A mass spectrometer, for example, can measure CO2 concentration to within 5 ppm from 0 to 2000 ppm. All laboratory measurement techniques are subject to errors resulting from
Amperometric Carbon Dioxide Sensor
Open-Cell Photoacoustic Carbon Dioxide Sensor
Fig. 16
Closed-Cell Photoacoustic Carbon Dioxide Sensor
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37.27
interfering agents. A gas chromatograph is typically used in conjunction with the mass spectrometer to eliminate interference from nitrous oxide (N2O), which has an equivalent mass, if samples are collected in a hospital or in another location where N2O might be present.
10.
ELECTRIC MEASUREMENT
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Ammeters Ammeters are low-resistance instruments for measuring current. They should be connected in series with the circuit being measured (Figure 17). Ideally, they have the appearance of a short circuit, but in practice, all ammeters have a nonzero input impedance that influences the measurement to some extent. Ammeters often have several ranges, and it is good practice when measuring unknown currents to start with the highest range and then reduce the range to the appropriate value to obtain the most sensitive reading. Ammeters with range switches maintain circuit continuity during switching. On some older instruments, it may be necessary to short-circuit the ammeter terminals when changing the range. Current transformers are often used to increase the operating range of ammeters. They may also provide isolation/protection from a high-voltage line. Current transformers have at least two separate windings on a magnetic core (Figure 18). The primary winding is connected in series with the circuit in which the current is measured. In a clamp-on probe, the transformer core is actually opened and then connected around a single conductor carrying the current to be measured. That conductor serves as the primary winding. The secondary winding carries a scaled-down version of the primary current, which is connected to an ammeter. Depending on instrument type, the ammeter reading may need to be multiplied by the ratio of the transformer.
Fig. 17 Ammeter Connected in Power Circuit
Fig. 20
Voltmeter with Potential Transformer
Fig. 18
When using an auxiliary current transformer, the secondary circuit must not be open when current is flowing in the primary winding; dangerously high voltage may exist across the secondary terminals. A short-circuiting blade between the secondary terminals should be closed before the secondary circuit is opened at any point. Transformer accuracy can be impaired by residual magnetism in the core when the primary circuit is opened at an instant when flux is large. The transformer core may be left magnetized, resulting in ratio and phase angle errors. The primary and secondary windings should be short-circuited before making changes.
Voltmeters Voltmeters are high-resistance instruments that should be connected across the load (in parallel), as shown in Figure 19. Ideally, they have the appearance of an open circuit, but in practice, all voltmeters have some finite impedance that influences measurement to some extent. Voltage transformers are often used to increase the operating range of a voltmeter (Figure 20). They also provide isolation from high voltages and prevent operator injury. Like current transformers, voltage transformers consist of two or more windings on a magnetic core. The primary winding is generally connected across the high voltage to be measured, and the secondary winding is connected to the voltmeter. It is important not to short-circuit the secondary winding of a voltage transformer.
Wattmeters Wattmeters measure the active power of an AC circuit, which equals the voltage multiplied by that part of the current in phase with the voltage. There are generally two sets of terminals: one to connect the load voltage and the other to connect in series with the load current. Current and voltage transformers can be used to extend the range of a wattmeter or to isolate it from high voltage. Figures 21 and
Ammeter with Current Transformer
Fig. 21 Wattmeter in Single-Phase Circuit Measuring Power Load plus Loss in Current-Coil Circuit
Fig. 19 Voltmeter Connected Across Load
Fig. 22 Wattmeter in Single-Phase Circuit Measuring Power Load plus Loss in Potential-Coil Circuit
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Fig. 23
Wattmeter with Current and Potential Transformer
Fig. 24 Polyphase Wattmeter in TwoPhase, Three-Wire Circuit with Balanced or Unbalanced Voltage or Load
Fig. 25 Polyphase Wattmeter in Three-Phase, Three-Wire Circuit
Fig. 26 Single-Phase Power-Factor Meter
Fig. 27 Three-Wire, Three-Phase Power-Factor Meter
22 show connections for single-phase wattmeters, and Figure 23 shows use of current and voltage transformers with a single-phase wattmeter. Wattmeters with multiple current and voltage elements are available to measure polyphase power. Polyphase wattmeter connections are shown in Figures 24 and 25.
marks on the shaft, such as a bar on one side and a circle on the opposite side. If, for example, the two are seen superimposed, then the strobe light is flashing at an even multiple of the true rpm.
Power-Factor Meters Power-factor meters measure the ratio of active to apparent power (product of voltage and current). Connections for powerfactor meters and wattmeters are similar, and current and voltage transformers can be used to extend their range. Connections for single-phase and polyphase power-factor meters are shown in Figures 26 and 27, respectively.
11.
ROTATIVE SPEED AND POSITION MEASUREMENT
Tachometers Tachometers, or direct-measuring rpm counters, vary from handheld mechanical or electric meters to shaft-driven and electronic pulse counters. They are used in general laboratory and shop work to check rotative speeds of motors, engines, and turbines.
Stroboscopes Optical rpm counters produce a controlled high-speed electronic flashing light, which the operator directs on a rotating member, increasing the rate of flashes until reaching synchronism (the optical effect that rotation has stopped). At this point, the rpm measured is equal to the flashes per minute emitted by the strobe unit. Care must be taken to start at the bottom of the instrument scale and work up because multiples of the rpm produce almost the same optical effect as true synchronism. Multiples can be indicated by positioning suitable
AC Tachometer-Generators A tachometer-generator consists of a rotor and a stator. The rotor is a permanent magnet driven by the equipment. The stator is a winding with a hole through the center for the rotor. Concentricity is not critical; bearings are not required between rotor and stator. The output can be a single-cycle-per-revolution signal whose voltage is a linear function of rotor speed. The polypole configuration that generates 10 cycles per revolution allows measurement of speeds as low as 20 rpm without causing the indicating needle to flutter. The output of the AC tachometer-generator is rectified and connected to a DC voltmeter.
Optical (Shaft) Encoders Optical encoders and linear glass scales can be used to measure rotational speed and both linear and angular position, depending on the geometry of the encoder. The output from an encoder is typically in the form of a pulse train. If this pulse train provides an indication of the absolute orientation (angle or position), it is called an absolute encoder. If it produces a relative pulse train with no absolute angle or position reference, it is called an incremental encoder. Modern high-resolution rotational (shaft) encoders typically consist of a shaft-mounted rotating disk containing regularly spaced slits. A light source is located on one side of the disk slits and a detector on the other side of the slits. As the disk rotates at a given frequency, a pulse train is produced at the detector from the intermittent pulses of light that pass through the rotating slits. Linear displacement encoders (absolute or incremental) operate on a similar principle, except that the slits are formed in a linear axis. Although modern encoders are typically optical, a resistive potentiometer can also be used as a position or rotation encoder, the output resistance being a direct indication of position or orientation.
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Measurement and Instruments
12.
SOUND AND VIBRATION MEASUREMENT
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Measurement systems for determining sound pressure level, intensity level, and mechanical vibration generally use transducers to convert mechanical signals into electrical signals, which are then processed electronically or digitally to characterize the measured mechanical signals. These measurement systems contain one or more of the following elements, which may or may not be contained in a single instrument: • A transducer, or an assembly of transducers, to convert sound pressure or mechanical vibration (time-varying strain, displacement, velocity, acceleration, or force) into an electrical signal that is quantitatively related to the mechanical quantity being measured • Preamplifiers and amplifiers to provide functions such as preconditioning and amplification of signals, electrical impedance matching, signal conditioning, and gain • Signal-processing equipment to quantify those aspects of the signal that are being measured (peak value, rms value, time-weighted average level, power spectral density, or magnitude or phase of a complex linear spectrum or transfer function) and conduct integration, differentiation, and frequency weighting of the signal • Display and storage devices such as meters, oscilloscopes, digital displays, or level recorder to display and record the signal or the aspects of it that are being quantified • An interface that allows cable, wireless, or memory card output The relevant range of sound signals (i.e., audible to humans) can vary over more than six orders of magnitude in amplitude and more than three orders of magnitude in frequency, depending on the application. The relevant range of vibration signals may be slightly larger than this. References on instrumentation, measurement procedures, and signal analysis are given in the Bibliography. Product and application notes, technical reviews, and books published by instrumentation manufacturers are sources of additional reference material. See Chapter 48 of the 2015 ASHRAE Handbook—HVAC Applications and Chapter 8 of this volume for further information on sound and vibration.
12.1
SOUND MEASUREMENT
37.29 measurement; and any other pertinent considerations, such as size and directional characteristics.
Sound Measurement Systems Microphone preamplifiers, amplifiers, weighting networks, filters, analyzers, and displays are available either separately or integrated into a measuring instrument such as a sound level meter, personal noise exposure meter (often called a noise dose meter or dosimeter), measuring amplifier, or real-time constant-percentage bandwidth (e.g., octave band) or narrow-band [e.g., fast Fourier transform (FFT)] frequency analyzer. Instruments included in a sound measurement system depend on the purpose of the measurement, the frequency range, and the resolution of the signal analysis. For community and industrial noise measurements for regulatory purposes, the instrument, signal processing, and quantity to be measured are usually dictated by the pertinent regulation. The optimal instrument set generally varies for measurement of different characteristics such as sound power in HVAC ducts, sound power emitted by machinery, noise criteria (NC) numbers, sound absorption coefficients, sound transmission loss of building partitions, and reverberation times (T60).
Frequency Analysis Measurement criteria often dictate using filters to analyze the signal, to indicate the spectrum of the sound being measured. Filters of different bandwidths for different purposes include fractional octave band (one, one-third, one-twelfth, etc.), constant-percentage bandwidth, and constant (typically narrow) bandwidth. The filters may be analog or digital and, if digital, may or may not be capable of real-time data acquisition during measurement, depending on the bandwidth of frequency analysis. FFT signal analyzers are generally used in situations that require very narrow-resolution signal analysis at constant bandwidth when the amplitudes of the sound spectra vary significantly with respect to frequency. This may occur in regions of resonance or when it is necessary to identify narrowband or discrete sine-wave signal components of a spectrum in the presence of other such components or of broadband noise. However, when the frequency varies (e.g., because of nonconstant rpm of a motor), results from FFT analyzers can be difficult to interpret because the change in rpm provides what looks like a broadband signal.
Microphones
Sound Chambers
A microphone is a transducer that transforms an acoustical signal into an electrical signal. The two predominant transduction principles used in sound measurement (as opposed to broadcasting) are the electrostatic and the piezoelectric. Electrostatic (capacitor) microphones are available either as electret microphones, which do not require an external polarizing voltage, or as condenser microphones, which do require an external polarizing voltage, typically in the range of 28 to 200 V (DC). Piezoelectric microphones may be manufactured using either natural piezoelectric crystals or poled ferroelectric crystals. The types of response characteristics of measuring microphones are pressure, free field, and random incidence (diffuse field). The sensitivity and the frequency range over which the microphone has uniform sensitivity (flat frequency response) vary with sensing element diameter (surface area) and microphone type. Other critical factors that may affect microphone/preamplifier performance or response are atmospheric pressure, temperature, relative humidity, external magnetic and electrostatic fields, mechanical vibration, and radiation. Microphone selection is based on longand short-term stability; the match between performance characteristics (e.g., sensitivity, frequency response, amplitude linearity, self-noise) and the expected amplitude of sound pressure, frequency, range of analysis, and expected environmental conditions of
Special rooms and procedures are required to characterize and calibrate sound sources and receivers. The rooms are generally classified into three types: anechoic, hemianechoic, and reverberant. In the ideal anechoic room, all boundary surfaces completely absorb sound energy at all frequencies of interest. The ideal hemianechoic room would be identical to the ideal anechoic room, except that one surface would totally reflect sound energy at all frequencies. The ideal reverberant room would have boundary surfaces that totally reflect sound energy at all frequencies of interest. Anechoic chambers are used to perform measurements under conditions approximating those of a free sound field. They can be used in calibrating and characterizing individual microphones, microphone arrays, acoustic intensity probes, reference sound power sources, loudspeakers, sirens, and other individual or complex sources of sound. Hemianechoic chambers have a hard reflecting floor to accommodate heavy machinery or to simulate large factory floor or outdoor conditions. They can be used in calibrating and characterizing reference sound power sources, obtaining sound power levels of noise sources, and characterizing sound output of emergency vehicle sirens when mounted on an emergency motor vehicle. Reverberation chambers are used to perform measurements under conditions approximating those of a diffuse sound field. They
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2017 ASHRAE Handbook—Fundamentals (SI)
can be used in calibrating and characterizing random-incidence microphones and reference sound power sources, obtaining sound power ratings of equipment and sound power levels of noise sources, measuring sound absorption coefficients of building materials and panels, and measuring transmission loss through building partitions and components such as doors and windows. The choice of which room type to use often depends on the test method required for the subject units, testing costs, or room availability.
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Calibration A measurement system should be calibrated as a system from microphone or probe to indicating device before it is used to perform absolute measurements of sound. Acoustic calibrators and pistonphones of fixed or variable frequency and amplitude are available for this purpose. These calibrators should be used at a frequency low enough that the pressure, free-field, and random-incidence response characteristics of the measuring microphone(s) are, for practical purposes, equivalent, or at least related in a known quantitative manner for that specific measurement system. In general, the sound pressure produced by these calibrators may vary, depending on microphone type, whether the microphone has a protective grid, atmospheric pressure, temperature, and relative humidity. Correction factors and coefficients are required when conditions of use differ from those existing during the calibration of the acoustic calibrator or pistonphone. For demanding applications, precision sound sources and measuring microphones should periodically be sent to the manufacturer, a private testing laboratory, or a national standards laboratory for calibration.
12.2
VIBRATION MEASUREMENT
Except for seismic instruments that record or indicate vibration directly with a mechanical or optomechanical device connected to the test surface, vibration measurements use an electromechanical or interferometric vibration transducer. Here, the term vibration transducer refers to a generic electromechanical vibration transducer. Electromechanical and interferometric vibration transducers belong to a large and varied group of transducers that detect mechanical motion and furnish an electrical signal that is quantitatively related to a particular physical characteristic of the motion. Depending on design, the electrical signal may be related to mechanical strain, displacement, velocity, acceleration, or force. The operating principles of vibration transducers may involve optical interference; electrodynamic coupling; piezoelectric (including poled ferroelectric) or piezoresistive crystals; or variable capacitance, inductance, reluctance, or resistance. A considerable variety of vibration transducers with a wide range of sensitivities and bandwidths is commercially available. Vibration transducers may be contacting (e.g., seismic transducers) or noncontacting (e.g., interferometric, optical, or capacitive).
Transducers Seismic transducers use a spring-mass resonator within the transducer. At frequencies much greater than the fundamental natural frequency of the mechanical resonator, the relative displacement between the base and the seismic mass of the transducer is nearly proportional to the displacement of the transducer base. At frequencies much lower than the fundamental resonant frequency, the relative displacement between the base and the seismic mass of the transducer is nearly proportional to the acceleration of the transducer base. Therefore, seismic displacement transducers and seismic electrodynamic velocity transducers tend to have a relatively compliant suspension with a low resonant frequency; piezoelectric accelerometers and force transducers have a relatively stiff suspension with a high resonant frequency.
Strain transducers include the metallic resistance gage and piezoresistive strain gage. For dynamic strain measurements, these are usually bonded directly to the test surface. The accuracy with which a bonded strain gage replicates strain occurring in the test structure is largely a function of how well the strain gage was oriented and bonded to the test surface. Displacement transducers include the capacitance gage, fringecounting interferometer, seismic displacement transducer, optical approaches, and the linear variable differential transformer (LVDT). Velocity transducers include the reluctance (magnetic) gage, laser Doppler interferometer, and seismic electrodynamic velocity transducer. Accelerometers and force transducers include the piezoelectric, piezoresistive, and force-balance servo.
Vibration Measurement Systems Sensitivity, frequency limitations, bandwidth, and amplitude linearity of vibration transducers vary greatly with the transduction mechanism and the manner in which the transducer is applied in a given measurement apparatus. Contacting transducers’ performance can be significantly affected by the mechanical mounting methods and points of attachment of the transducer and connecting cable and by the mechanical impedance of the structure loading the transducer. Amplitude linearity varies significantly over the operating range of the transducer, with some transducer types or configurations being inherently more linear than others. Other factors that may critically affect performance or response are temperature; relative humidity; external acoustic, magnetic, and electrostatic fields; transverse vibration; base strain; chemicals; and radiation. A vibration transducer should be selected based on its long- and short-term stability; the match between its performance characteristics (e.g., sensitivity, frequency response, amplitude linearity, self-noise) and the expected amplitude of vibration, frequency range of analysis, and expected environmental conditions of measurement; and any other pertinent considerations (e.g., size, mass, resonant frequency). Vibration exciters, or shakers, are used in structural analysis, vibration analysis of machinery, fatigue testing, mechanical impedance measurements, and vibration calibration systems. Vibration exciters have a table or moving element with a drive mechanism that may be mechanical, electrodynamic, piezoelectric, or hydraulic. They range from relatively small, low-power units for calibrating transducers (e.g., accelerometers) to relatively large, high-power units for structural and fatigue testing. Conditioning amplifiers, power supplies, preamplifiers, charge amplifiers, voltage amplifiers, power amplifiers, filters, controllers, and displays are available either separately or integrated into a measuring instrument or system, such as a structural analysis system, vibration analyzer, vibration monitoring system, vibration meter, measuring amplifier, multichannel data-acquisition and modal analysis system, or real-time fractional-octave or FFT signal analyzer. The choice of instruments to include in a vibration measurement system depends on the mechanical quantity to be determined, purpose of measurement, and frequency range and resolution of signal analysis. For vibration measurements, the signal analysis is relatively narrow in bandwidth and may be relatively low in frequency, to accurately characterize structural resonances. Accelerometers with internal integrated circuitry are available to provide impedance matching or servo control for measuring very-low-frequency acceleration (servo accelerometers). Analog integration and differentiation of vibration signals are available through integrating and differentiating networks and amplifiers, and digital is available through FFT analyzers. Vibration measurements made for different purposes (e.g., machinery diagnostics and health monitoring, balancing rotating machinery, analysis of torsional vibration, analysis of machine-tool vibration, modal analysis, analysis of vibration isolation, stress monitoring, industrial control) generally have
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Measurement and Instruments different mechanical measurement requirements and a different optimal set of instrumentation.
Calibration Because of their inherent long- and short-term stability, amplitude linearity, wide bandwidth, wide dynamic range, low noise, and wide range of sensitivities, seismic accelerometers have traditionally been used as a reference standard for dynamic mechanical measurements. A measurement system should be calibrated as a system from transducer to indicating device before it is used to perform absolute dynamic measurements of mechanical quantities. Calibrated reference vibration exciters, standard reference accelerometers, precision conditioning amplifiers, and precision calibration exciters are available for this purpose. These exciters and standard reference accelerometers can be used to transfer a calibration to another transducer. For demanding applications, a calibrated exciter or standard reference accelerometer with connecting cable and conditioning amplifier should periodically be sent to the manufacturer, a private testing laboratory, or a national standards laboratory for calibration.
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13. LIGHTING MEASUREMENT Light level, or illuminance, is usually measured with a photocell made from a semiconductor such as silicon or selenium. Photocells produce an output current proportional to incident luminous flux when linked with a microammeter, color- and cosine-corrected filters, and multirange switches; they are used in inexpensive handheld light meters and more precise instruments. Different cell heads allow multirange use in precision meters. Cadmium sulfide photocells, in which resistance varies with illumination, are also used in light meters. Both gas-filled and vacuum photoelectric cells are in use. Small survey-type meters are not as accurate as laboratory meters; their readings should be considered approximate, although consistent, for a given condition. Their range is usually from 50 to 50 000 lux. Precision low-level meters have cell heads with ranges down to 0 to 20 lux. A photometer installed in a revolving head is called a goniophotometer and is used to measure the distribution of light sources or luminaires. To measure total luminous flux, the luminaire is placed in the center of a sphere painted inside with a high-reflectance white with a near-perfect diffusing matte surface. Total light output is measured through a small baffled window in the sphere wall. To measure irradiation from germicidal lamps, a filter of fused quartz with fluorescent phosphor is placed over the light meter cell. If meters are used to measure the number of lumens per unit area diffusely leaving a surface, luminance (cd/m2) instead of illumination (lux) is read. Light meters can be used to measure luminance, or electronic lux meters containing a phototube, an amplifier, and a microammeter can read luminance directly. Chapter 9 of the IES (2011) Lighting Handbook gives detailed information on measurement of light.
14. THERMAL COMFORT MEASUREMENT Thermal comfort depends on the combined influence of clothing, activity, air temperature, air velocity, mean radiant temperature, and air humidity. Thermal comfort is influenced by heating or cooling of particular body parts through radiant temperature asymmetry (plane radiant temperature), draft (air temperature, air velocity, turbulence), vertical air temperature differences, and floor temperature (surface temperature).
37.31 A general description of thermal comfort is given in Chapter 9, and guidelines for an acceptable thermal environment are given in ASHRAE Standard 55 and ISO Standard 7730. ASHRAE Standard 55 also includes required measuring accuracy. In addition to specified accuracy, ISO Standard 7726 includes recommended measuring locations and a detailed description of instruments and methods.
Clothing and Activity Level These values are estimated from tables (Chapter 9; ISO Standards 8996, 9920). Thermal insulation of clothing [(m2 ·K)/W] can be measured on a thermal mannequin (McCullough et al. 1985; Olesen 1985). Activity (W/m2) can be estimated from measuring CO2 and O2 in a person’s expired air.
Air Temperature Various types of thermometers may be used to measure air temperature. Placed in a room, the sensor registers a temperature between air temperature and mean radiant temperature. One way of reducing the radiant error is to make the sensor as small as possible, because the convective heat transfer coefficient increases as size decreases, whereas the radiant heat transfer coefficient is constant. A smaller sensor also provides a favorably low time constant. Radiant error can also be reduced by using a shield (an open, polished aluminum cylinder) around the sensor, using a sensor with a low-emittance surface, or artificially increasing air velocity around the sensor (aspirating air through a tube in which the sensor is placed).
Air Velocity In occupied zones, air velocities are usually small (0 to 0.5 m/s), but do affect thermal sensation. Because velocity fluctuates, the mean value should be measured over a suitable period, typically 3 min. Velocity fluctuations with frequencies up to 1 Hz significantly increase human discomfort caused by draft, which is a function of air temperature, mean air velocity, and turbulence (see Chapter 9). Fluctuations can be given as the standard deviation of air velocity over the measuring period (3 min) or as the turbulence intensity (standard deviation divided by mean air velocity). Velocity direction may change and is difficult to identify at low air velocities. An omnidirectional sensor with a short response time should be used. A thermal anemometer is suitable. If a hot-wire anemometer is used, the direction of measured flow must be perpendicular to the hot wire. Smoke puffs can be used to identify the direction.
Plane Radiant Temperature This refers to the uniform temperature of an enclosure in which the radiant flux on one side of a small plane element is the same as in the actual nonuniform environment. It describes the radiation in one direction. Plane radiant temperature can be calculated from surface temperatures of the environment (half-room) and angle factors between the surfaces and a plane element (ASHRAE Standard 55). It may also be measured by a net-radiometer or a radiometer with a sensor consisting of a reflective disk (polished) and an absorbent disk (painted black) (Olesen et al. 1989).
Mean Radiant Temperature This is the uniform temperature of an imaginary black enclosure in which an occupant would exchange the same amount of radiant heat as in the actual nonuniform enclosure. Mean radiant temperature can be calculated from measured surface temperatures and the corresponding angle factors between the person and surfaces. It can also be determined from the plane radiant temperature in six opposite directions, weighted according to the projected area factors for a person. For more information, see Chapter 9.
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2017 ASHRAE Handbook—Fundamentals (SI)
Because of its simplicity, the instrument most commonly used to determine the mean radiant temperature is a black globe thermometer (Bedford and Warmer 1935; Vernon 1932). This thermometer consists of a hollow sphere usually 150 mm in diameter, coated in flat black paint with a thermocouple or thermometer bulb at its center. The temperature assumed by the globe at equilibrium results from a balance between heat gained and lost by radiation and convection. Mean radiant temperatures are calculated from 8
0.6
1.10 10 V a t r = t g + 273 4 + ---------------------------------- tg – ta 0.4 D
14
– 273
(12)
where
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tr tg Va ta D
= = = = = =
mean radiant temperature, °C globe temperature, °C air velocity, m/s air temperature, °C globe diameter, m emissivity (0.95 for black globe)
According to Equation (12), air temperature and velocity around the globe must also be determined. The globe thermometer is spherical, but mean radiant temperature is defined in relation to the human body. For sedentary people, the globe represents a good approximation. For people who are standing, the globe, in a radiant nonuniform environment, overestimates the radiation from floor or ceiling; an ellipsoidal sensor gives a closer approximation. A black globe also overestimates the influence of short-wave radiation (e.g., sunshine). A flat gray color better represents the radiant characteristic of normal clothing (Olesen et al. 1989). The hollow sphere is usually made of copper, which results in an undesirably high time constant. This can be overcome by using lighter materials (e.g., a thin plastic bubble).
Air Humidity The water vapor pressure (absolute humidity) is usually uniform in the occupied zone of a space; therefore, it is sufficient to measure absolute humidity at one location. Many of the instruments listed in Table 3 are applicable. At ambient temperatures that provide comfort or slight discomfort, the thermal effect of humidity is only moderate, and highly accurate humidity measurements are unnecessary.
14.1
CALCULATING THERMAL COMFORT
When the thermal parameters have been measured, their combined effect can be calculated by the thermal indices in Chapter 9. For example, the effective temperature (Gagge et al. 1971) can be determined from air temperature and humidity. Based on the four environmental parameters and an estimation of clothing and activity, the predicted mean vote (PMV) can be determined with the aid of tables (Chapter 9; Fanger 1982; ISO Standard 7730). The PMV is an index predicting the average thermal sensation that a group of occupants may experience in a given space. For certain types of normal activity and clothing, measured environmental parameters can be compared directly with those in ASHRAE Standard 55 or ISO Standard 7730.
14.2
INTEGRATING INSTRUMENTS
Several instruments have been developed to evaluate the combined effect of two or more thermal parameters on human comfort. Madsen (1976) developed an instrument that gives information on the occupants’ expected thermal sensation by directly measuring the PMV value. The comfort meter has a heated elliptical sensor that simulates the body (Figure 28). The estimated clothing (insulation value), activity in the actual space, and humidity are set on the instru-
Fig. 28 Madsen’s Comfort Meter (Madsen 1976)
ment. The sensor then integrates the thermal effect of air temperature, mean radiant temperature, and air velocity in approximately the same way the body does. The electronic instrument gives the measured operative and equivalent temperature, calculated PMV, and predicted percentage of dissatisfied (PPD).
15. MOISTURE CONTENT AND TRANSFER MEASUREMENT Moisture commonly refers to the presence of liquid and vapor states of water, which has two positively charged hydrogen atoms and one negatively charged oxygen atom. As moisture vapor and liquid molecules adsorb to a hygroscopic material (e.g., most porous building materials), the material’s moisture content (MC) increases significantly. The presence of moisture in materials influences their performance (e.g., thermal insulation, acoustics, processability, dielectricity, storage life, unhealthy microbe growth, corrosion, chemical aspects). However, little off-the-shelf instrumentation exists to measure the moisture content or transfer of porous materials, although many measurements can be set up with a small investment of time and money. Three moisture properties are most commonly sought: (1) moisture content; (2) vapor permeability (rate at which water vapor passes through a given material); and (3) liquid diffusivity (rate at which liquid water passes through a porous material).
Moisture Content MC can be classified into four categories: • Wet-basis volumetric MC: ratio of moisture’s volume to wet sample’s bulk volume • Wet-basis mass MC: ratio of moisture’s mass to wet sample’s bulk mass • Dry-basis volumetric MC: ratio of moisture’s volume to dry sample’s bulk volume • Dry-basis mass MC: ratio of moisture’s mass to dry sample’s bulk mass Dry basis is often used in building materials. The MC of a material is usually described with a sorption isotherm, which relates the equilibrium moisture content (EMC) of a hygroscopic material to the ambient relative humidity under constant temperature. Determining a sorption isotherm involves exposing a sample of material to a known relative humidity at a known temperature and then measuring the sample’s moisture content after enough time has elapsed for the sample to reach equilibrium with its surroundings. Hysteresis in the sorption behavior of most hygroscopic materials requires that measurements be made for both increasing (adsorption isotherm) and decreasing relative humidity (desorption isotherm). Figure 29 shows hysteresis through the
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Measurement and Instruments adsorption and desorption isotherms for a typical hygroscopic porous material (Straube 1998). Three different regimes of moisture storage are shown: sorption or hygroscopic (A), capillary (B), and oversaturated (C). In the hygroscopic regime, water vapor adsorbs to the pore walls. As relative humidity increases, adsorbed moisture molecules grow from single layers to multiple interconnected layers (internal capillary condensation). The capillary regime (B) is somewhat arbitrarily designated as that part of the moisture storage function above the critical moisture content. Physically, it is presumed that a continuous liquid phase forms. Finally, in the supersaturated state, the relative humidity is always 100% and no more water will wick into a material. Ambient relative humidity can be controlled using saturated salt solutions or mechanical refrigeration equipment (Carotenuto et al. 1991; Cunningham and Sprott 1984; Tveit 1966). Precise measurements of the relative humidity produced by various salt solutions were reported by Greenspan (1977). ASTM Standard E104 describes the use of saturated salt solutions. A sample’s EMC is usually determined gravimetrically using a precision balance. The sample’s dry mass, necessary to calculate moisture content, can be found by oven or desiccant drying. Oven dry mass may be lower than desiccant dry mass because of the loss of volatiles other than water in the oven (Richards et al. 1992). A major difficulty in measuring sorption isotherms of engineering materials is the long time required for many materials to reach equilibrium (often as long as weeks or months). The rate-limiting mechanism for these measurements is usually the slow process of vapor diffusion into the pores of the material. Using smaller samples can reduce diffusion time. Note that, although EMC isotherms are traditionally plotted as a function of relative humidity, the actual transport to or from materials is determined by vapor pressure differences. Thus, significant moisture content changes can occur because of changes in either the material vapor pressure or the surrounding air long before equilibrium is reached. Moisture content can be directly or indirectly determined. Direct techniques measure moisture by a chemical reaction (e.g., Karl Fisher titration) or using the difference in mass before and after drying a test sample. Using chemical reactions is complex, slow, and expensive, and uses toxic chemicals, so it is seldom used. Mass loss after drying is usually referred to as the loss on drying (LOD) or thermo-gravimetric method (or simply gravimetric), in which a sample is weighed, heated in an oven at a set temperature (e.g., 100 to 105°C) for an appropriate period to a
37.33 constant weight, cooled in the dry atmosphere, and reweighed. Although these direct methods are accurate (within ±0.01 m3/m3), they are time-consuming, destructive, and do not allow for in situ measurement. Therefore, many indirect methods have been developed. Unlike direct measurements, indirect techniques do not involve removing moisture in a sample: they estimate moisture content by a strong or calibrated relationship with some other measurable variables such as dielectric permittivity or thermal conductivity. Both empirical and theoretical equations between the moisture content and measurable variable are used for the calibration. Major types of indirect methods include neutron moderation, gamma ray attenuation, nuclear magnetic resonance, microwave reflectance and attenuation, near-infrared reflectance or transmission (NIR/T), dielectric techniques, and thermal methods. Safety issues may limit application of the first three techniques, although they have some advantages such as robustness or high accuracy. NIR/T is a rapid, noninvasive technique and is officially sanctioned for determining moisture content in several applications (e.g., in grain). Its limiting factors include high cost, low penetration depth, and a large sample size required for calibration. Dielectric techniques take advantage of the strong dependence of the composite material’s bulk permittivity (or dielectric constant) on moisture content. Because the permittivity of water (Kaw = 81) is much higher than that of the porous material’s other constituents (e.g., 1 for air, 4.7 for fiberglass), the total permittivity of the composite porous material is mainly determined by the moisture content. Dielectric methods include time domain reflectometry (TDR), frequency domain (FD), amplitude domain reflectometry (ADR), phase transmission (Virrib), and time domain transmission (TDT). These dielectric moisture content sensors are becoming popular in various field and laboratory applications (e.g., soil, wood) because they have short response time (almost instantaneous measurements), do not require maintenance, and can provide good accuracy (commercial devices can be accurate within ±0.02 or 0.03 m3/m3 and up to ±0.01 m3/m3 in some more specific situations). However, their sensitivity to boundary water or salinity limits their applications in a highly saline/conductive condition. Thermal techniques are increasingly considered as an alternative to other methodologies to determine moisture content in porous materials. The main attractions in pursuing this technology are simultaneous measurements of other thermophysical properties (e.g., thermal conductivity, thermal diffusivity, volumetric heat capacity, and water matric potential), wide measurement range, invulnerability of salinity, low cost, high accuracy and robustness. Two predominant concerns limiting the development of the thermal technique include slow reaction time (one commercial product with the accuracy of ±0.01 m3/m3 has the response time of two minutes) and requirement of good contact with the test sample. One typical example of the thermal method identifies moisture content y by its relationship with volumetric specific heat capacity c, as described by Bristow et al. (1993, 1994), Campbell et al. (1991), and Yang et al. (2015). There is no universal method of moisture content measurement suitable to every application, and detected moisture content varies with measurement method. The suitability of each method relies on different application considerations such as safety, cost, accuracy, response time, installation, ease of operation, management, and durability.
Vapor Permeability Diffusive transfer of water vapor through porous materials is often described by a modified form of Fick’s law: Fig. 29
Adsorption Isotherm and Desorption Isotherm for Hygroscopic Material (Straube 1998)
dp wv = – -----dx
(13)
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2017 ASHRAE Handbook—Fundamentals (SI)
where
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wv = mass of vapor diffusing through unit area per unit time, mg/(s·m2) dp/dx = vapor pressure gradient, kPa/m = vapor permeability, mg/(s·m·kPa)
In engineering practice, permeance may be used instead of permeability. Permeance is simply permeability divided by the material thickness in the direction of vapor flow; thus, permeability is a material property, whereas permeance depends on thickness. Permeability is measured with wet-cup, dry-cup, or modified cup tests. Specific test methods for measuring water vapor permeability are given in ASTM Standard E96. For many engineering materials, vapor permeability is a strong function of mean relative humidity. Wet and dry cups cannot adequately characterize this dependence on relative humidity. Instead, a modified cup method can be used, in which pure water or desiccant in a cup is replaced with a saturated salt solution (Burch et al. 1992; McLean et al. 1990). A second saturated salt solution is used to condition the environment outside the cup. Relative humidities on both sides of the sample material can be varied from 0 to 100%. Several cups with a range of mean relative humidities are used to map out the dependence of vapor permeability on relative humidity. In measuring materials of high permeability, the finite rate of vapor diffusion through air in the cup may become a factor. Air-film resistance could then be a significant fraction of the sample’s resistance to vapor flow. Accurate measurement of high-permeability materials may require an accounting of diffusive rates across all air gaps (Fanney et al. 1991).
Liquid Diffusivity Transfer of liquid water through porous materials may be characterized as a diffusion-like process: d wl = – D l -----dx
(14)
where wl = mass of liquid transferred through unit area per unit time, kg/(s·m2) = liquid density, kg/m3 Dl = liquid diffusivity, m2/s d/dx = moisture content gradient, m1
Dl typically depends strongly on moisture content. Transient measurement methods deduce the functional form of Dl by observing the evolution of a one-dimensional moisture content profile over time. An initially dry specimen is brought into contact with liquid water. Free water migrates into the specimen, drawn in by surface tension. The resulting moisture content profile, which changes with time, must be differentiated to find the material’s liquid diffusivity (Bruce and Klute 1956). Determining the transient moisture content profile typically involves a noninvasive and nondestructive method of measuring local moisture content. Methods include gamma ray absorption (Freitas et al. 1991; Kumaran and Bomberg 1985; Quenard and Sallee 1989), x-ray radiography (Ambrose et al. 1990), neutron radiography (Prazak et al. 1990), and nuclear magnetic resonance (NMR) (Gummerson et al. 1979). Uncertainty in liquid diffusivity measurement is often large because of the need to differentiate noisy experimental data.
16. HEAT TRANSFER THROUGH BUILDING MATERIALS Thermal Conductivity The thermal conductivity of a heat insulator, as defined in Chapter 25, is a unit heat transfer factor. Two methods of determining the
thermal conductivity of flat insulation are the guarded hot plate and the heat flow meter apparatus, according to ASTM Standards C177 and C518, respectively. Both methods use parallel isothermal plates to induce a steady temperature gradient across the thickness of the specimen(s). The guarded hot plate is considered an absolute method for determining thermal conductivity. The heat flow meter apparatus requires calibration with a specimen of known thermal conductivity, usually determined in the guarded hot plate. The heat flow meter apparatus is calibrated by determining the voltage output of its heat flux transducer(s) as a function of the heat flux through the transducer(s). Basic guarded hot plate design consists of an electrically heated plate and two liquid-cooled plates. Two similar specimens of a material are required for a test; one is mounted on each side of the hot plate. A cold plate is then pressed against the outside of each specimen by a clamp screw. The heated plate consists of two sections separated by a small gap. During tests, the central (metering) and outer (guard) sections are maintained at the same temperature to minimize errors caused by edge effects. The electric energy required to heat the metering section is measured carefully and converted to heat flow. Thermal conductivity of the material can be calculated under steady-state conditions using this heat flow quantity, area of the metering section, temperature gradient, and specimen thickness. Thermal conductivity of cylindrical or pipe insulation (Chapter 25) is determined similarly, but an equivalent thickness must be calculated to account for the cylindrical shape (ASTM Standard C335). Transient methods have been developed by D’Eustachio and Schreiner (1952), Hooper and Chang (1953), and Hooper and Lepper (1950) using a line heat source within a slender probe. These instruments are available commercially and have the advantages of rapidity and a small test specimen requirement. The probe is a useful research and development tool, but it has not been as accepted as the guarded hot plate, heat flow meter apparatus, or pipe insulation apparatus.
Thermal Conductance and Resistance Thermal conductances (C-factors) and resistances (R-values) of many building assemblies can be calculated from the conductivities and dimensions of their components, as described in Chapter 27. Test values can also be determined experimentally by testing large, representative specimens in the hot box apparatus described in ASTM Standards C976 and C1363. This laboratory apparatus measures heat transfer through a specimen under controlled air temperature, air velocity, and radiation conditions. It is especially suited for large, nonhomogeneous specimens. For in situ measurements, heat flux and temperature transducers are useful in measuring the dynamic or steady-state behavior of opaque building components (ASTM Standard C1046). A heat flux transducer is simply a differential thermopile within a core or substrate material. Two types of construction are used: (1) multiple thermocouple junctions wrapped around a core material, or (2) printed circuits with a uniform array of thermocouple junctions. The transducer is calibrated by determining its voltage output as a function of the heat flux through the transducer. For in situ measurements, the transducer is installed in either the wall or roof, or mounted on an exterior surface with tape or adhesive. Data obtained can be used to compute the thermal conductance or resistance of the building component (ASTM Standard C1155).
17.
AIR CONTAMINANT MEASUREMENT
Three measures of particulate air contamination include the number, projected area, and mass of particles per unit volume of air (ASTM 2012). Each requires an appropriate sampling technique.
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Measurement and Instruments Particles are counted by capturing them in impingers, impactors, membrane filters, or thermal or electrostatic precipitators. Counting may be done by microscope, using stage counts if the sample covers a broad range of sizes (Nagda and Rector 2001). Electronic particle counters can give rapid data on particle size distribution and concentration. Inertial particle counters use acceleration to separate sampled particles into different sizes. Real-time aerodynamic particle sizers (APS) use inertial effects to separate particles by size, but instead of capturing the particles, they are sized optically (Cox and Miro 1997), and can provide continuous sampling; however, they tend to be very expensive. Other, less costly types of optical particle counters (OPCs) are also available, but they typically require careful calibration using the type of particle that is being measured for accurate results (Baron and Willeke 2001). Their accuracy also depends heavily on appropriate maintenance and proper application. Correction for particle losses (dropout in the sampling lines) during sampling can be particularly important for accurate concentration measurements. Concentration uncertainty (random measurement uncertainty) also depends on the number of particles sampled in a given sampling interval. Particle counters have been used in indoor office environments as well as in cleanrooms, and in aircraft cabin air quality testing (Cox and Miro 1997). Projected area determinations are usually made by sampling onto a filter paper and comparing the light transmitted or scattered by this filter to a standard filter. The staining ability of dusts depends on the projected area and refractive index per unit volume. For sampling, filters must collect the minimum-sized particle of interest, so membrane or glass fiber filters are recommended. To determine particle mass, a measured quantity of air is drawn through filters, preferably of membrane or glass fiber, and the filter mass is compared to the mass before sampling. Electrostatic or thermal precipitators and various impactors have also been used. For further information, see ACGIH (2001), Lodge (1989), and Lundgren et al. (1979). Chapter 46 of the 2015 ASHRAE Handbook—HVAC Applications presents information on measuring and monitoring gaseous contaminants. Relatively costly analytical equipment, which must be calibrated and operated carefully by experienced personnel, is needed. Numerous methods of sampling the contaminants, as well as the laboratory analysis techniques used after sampling, are specified. Some of the analytical methods are specific to a single pollutant; others can present a concentration spectrum for many compounds simultaneously.
18.
COMBUSTION ANALYSIS
Two approaches are used to measure the thermal output or capacity of a boiler, furnace, or other fuel-burning device. The direct or calorimetric test measures change in enthalpy or heat content of the fluid, air, or water heated by the device, and multiplies this by the flow rate to arrive at the unit’s capacity. The indirect test or flue gas analysis method determines heat losses in flue gases and the jacket and deducts them from the heat content (higher heating value) of measured fuel input to the appliance. A heat balance simultaneously applies both tests to the same device. The indirect test usually indicates the greater capacity, and the difference is credited to radiation from the casing or jacket and unaccounted-for losses. With small equipment, the expense of the direct test is usually not justified, and the indirect test is used with an arbitrary radiation and unaccounted-for loss factor.
18.1
FLUE GAS ANALYSIS
Flue gases from burning fossil fuels generally contain carbon dioxide (CO2) and water, with some small amounts of hydrogen
37.35 (H2), carbon monoxide (CO), nitrogen oxides (NOx), sulfur oxides (SOx), and unburned hydrocarbons. However, generally only concentrations of CO2 (or O2) and CO are measured to determine completeness of combustion and efficiency. Nondispersive infrared (NDIR) analyzers are the most common laboratory instruments for measuring CO and CO2. Their advantages include the following: (1) they are not very sensitive to flow rate, (2) no wet chemicals are required, (3) they have a relatively fast response, (4) measurements can be made over a wide range of concentrations, and (5) they are not sensitive to the presence of contaminants in ambient air. In the laboratory, oxygen is generally measured with an instrument that uses O2’s paramagnetic properties. Paramagnetic instruments are generally used because of their excellent accuracy and because they can be made specific to the measurement of oxygen. For field testing and burner adjustment, portable combustion testing equipment is available. These instruments generally measure O2 and CO with electrochemical cells. The CO2 is then calculated by an on-board microprocessor and, together with temperature, is used to calculate thermal efficiency.
19. DATA ACQUISITION AND RECORDING Almost every type of transducer and sensor is available with the necessary interface system to make it computer compatible. The transducer itself begins to lose its identity when integrated into a system with features such as linearization, offset correction, selfcalibration, and so forth. This has eliminated concern about the details of signal conditioning and amplification of basic transducer outputs, although engineering judgment is still required to review all data for validity, accuracy, and acceptability before making decisions based on the results. The personal computer is integrated into every aspect of data recording, including sophisticated graphics, acquisition and control, and analysis. Internet or intranet connections allow easy access to remote personal-computer-based datarecording systems from virtually any locale. Direct output devices can be either multipurpose or specifically designed for a given sensor. Traditional chart recorders still provide a visual indication and a hard-copy record of the data, but their output is now rarely used to process data. These older mechanical stylus-type devices use ink, hot wire, pressure, or electrically sensitive paper to provide a continuous trace. They are useful up to a few hundred hertz. Thermal and ink recorders are confined to chart speeds of several centimetres per second for recording relatively slow processes. Simple indicators and readouts are used mostly to monitor the output of a sensor visually, and have usually been replaced by modern digital indicators. Industrial environments commonly use signal transmitters for control or computer data-handling systems to convert the signal output of the primary sensor into a compatible common signal span (e.g., the standard 4-20 mA current loop). All signal conditioning (ranging, zero suppression, reference-junction compensation) is provided at the transmitter. Thus, all recorders and controllers in the system can have an identical electrical span, with variations only in charts and scales offering the advantages of interchangeability and economy in equipment cost. Long signal transmission lines can be used, and receiving devices can be added to the loop without degrading performance. Newer instruments may be digitally bus based, which removes the degradation that may occur with analog signals. These digital instruments are usually immune to noise, based on the communications scheme that is used. They also may allow for self-configuration of the sensor in the field to the final data acquisition device. The vast selection of available hardware, often confusing terminology, and challenge of optimizing the performance/cost ratio for
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37.36 a specific application make configuring a data acquisition system difficult. A system specifically configured to meet a particular measurement need can quickly become obsolete if it has inadequate flexibility. Memory size, recording speed, and signal processing capability are major considerations in determining the correct recording system. Thermal, mechanical, electromagnetic interference, portability, and meteorological factors also influence the selection.
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Digital Recording A digital data acquisition system must contain an interface, which is a system involving one or several analog-to-digital converters, and, in the case of multichannel inputs, circuitry for multiplexing. The interface may also provide excitation for transducers, calibration, and conversion of units. The digital data are arranged into one or several standard digital bus formats. Many data acquisition systems are designed to acquire data rapidly and store large records of data for later recording and analysis. Once the input signals have been digitized, the digital data are essentially immune to noise and can be transmitted over great distances. The most popular physical layer bus standards used for data transmission are the TIA/EIA-485, IEEE 802.3, USB, IEEE 488, or general-purpose interface bus (GPIB) and the RS232 serial interface. The TIA/EIA-485 bus system can be based on either two twisted pairs for full duplex (simultaneous signaling) operation or half duplex using only one twisted pair. A ground wire is generally required to provide a common reference ground for all the devices on the bus to eliminate common mode voltage problems caused by unequal ground potentials. The TIA/EIA-485 standard does not specify cabling length and data rate specifications, but refers to application guidelines (TSB-89). For practical applications, it gives an example of a specific manufacturer’s 24AWG cable, for which the maximum distance between two devices is 1200 m at 100 kbps (TSB-89). The actual distance between devices for a given data rate depends on cable capacitance and impedance, network configuration, and grounding arrangements (TIA Standards TIA-485, TSB-89). The IEEE 802.3 bus standard specifies both copper and optical fiber for physical linking of devices (IEEE Standard 802.3). The copper version uses either twin coaxial cables or four twisted-wire pairs with the twisted wire pairs being used more commonly. The twisted-wire pairs can support up to 10 Gbps with maximum distance of 100 m between devices depending on type of CAT (cable and telephone) cable used (IEEE 802.3). A 40 Gbps data rate specification over twisted pairs is expected to be released by end of 2016. The USB 2.0 and 3.1 uses one and three twisted pairs respectively for data transmission. Three additional wires are used for power, ground and shield in both USB 2.0 and USB 3.0. The USB 2.0 can provide 480 Mbps with a recommended length of 5 m based on the maximum 26 ns signal delay and cable specifications given in the USB 2.0 standard. The new USB 3.1 standard can support data transmission speeds of up to 10 Gbps with a recommended length of 3 m for a cable satisfying the standard’s electrical requirements. Device distances can be increased by using either USB bridges/hubs or physical layer interfacing converters. The IEEE 488 bus system feeds data down eight parallel wires, one data byte at a time. This parallel operation allows it to transfer data rapidly at up to 1 million characters per second. However, the IEEE 488 bus is limited to a cable length of 20 m and requires an interface connection on every meter for proper termination. The RS232 system feeds data serially down two wires, one bit at a time. The distance between the devices is limited by maximum cable capacitance of 2500 pF at a maximum data rate of 20 kbps (TIA Standard TIA-232). This translates to typical distance of 15 m, which was the maximum length limit specified in an earlier version
2017 ASHRAE Handbook—Fundamentals (SI) of the RS232 standard. For longer distances, it may feed a modem to send data over standard telephone lines. These physical layer buses can be connected to a data acquisition unit or a personal computer directly or through a local area network (LAN) available in a facility for transmitting information. With appropriate interfacing, transducer data are available to any computer connected to the network. Bus measurements can greatly simplify three basic applications: data gathering, automated limit testing, and computer-controlled processes. Data gathering collects readings over time. The most common applications include aging tests in quality control, temperature tests in quality assurance, and testing for intermittency in service. A controller can monitor any output indefinitely and then display the data directly on screen or record it on magnetic tape or disks for future use. In automated limit testing, the computer compares each measurement with programmed limits. The controller converts readings to a good/bad readout. Automatic limit testing is highly costeffective when working with large number of parameters of a particular unit under test. In computer-controlled processes, the IEEE 488 bus system becomes a permanent part of a larger, completely automated system. For example, a large industrial process may require many electrical sensors that feed a central computer controlling many parts of the manufacturing process. An IEEE 488 bus controller collects readings from several sensors and saves the data until asked to dump an entire batch of readings to a larger central computer at one time. Used in this manner, the IEEE 488 bus controller serves as a slave of the central computer. Dynamic range and accuracy must be considered in a digital recording system. Dynamic range refers to the ratio of the maximum input signal for which the system is useful to the noise floor of the system. The accuracy figure for a system is affected by the signal noise level, nonlinearity, temperature, time, crosstalk, and so forth. In selecting an 8-, 12-, or 16-bit analog-to-digital converter, the designer cannot assume that system accuracy is necessarily determined by the resolution of the encoders (i.e., 0.4%, 0.025%, and 0.0016%, respectively). If the sensor preceding the converter is limited to 1% full-scale accuracy, for example, no significant benefits are gained by using a 12-bit system over an 8-bit system and suppressing the least significant bit. However, a greater number of bits may be required to cover a larger dynamic range.
Data-Logging Devices Data loggers digitally store electrical signals (analog or digital) to an internal memory storage component. The signal from connected sensors is typically stored to memory at timed intervals ranging from MHz to hourly sampling. Some data loggers store data based on an event (e.g., button push, contact closure). Many data loggers can perform linearization, scaling, or other signal conditioning and allow logged readings to be either instantaneous or averaged values. Most data loggers have built-in clocks that record the time and date together with transducer signal information. Data loggers range from single-channel input to 256 or more channels. Some are general-purpose devices that accept a multitude of analog and/or digital inputs, whereas others are more specialized to a specific measurement (e.g., a portable anemometer with built-in data-logging capability) or application (e.g., a temperature, relative humidity, CO2, and CO monitor with data logging for IAQ applications). Stored data are generally downloaded using a serial interface with a temporary direct connection to a personal computer. Some data loggers allow downloading directly to a printer, or to an external hard drive or tape drive that can later be connected to a PC. With the reduction in size of personal computers (laptops, notebooks, hand-held PCs, and palmtops), the computer itself is now being used as the data logger. These mobile computers may be left
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Measurement and Instruments in the field, storing measurements from sensors directly interfaced into the computer. Depending on the particular application and number of sensors to be read, a computer card mounted directly into the PC may eliminate the external data acquisition device completely.
20. MECHANICAL POWER MEASUREMENT Power measurement quantities mechanical power, and electrical power for both AC and DC loading (described in the section on Electric Measurement). Examples of mechanical power include shaft power and pumping power.
Measurement of Shaft Power The measurement of shaft power associated with rotating machinery can be accomplished using a dynamometer setup which independently measures torque and rotation rate (or angular velocity). P = T
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where P = power, W = shaft angular velocity, rad/s T = torque, N·m
Measurement of Fluid Pumping Power Pumping power can be determined by independent measurements of volumetric flow rate and pressure differential across the pump. Q p P = ---------------60 where P = power, W Q = volumetric flow rate, L/min p = pressure differential across the pump, kPa
20.1 A a C Cp cp D d Dl d/dx dp/dx E Fa H J K
= = = = = = = = = = = = = = =
n= p= pw = Pwet = P= Q= R= r= S= t=
m2
SYMBOLS
flow area, thermocouple constant correction factor pitot-static probe pressure difference coefficient specific heat at constant pressure, kJ/(kg·K) distance; diameter throat diameter liquid diffusivity, m2/s moisture content gradient, m–1 vapor pressure gradient, kPa/m voltage thermal expansion correction factor height mechanical equivalent of heat = 100 (N·m)/kJ sensitivity (Figure 1); differential expansion coefficient for liquid in glass; constant (function of geometry and Reynolds number) number of degrees that liquid column emerged from bath absolute pressure, Pa velocity pressure (pitot-tube manometer reading), Pa wetted perimeter mechanical power, W discharge flow rate, m3/s; L/min resistance, (see Figure 9) spot size temperature, °C; wall thickness
37.37 mean radiant temperature, °C torque, N·m velocity, m/s; volume width mass flow rate, kg/s mass of liquid transferred through unit area per unit time, kg/(s·m2) wv = mass of vapor diffusing through unit area per unit time, mg/(s·m2) X = variable; velocity of stream, m/s tr T V W w wl
= = = = = =
Greek = systematic (bias) error; ratio of diameters D2/D1 for venturi and sharp-edge orifice and d/D for flow nozzle p = pressure differential across the pump, Pa = deviation = random error; emissivity (0.95 for black globe) = tilt angle, ° y = moisture content = mean; vapor permeability, mg/(s·m·kPa) = density, kg/m3 c = volumetric specific heat capacity = shaft angular velocity, rev/s
Subscripts 1 2 a b c e eff g h i k s true
= = = = = = = = = = = = =
entering conditions; state 1 throat conditions; state 2 air bath cross-sectional equivalent of stream velocity effective globe hydraulic pertaining to variable X reading number average of emergent liquid column of n degrees true
STANDARDS ASA. 2011. Reference quantities for acoustical levels. ANSI Standard S1.81989 (R2011). Acoustical Society of America, New York. ASA. 2010. Measurement of sound pressure levels in air. ANSI Standard S1.13-2005 (R2010). Acoustical Society of America, New York. ASA. 2011. Specification and verification procedures for sound calibrators. ANSI Standard S1.40-2006 (R2011). Acoustical Society of America, New York. ASA. 2015. Guide to the mechanical mounting of accelerometers. ANSI Standard S2.61-1989 (R2015). Acoustical Society of America, New York. ASA. 2016. Statistical methods for determining and verifying stated noise emission values of machinery and equipment. ANSI Standard S12.31985 (R2016). Acoustical Society of America, New York. ASA. 2013. Methods for determining the insertion loss of outdoor noise barriers. ANSI Standard S12.8-1998 (R2013). Acoustical Society of America, New York. ASA. 2016. Method for the designation of sound power emitted by machinery and equipment. ANSI Standard S12.23-1989 (R2016). Acoustical Society of America, New York. ASHRAE. 2013. Standard methods for temperature measurement. ANSI/ ASHRAE Standard 41.1-2013. ASHRAE. 1987. Standard methods for laboratory air flow measurement. Standard 41.2-1987 (RA 1992). ASHRAE. 2014. Standard methods for pressure measurement. ANSI/ ASHRAE Standard 41.3-2014. ASHRAE. 2015. Standard methods for measurement of proportion of lubricant in liquid refrigerant. ANSI/ASHRAE Standard 41.4-2015. ASHRAE. 2014. Standard methods for humidity measurement. ANSI/ ASHRAE Standard 41.6-2014. ASHRAE. 2015. Method of test for measurement of flow of gas. ANSI/ ASHRAE Standard 41.7-2015.
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37.38 ASHRAE. 2016. Standard methods of measurement of flow of liquids in pipes using orifice flowmeters. ANSI/ASHRAE Standard 41.8-2016. ASHRAE. 2011. Standard methods for volatile-refrigerant mass flow measurements using calorimeters. ANSI/ASHRAE Standard 41.9-2011. ASHRAE. 2014. Standard methods for power measurement. ANSI/ ASHRAE Standard 41.11-2014. ASHRAE. 2007. Laboratory methods of testing fans for aerodynamic performance rating. ANSI/ASHRAE Standard 51-07, also ANSI/AMCA Standard 210-07. ASHRAE. 2010. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2010. ASHRAE. 2010. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2010. ASHRAE. 1997. Laboratory method of testing to determine the sound power in a duct. ANSI/ASHRAE Standard 68-1997, also ANSI/AMCA Standard 330-97. ASHRAE. 2008. Measurement, testing, adjusting, and balancing of building HVAC systems. ANSI/ASHRAE Standard 111-2008. ASHRAE. 2014. Engineering analysis of experimental data. Guideline 22010 (RA2014). ASME. 2013. Pressure gauges and gauge attachments. ANSI/ASME Standard B40.100-2013. American Society of Mechanical Engineers, New York. ASME. 2014. Glossary of terms used in the measurement of fluid flow in pipes. ANSI/ASME Standard MFC-1-2014. American Society of Mechanical Engineers, New York. ASME. 2013. Measurement uncertainty for fluid flow in closed conduits. ANSI/ASME Standard MFC-2M-1983 (RA13). American Society of Mechanical Engineers, New York. ASME. 2004. Measurement of fluid flow in pipes using orifice, nozzle, and venturi. Standard MFC-3M-2004. American Society of Mechanical Engineers, New York. ASME. 2011. Measurement of liquid flow in closed conduits by weighing methods. ANSI/ASME Standard MFC-9M-1988 (RA11). American Society of Mechanical Engineers, New York. ASME. 2011. Method for establishing installation effects on flowmeters. ANSI/ASME Standard MFC-10M-2000 (RA11). American Society of Mechanical Engineers, New York. ASME. 2013. Test uncertainty. ANSI/ASME Standard PTC 19.1-2013. American Society of Mechanical Engineers, New York. ASME. 1974. Temperature measurement. ANSI/ASME Standard PTC 19.31974 (RA98). American Society of Mechanical Engineers, New York. ASME. 2013. Flow measurement. ANSI/ASME Standard PTC 19.5-2004 (RA13). American Society of Mechanical Engineers, New York. ASTM. 2010. Standard test method for steady-state heat flux measurements and thermal transmission properties by means of the guarded-hot-plate apparatus. Standard C177-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test method for steady-state heat transfer properties of pipe insulation. Standard C335-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test method for steady-state thermal transmission properties by means of the heat flow meter apparatus. Standard C518-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard test method for thermal performance of building assemblies by means of a calibrated hot box. Standard C976-11. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2013. Standard practice for in-situ measurement of heat flux and temperature on building envelope components. Standard C1046-95 (R2013). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2015. Standard practice for thermographic inspection of insulation installations in envelope cavities of frame buildings. Standard C106011a (R2015). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2013. Standard practice for determining thermal resistance of building envelope components from the in-situ data. Standard C1155-95 (R2013). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard test method for thermal performance of building materials and envelope assemblies by means of a hot box apparatus. Standard C1363-11. American Society for Testing and Materials, West Conshohocken, PA.
2017 ASHRAE Handbook—Fundamentals (SI) ASTM. 2012. Standard guide for using indoor carbon dioxide concentrations to evaluate indoor air quality and ventilation. Standard D6245-12. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2010. Standard test methods for water vapor transmission of materials. Standard E96/E96M-10. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2012. Standard practice for maintaining constant relative humidity by means of aqueous solutions. Standard E104-02 (2012). American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2012. Standard specification and temperature-electromotive force (emf) tables for standardized thermocouples. Standard E230/E230M-12. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2011. Standard test method for determining air change in a single zone by means of a tracer gas dilution. Standard E741-2011. American Society for Testing and Materials, West Conshohocken, PA. ASTM. 2012. Occupational health and safety; protective clothing. (79 standards.) American Society for Testing and Materials, West Conshohocken, PA. IEEE. 2004. Standard test procedure for polyphase induction motors and generators. Standard 112-2004. Institute of Electrical and Electronics Engineers, New York. IEEE. 2012. IEEE standard for Ethernet. Standard 802.3. Institute of Electrical and Electronics Engineers, New York. ISO. 2003 Measurement of fluid flow by means of pressure differential devices inserted in circular cross-section conduits running full. Standard 5167:2003. International Organization for Standardization, Geneva. ISO. 2007. Industrial fans—Performance testing using standardized airways. Standard 5801:2007. International Organization for Standardization, Geneva. ISO. 1998. Ergonomics of the thermal environment—Instruments for measuring physical quantities. Standard 7726:1998. International Organization for Standardization, Geneva. ISO. 2005. Ergonomics of the thermal environment—Analytical determination and interpretation of thermal comfort using calculation of the PMV and PPD indices and local thermal comfort criteria. Standard 7730:2005. International Organization for Standardization, Geneva. ISO. 2004. Ergonomics of the thermal environment—Determination of metabolic rate. Standard 8996:2004. International Organization for Standardization, Geneva. ISO. 2007. Ergonomics of the thermal environment—Estimation of thermal insulation and water vapour resistance of a clothing ensemble. Standard 9920:2007. International Organization for Standardization, Geneva.
REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. Abernethy, R.B., R.P. Benedict, and R.B. Dowdell. 1985. ASME measurement uncertainty. Transactions of ASME 107:161-164. ACGIH. 2001. Air sampling instruments for evaluation of atmospheric contaminants, 9th ed. American Conference of Governmental Industrial Hygienists, Cincinnati, OH. Ambrose, J.H., L.C. Chow, and J.E. Beam. 1990. Capillary flow properties of mesh wicks. AIAA Journal of Thermophysics 4:318-324. Amdur, E.J. 1965. Two-pressure relative humidity standards. In Humidity and moisture, vol. 3, p. 445. Reinhold, New York. ASTM. 1993. Manual on the use of thermocouples in temperature measurement. Manual 12. American Society for Testing and Materials, West Conshohocken, PA. Bahnfleth, W.P., G.K. Yuill, and B.W. Lee. 1999. Protocol for field testing of tall buildings to determine envelope air leakage rates. ASHRAE Transactions 105(2):27-38. Baron, P.A., and K. Willeke. 2001. Aerosol measurement. Wiley, New York. Bedford, T., and C.G. Warmer. 1935. The globe thermometer in studies of heating and ventilating. Journal of the Institution of Heating and Ventilating Engineers 2:544. Benedict, R.P. 1984. Fundamentals of temperature, pressure and flow measurements, 3rd ed. John Wiley & Sons, New York. Bentz, D.P., and J.W. Martin. 1987. Using the computer to analyze coating defects. Journal of Protective Coatings and Linings 4(5).
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Measurement and Instruments Bristow, K.L., G.S. Campbell, and K. Calissendorff. 1993. Test of heat-pulse probe for measuring changes in soil water content. Soil Science Society of America Journal 57:930-934. Bristow, K.L, G.J. Kluitenberg, and R. Horton. 1994. Measurement of soil thermal properties with a dual-probe heat-pulse technique. Soil Science Society of America Journal 58:1288-1294. Brown, K.K., H.W. Coleman, and W.G. Steele. 1998. A methodology for determining experimental uncertainties in regressions. ASME Journal of Fluids Engineering, Transactions of ASME 120:445-456. Bruce, R.R., and A. Klute. 1956. The measurement of soil moisture diffusivity. Proceedings of the Soil Science Society of America 20:458-462. Burch, D.M. 1980. Infrared audits of roof heat loss. ASHRAE Transactions 86(2). Burch, D.M., and C.M. Hunt. 1978. Retrofitting an existing residence for energy conservation—An experimental study. Building Science Series 105. National Institute of Standards and Technology, Gaithersburg, MD. Burch, D.M., W.C. Thomas, and A.H. Fanney. 1992. Water vapor permeability measurements of common building materials. ASHRAE Transactions 98(2):486-494. Burns, G.W., M.G. Scroger, G.F. Strouse, M.C. Croarkin, and W.F. Guthrie. 1992. Temperature-electromotive force reference functions and tables for the letter-designated thermocouple types based on the ITS-90. NIST Monograph 175. U.S. Government Printing Office, Washington, D.C. Campbell, G.S., C. Calissendorff, and J.H. Williams. 1991. Probe for measuring soil specific heat using a heat-pulse method. Soil Science Society of America Journal 55:291-293. Carotenuto, A., F. Fucci, and G. LaFianzi. 1991. Adsorption phenomena in porous media in the presence of moist air. International Journal of Heat and Mass Transfer 18:71-81. Coleman, H.W., and W.G. Steele. 1995. Engineering application of experimental uncertainty analysis. AIAA Journal 33(10):1888-1896. Coleman, H.W., and W.G. Steele. 2009. Experimentation, validation and uncertainty analysis for engineers, 3rd ed. John Wiley & Sons, New York. Considine, D.M. 1985. Process instruments and controls handbook, 3rd ed. McGraw-Hill, New York. Cox, J.E., and C.R. Miro. 1997. Aircraft cabin air quality. ASHRAE Journal 22. Cunningham, M.J., and T.J. Sprott. 1984. Sorption properties of New Zealand building materials. Building Research Association of New Zealand Research Report 45, Judgeford. DeCarlo, J.P. 1984. Fundamentals of flow measurement. Instrumentation Society of America, Research Triangle Park, NC. D’Eustachio, D., and R.E. Schreiner. 1952. A study of transient heat method for measuring thermal conductivity. ASHVE Transactions 58:331. DeWitt, D.P., and G.D. Nutter. 1988. Theory and practice of radiation thermometry. John Wiley & Sons, New York. Fanger, P.O. 1982. Thermal comfort. Robert E. Krieger, Malabar, FL. Fanney, A.H., W.C. Thomas, D.M. Burch, and L.R. Mathena. 1991. Measurements of moisture diffusion in building materials. ASHRAE Transactions 97:99-113. Freitas, V., P. Crausse, and V. Abrantes. 1991. Moisture diffusion in thermal insulating materials. In Insulation materials: Testing and applications, vol. 2. ASTM Special Technical Publication STP 1116. American Society for Testing and Materials, West Conshohocken, PA. Gagge, A.P., J.A.J. Stolwijk, and Y. Nishi. 1971. An effective temperature scale based on a simple model of human physiological regulatory response. ASHRAE Transactions 77(1). Goldstein, R.J. 1978. Application of aerial infrared thermography. ASHRAE Transactions 84(1). Greenspan, L. 1977. Humidity fixed points of binary saturated aqueous solutions. Journal of Research of the National Bureau of Standards 81A: 89-95. Greenspan, L., and A. Wexler. 1968. An adiabatic saturation psychrometer. Journal of Research of the National Bureau of Standards 72C(1):33. Gummerson, R.J., C. Hall, W.D. Hoff, R. Hawkes, G.N. Holland, and W.S. Moore. 1979. Unsaturated water flow within porous materials observed by NMR imaging. Nature 281:56-57. Hasegawa, S. 1976. The NBS two-pressure humidity generator, mark 2. Journal of Research of the National Bureau of Standards 81A:81. Hickman, C., B.T. Beck, and B. Babin. 2012. Determining the effects of duct fittings on volumetric air flow measurements. Final Report, ASHRAE Research Project RP-1245.
37.39 Hickman, C., B.T. Beck, and B. Babin. 2015a. Effect of fittings on volumetric airflow measurements (RP-1245): Single-path duct disturbances. Science and Technology for the Built Environment 21:190-206. Hickman, C., B.T. Beck, and B. Babin, 2015b. Effect of fittings on volumetric airflow measurements (RP-1245): Multiple-path (tee) duct disturbances. Science and Technology for the Built Environment 21:957-975. Holman, J.P. 2001. Experimental methods for engineers, 7th ed., pp. 383389. McGraw-Hill, New York. Hooper, F.C., and S.C. Chang. 1953. Development of thermal conductivity probe. ASHVE Transactions 59:463. Hooper, F.C., and F.C. Lepper. 1950. Transient heat flow apparatus for the determination of thermal conductivity. ASHVE Transactions 56:309. Hudson, R.D., Jr. 1969. Infrared system engineering. John Wiley & Sons, New York. IES. 2011. Lighting handbook, 10th ed. Illuminating Engineering Society of North America, New York. Kumaran, M.K., and M. Bomberg. 1985. A gamma-spectrometer for determination of density distribution and moisture distribution in building materials. Proceedings of the International Symposium on Moisture and Humidity, Washington, D.C., pp. 485-490. Kusuda, T. 1965. Calculation of the temperature of a flat-plate wet surface under adiabatic conditions with respect to the Lewis relation. In Humidity and moisture, vol. 1, p. 16. Reinhold, New York. Liptak, B.G., ed. 1972. Instrument engineers handbook, vol. 1. Chilton, Philadelphia, PA. Lodge, J.P., ed. 1989. Methods of air sampling and analysis, 3rd ed. Lewis Publishers, MI. Lundgren, D.A., M. Lippmann, F.S. Harris, Jr., W.H. Marlow, W.E. Clark, and M.D. Durham, eds. 1979. Aerosol measurement. University Presses of Florida, Gainesville. Mack, R.T. 1986. Energy loss profiles: Foundation for future profit in thermal imager sales and service. Proceedings of the 5th Infrared Information Exchange, Book 1, AGEMA Infrared Systems, Secaucus, NJ. Madding, R. 1989. Infrared thermography. McGraw-Hill, New York. Madsen, T.L. 1976. Thermal comfort measurements. ASHRAE Transactions 82(1). Mattingly, G.E. 1984. Workshop on fundamental research issues in orifice metering. GRI Report 84/0190. Gas Research Institute, Chicago. Mattingly, G.E. 1992. The characterization of a piston displacement-type flowmeter calibration facility and the calibration and use of pulsed output type flowmeters. Journal of Research of the National Institute of Standards and Technology 97(5):509. McCullough, E.A., B.W. Jones, and J. Huck. 1985. A comprehensive data base for estimating clothing insulation. ASHRAE Transactions 92:29-47. McLean, R.C., G.H. Galbraith, and C.H. Sanders. 1990. Moisture transmission testing of building materials and the presentation of vapour permeability values. Building Research and Practice 18(2):82-103. Mease, N.E., W.G. Cleveland, Jr., G.E. Mattingly, J.M. Hall. 1992. Air speed calibrations at the National Institute of Standards and Technology. Proceedings of the 1992 Measurement Science Conference, Anaheim, CA. Miller, R.W. 1983. Measurement engineering handbook. McGraw-Hill, New York. Nagda, N.L., and H.E. Rector. 2001. Instruments and methods for measuring indoor air quality. In Indoor air quality handbook, pp. 51.1-51.37. J.D. Spengler, J.M. Samet, and J.F. McCarthy, eds. McGraw-Hill. NIST. 1976. Liquid-in-glass thermometry. NIST Monograph 150. National Institute of Standards and Technology, Gaithersburg, MD. NIST. 1986. Thermometer calibrations. NIST Monograph 174. National Institute of Standards and Technology, Gaithersburg, MD. Nottage, H.B., J.G. Slaby, and W.P. Gojsza. 1952. A smoke-filament technique for experimental research in room air distribution. ASHVE Transactions 58:399. Olesen, B.W. 1985. A new and simpler method for estimating the thermal insulation of a clothing ensemble. ASHRAE Transactions 92:478-492. Olesen, B.W., J. Rosendahl, L.N. Kalisperis, L.H. Summers, and M. Steinman. 1989. Methods for measuring and evaluating the thermal radiation in a room. ASHRAE Transactions 95(1). Paljak, I., and B. Pettersson. 1972. Thermography of buildings. National Swedish Institute for Materials Testing, Stockholm. Parmelee, G.V., and R.G. Huebscher. 1946. The shielding of thermocouples from the effects of radiation. ASHVE Transactions 52:183.
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2017 ASHRAE Handbook—Fundamentals (SI)
Persily, A., and S.J. Emmerich. 2001. State-of-the-art review of CO2 demand control ventilation and application. NIST IR6729. National Institute of Standards and Technology, Gaithersburg, MD. Prazak, J., J. Tywoniak, F. Peterka, and T. Slonc. 1990. Description of transport of liquid in porous media—A study based on neutron radiography data. International Journal of Heat and Mass Transfer 33:1105-1120. Quenard, D., and H. Sallee. 1989. A gamma-ray spectrometer for measurement of the water diffusivity of cementitious materials. Proceedings of the Materials Research Society Symposium, vol. 137. Quinn, T.J. 1990. Temperature, 2nd ed. Academic Press, New York. Raffel, M., C. Willert, and J. Kompenhans. 1998. Particle image velocimetry: A practical guide. Springer. Richards, R.F., D.M. Burch, W.C. Thomas. 1992. Water vapor sorption measurements of common building materials. ASHRAE Transactions 98(1). Richardson, L. 1965. A thermocouple recording psychrometer for measurement of relative humidity in hot, arid atmosphere. In Humidity and moisture, vol. 1, p. 101. Reinhold, New York. Schooley, J.F. 1986. Thermometry. CRC, Boca Raton, FL. Schooley, J.F., ed. 1992. Temperature: Its measurement and control in science and in industry, vol. 6. American Institute of Physics, New York. Seely, R.E. 1955. A circuit for measuring the resistance of energized A-C windings. AIEE Transactions, p. 214. Shafer, M.R. 1961. Performance characteristics of turbine flowmeters. Proceedings of the Winter Annual Meeting, Paper 61-WA-25. American Society of Mechanical Engineers, New York. Straube, J.F. 1998. Moisture control and enclosure wall systems. Ph.D. dissertation, University of Waterloo. Waterloo, Ontario, Canada. TIA. 1997. Interface between data terminal equipment and data circuitterminating equipment employing serial binary data interchange. Standard TIA-232. Telecommunication Industry Alliance, Revision F, October. TIA. 1998. Electrical characteristics of generators and receivers for use in balanced digital multipoint systems. Standard TIA-485. Telecommunication Industry Alliance, Revision A, March. TIA. 2006. Application guidelines for TIA/EIA-485-A. Standard TSB-89. Telecommunication Industry Alliance, Revision A, January 2006. Till, C.E., and G.E. Handegord. 1960. Proposed humidity standard. ASHRAE Transactions 66:288. Tobiasson, W., and C. Korhonen. 1985. Roofing moisture surveys: Yesterday, today, and tomorrow. Proceedings of the Second International Symposium on Roofing Technology, Gaithersburg, MD. Tveit, A. 1966. Measurement of moisture sorption and moisture permeability of porous materials. Report 45. Norwegian Building Research Institute, Oslo. Vernon, H.M. 1932. The globe thermometer. Proceedings of the Institution of Heating and Ventilating Engineers, vol. 39, p. 100. Wentzel, J.D. 1961. An instrument for measurement of the humidity of air. ASHRAE Journal 11:67.
Wile, D.D. 1947. Air flow measurement in the laboratory. Refrigerating Engineering 6:515. Woodring, E.D. 1969. Magnetic turbine flowmeters. Instruments and Control Systems 6:133. Worrall, R.W. 1965. Psychrometric determination of relative humidities in air with dry-bulb temperatures exceeding 212°F. In Humidity and Moisture, vol. 1, p. 105. Reinhold, New York. Yang, H., Y. Luo, and T. Zhang. 2015. A data sorting method for rapid measurement of moisture content in porous insulation materials. Proceedings of ISHVAC-COBEE 2015, Tianjin, China.
BIBLIOGRAPHY Beranek, L.L. 1988. Acoustical measurements. Published for the Acoustical Society of America by the American Institute of Physics, New York. Beranek, L.L. 1989. Noise and vibration control. Institute of Noise Control Engineering, Poughkeepsie, NY. Cohen, E.R. 1990. The expression of uncertainty in physical measurements. 1990 Measurement Science Conference Proceedings, Anaheim, CA. Contemporary Controls. 1999. Understanding EIA-485 networks. The Extension: A Technical Supplement to Control Network 1(1). www.ccontrols .com/pdf/ExtV1N1.pdf. EPA. 1991. Introduction to indoor air quality: A self-paced learning module. EPA/400/3-91/002, U.S. Environmental Protection Agency, Washington, D.C. Harris, C.M. 1987. Shock and vibration handbook, 3rd ed. McGraw-Hill, New York. IEEE. 1987. Standard digital interface for programmable instrumentation. ANSI/IEEE Standard 488.1-87 (R 1994). Institute of Electrical and Electronics Engineers, Piscataway, NJ. Jones, J.L., and A.M. Flynn, 1993. Mobile robots, A.K. Peters, Wellesley, MA. De Silva, C.W. 1989. Control sensors and actuators. Prentice Hall, Englewood Cliffs, NJ. Lord, H.W., W.S. Gatley, and H.A. Evensen. 1987. Noise control for engineers. Krieger, Melbourne, FL. Morrison, R. 1986. Grounding and shielding techniques in instrumentation, 3rd ed. John Wiley & Sons, New York. Spitzer, D.W., ed. 1991. Flow measurement. Instrumentation Society of America, Research Triangle Park, NC. Steele, W.G., R.A. Ferguson, R.P. Taylor, and H.W. Coleman. 1994. Comparison of ANSI/ASME and ISO models for calculation of uncertainty. ISA Transactions 33:339-352. Tilford, C.R. 1992. Pressure and vacuum measurements. In Physical methods of chemistry, 2nd ed., vol. 6, pp. 106-173. John Wiley & Sons, New York. Universal Serial Bus Specification. 2013. Universal serial bus revision 3.1 specification, section 5.5.7. July. Universal Serial Bus Specification. 2000. Universal serial bus revision 2.0 specification, section 7.1.6. April.
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ABBREVIATIONS AND SYMBOLS Abbreviations for Text, Drawings, and Computer Programs ....................................................... 37.1 Letter Symbols............................................................................................................................... 37.1 Dimensionless Numbers ................................................................................................................ 37.4 Mathematical Symbols .................................................................................................................. 37.4 Subscripts ...................................................................................................................................... 37.5 Graphical Symbols for Drawings ................................................................................................ 37.5 Piping System Identification ....................................................................................................... 37.10
HIS chapter contains information about abbreviations and symbols for HVAC&R engineers. Abbreviations are shortened forms of names and expressions used in text, drawings, and computer programs. This chapter discusses conventional English-language abbreviations that may be different in other languages. A letter symbol represents a quantity or a unit, not its name, and is independent of language. Because of this, use of a letter symbol is preferred over abbreviations for unit or quantity terms. Letter symbols necessary for individual chapters are defined in the chapters where they occur. Abbreviations are never used for mathematical signs, such as the equality sign (=) or division sign (/), except in computer programming, where the abbreviation functions as a letter symbol. Mathematical operations are performed only with symbols. Abbreviations should be used only where necessary to save time and space; avoid their use in documents circulated in foreign countries. Graphical symbols in this chapter of piping, ductwork, fittings, and in-line accessories can be used on scale drawings and diagrams. Identifying piping by legend and color promotes greater safety and lessens the chance of error in emergencies. Piping identification is now required throughout the United States by the Occupational Safety and Health Administration (OSHA) for some industries and by many federal, state, and local codes.
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T
ABBREVIATIONS FOR TEXT, DRAWINGS, AND COMPUTER PROGRAMS Table 1 gives some abbreviations, as well as others commonly found on mechanical drawings and abbreviations (symbols) used in computer programming. Abbreviations specific to a single subject are defined in the chapters in which they appear. Additional abbreviations used on drawings can be found in the section on Graphical Symbols for Drawings.
Computer Programs The abbreviations (symbols) used for computer programming for the HVAC&R industries have been developed by ASHRAE Technical Committee 1.5, Computer Applications. These symbols identify computer variables, subprograms, subroutines, and functions commonly applied in the industry. Using these symbols enhances comprehension of the program listings and provides a clearly defined nomenclature in applicable computer programs. Some symbols have two or more options listed. The longest abbreviation is preferred and should be used if possible. However, it is sometimes necessary to shorten the symbol to further identify the variable. For instance, the area of a wall cannot be defined as WALLAREA because some computer languages restrict the number The preparation of this chapter is assigned to TC 1.6, Terminology.
38.1 Copyright © 2017, ASHRAE
of letters in a variable name. Therefore, a shorter variable symbol is applied, and WALLAREA becomes WALLA or WAREA. Most modern computer programming languages do not have the character limitations of older computer languages, but the limitations still exist in some older building automation controllers. It is good programming practice to include the complete name of each variable and to define any abbreviations in the comments section at the beginning of each module of code. Abbreviations should be used to help clarify the variables in an equation and not to obscure the readability of the code. In Table 1, the same symbol is sometimes used for different terms. This liberty is taken because it is unlikely that the two terms would be used in the same program. If such were the case, one of the terms would require a suffix or prefix to differentiate it from the other.
LETTER SYMBOLS Letter symbols include symbols for physical quantities (quantity symbols) and symbols for the units in which these quantities are measured (unit symbols). Quantity symbols, such as I for electric current, are listed in this chapter and are printed in italic type. A unit symbol is a letter or group of letters such as mm for millimetre or a special sign such as ° for degrees, and is printed in Roman type. Subscripts and superscripts are governed by the same principles. Letter symbols are restricted mainly to the English and Greek alphabets. Quantity symbols may be used in mathematical expressions in any way consistent with good mathematical usage. The product of two quantities, a and b, is indicated by ab. The quotient is a/b, or ab1. To avoid misinterpretation, parentheses must be used if more than one slash (/) is used in an algebraic term; for example, (a/b)/c or a/(b/c) is correct, but not a/b/c. Subscripts and superscripts, or several of them separated by commas, may be attached to a single basic letter (kernel), but not to other subscripts or superscripts. A symbol that has been modified by a superscript should be enclosed in parentheses before an exponent is added (Xa)3. Symbols can also have alphanumeric marks such as (prime), + (plus), and * (asterisk). More detailed information on the general principles of letter symbol standardization are in standards listed at the end of this chapter. The letter symbols, in general, follow these standards, which are out of print: Y10.3M Y10.4-82
Letter Symbols for Mechanics and Time-Related Phenomena Letter Symbols for Heat and Thermodynamics
Other symbols chosen by an author for a physical magnitude not appearing in any standard list should be ones that do not already have different meanings in the field of the text.
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38.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 1 Abbreviations for Text, Drawings, and Computer Programs
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Term
Text
Drawings
Program
above finished floor absolute accumulat(e, -or) air condition(-ing, -ed) air-conditioning unit(s) air-handling unit air horsepower alteration alternating current altitude ambient American National Standards Institute1 American wire gage ampere (amp, amps) angle angle of incidence apparatus dew point approximate area atmosphere average azimuth azimuth, solar azimuth, wall
— abs acc — — — ahp altrn ac alt amb
AFF ABS ACCUM AIR COND ACU AHU AHP ALTRN AC ALT AMB
— ABS ACCUM — ACU AHU AHP — AC ALT AMB
ANSI AWG amp — — adp approx. — atm avg az — —
ANSI AWG AMP — — ADP APPROX — ATM AVG AZ — —
— — AMP, AMPS ANG ANGI ADP — A — AVG AZ SAZ WAZ
barometer(-tric) bill of material boiling point Brown & Sharpe wire gage
baro b/m bp B&S
BARO BOM BP B&S
— — BP —
Celsius center to center circuit clockwise coefficient coefficient, valve flow coil compressor condens(-er, -ing, -ation) conductance conductivity conductors, number of (3) contact factor cooling load counterclockwise cubic centimetre cubic metre
°C c to c ckt cw coeff. Cv — cprsr cond — cndct 3/c — clg load ccw cm3 m3
°C C TO C CKT CW COEF Cv — CMPR COND — CNDCT 3/c — CLG LOAD CCW CC CU M
°C — CKT — COEF CV COIL CMPR COND C K — CF CLOAD — CC CU M
decibel degree density depth or deep dew-point temperature diameter diameter, inside diameter, outside difference or delta diffuse radiation direct current direct radiation dry dry-bulb temperature
dB deg. or ° dens dp dpt dia. ID OD diff., — dc dir radn — dbt
DB DEG or ° DENS DP DPT DIA ID OD DIFF
DB DEG RHO DPTH DPT DIA ID OD D, DELTA DFRAD DC DIRAD DRY DB, DBT
effectiveness effective temperature2 efficiency efficiency, fin efficiency, surface electromotive force
— ET* eff — — emf
DC DIR RADN DBT ET* EFF EMF
EFT ET EFF FEFF SEFF —
Term
Text
Drawings
Program
elevation entering entering water temperature entering air temperature enthalpy entropy equivalent direct radiation evaporat(-e, -ing, -ed, -or) expansion
elev. entr EWT EAT — — edr evap exp
EL ENT EWT EAT — — EDR EVAP EXP
ELEV ENT EWT EAT H S — EVAP XPAN
face area face to face face velocity factor, correction factor, friction fan film coefficient,3 inside film coefficient,3 outside flow rate, air flow rate, fluid flow rate, gas freezing point frequency
fa f to f fvel — — — — — — — — fp Hz
FA F to F FVEL — — — — — — — — FP HZ
FA — FV CFAC, CFACT FFACT, FF FAN FI, HI FO, HO QAR, QAIR QFL QGA, QGAS FP —
gage or gauge gram gravitational constant greatest temp difference
ga g G GTD
GA g G GTD
GA, GAGE G G GTD
heat heater heat gain heat gain, latent heat gain, sensible heat loss heat transfer heat transfer coefficient height high-pressure steam high-temperature hot water hour(s) humidity, relative humidity ratio
— — HG LHG SHG — — U hgt hps hthw h rh W
— — HG LHG SHG — — U HGT HPS HTHW HR RH W
HT HTR HG, HEATG HGL HGS HL, HEATL Q U HGT, HT HPS HTHW HR RH W
incident angle indicated kilowatt International Pipe Std iron pipe size
— IkW IPS ips
— IkW IPS IPS
INANG — — —
joule
J
J
J
kelvin kilograms kilojoules kilometres per hour kilopascals kilowatt kilowatt hour
K kg kJ km/h kPa kW kWh
K kg kJ km/h kPa kW KWH
K KG KJ KPH KPA KW KWH
latent heat least mean temp. difference4 least temp. difference4 leaving air temperature leaving water temperature length liquid load-sharing (hybrid) HVAC system litre litres per second logarithm (natural) logarithm to base 10
LH LMTD LTD lat lwt lg liq
LH LMTD LTD LAT LWT LG LIQ
LH, LHEAT LMTD LTD LAT LWT LG, L LIQ
LSHVAC L L/s ln log
LSHVAC L L/s LN LOG
LSHVAC L LPS LN LOG
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Abbreviations and Symbols
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Term
Text
38.3 Drawings
Program
low-pressure steam low-temp. hot water
lps lthw
LPS LTHW
LPS LTHW
Mach number mass flow rate maximum mean effective temp. mean temp. difference medium-pressure steam medium-temp. hot water mercury metre metres per second millilitres per second mL/s standard minimum
Mach mfr max. MET MTD mps mthw Hg m m/s mL/s mL/sS min.
MACH MFR MAX MET MTD MPS MTHW HG m m/s mL/s mL/sS MIN
— MFR MAX MET MTD MPS MTHW HG M M/S MLPS MLPSS MIN
noise criteria normally open normally closed not applicable not in contract not to scale number number of circuits number of tubes
NC no nc na nic — no. — —
NC NO NC N/A NIC NTS NO — —
— — — — — — N, NO NC NT
outside air
oa
OA
OA
parts per million pascal Pa (absolute) Pa (gage) percent phase (electrical) pipe pressure pressure, barometric pressure, critical pressure, dynamic (velocity) pressure drop or difference pressure, static pressure, vapor primary
ppm Pa Pa (abs) Pa (gage) % ph — — baro pr — vp PD sp vap pr pri
PPM Pa Pa A Pa G % PH — PRESS BARO PR — VP PD SP VAP PR PRI
PPM PA PAA PAG PCT — PIPE PRES, P BP CRIP VP PD, DELTP SP VAP PRIM
radian radiat(-e, -or) radiant panel radiation radius receiver recirculate refrigerant (12, 22, etc.) relative humidity resist(-ance, -ivity, -or) return air revolutions revolutions per minute revolutions per second roughness
— — RP — — rcvr recirc. R-12, R-22 rh res ra rev rpm rps rgh
— RAD RP RADN — RCVR RECIRC R12, R22 RH RES RA REV RPM RPS RGH
RAD — RP RAD R REC RCIR, RECIR R12, R22 RH RES, OHMS RA REV RPM RPS RGH, E
safety factor saturation Saybolt seconds Furol Saybolt seconds Universal sea level second sensible heat sensible heat gain sensible heat ratio
sf sat. ssf ssu sl s SH SHG SHR
SF SAT SSF SSU SL s SH SHG SHR
SF SAT SSF SSU SE SEC SH SHG SHR
Term
Text
Drawings
Program
shading coefficient solar specification specific heat sp ht at constant pressure sp ht at constant volume specific volume square standard standard time meridian static pressure suction summ(-er, -ary, -ation) supply supply air surface surface, dry surface, wet system
— — spec sp ht cp cv sp vol sq. std — SP suct. — sply sa — — — —
— — SPEC SP HT cp cv SP VOL SQ STD — SP SUCT — SPLY SA — — — —
SC SOL — C CP CV V, CVOL SQ STD STM SP SUCT, SUC SUM SUP, SPLY SA SUR, S SURD SURW SYS
tabulat(-e, -ion) tee temperature temperature difference temperature entering temperature leaving thermal conductivity thermal expansion coeff. thermal resistance thermocouple thermostat thick(-ness) total total heat transmissivity
tab — temp. TD, t TE TL k — R tc T STAT thkns — tot ht —
TAB — TEMP TD TE TL K — R TC T STAT THKNS — TOT HT —
TAB TEE T, TEMP TD, TDIF TE, TENT TL, TLEA K TXPC RES, R TC, TCPL T STAT THK TOT — TAU
U-factor unit
— —
— —
U UNIT
vacuum valve vapor proof variable variable air volume velocity velocity, wind ventilation, vent vertical viscosity volt volt ampere volume volumetric flow rate
vac v vap prf var VAV vel. w vel. vent vert. visc V VA vol. —
VAC V VAP PRF VAR VAV VEL W VEL VENT VERT VISC V VA VOL —
VAC VLV — VAR VAV VEL, V W VEL VENT VERT MU, VISC E, VOLTS VA VOL VFR
wall water watt wet bulb wet-bulb temperature width wind wind direction wind pressure
— — W wb wbt — — wdir wpr
— — W WB WBT — — WDIR WPR
W, WAL WTR WAT, W WB WBT WI WD WDIR WP, WPRES
year
yr
YR
YR
zone
z
Z
Z, ZN
1Abbreviations
of most proper names use capital letters in both text and drawings. 2The asterisk (*) is used with ET*, effective temperature, as in Chapter 9 of this volume. 3These are surface heat transfer coefficients. 4Letter L also used for Logarithm of these temperature differences in computer programming.
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38.4
2017 ASHRAE Handbook—Fundamentals (SI) LETTER SYMBOLS
Symbol
acoustic velocity area breadth or width barometric pressure concentration specific heat specific heat at constant pressure specific heat at constant volume coefficient fluid capacity rate thermal conductance loss coefficient coefficient of performance prefix meaning differential diameter equivalent or hydraulic diameter mass diffusivity base of natural logarithms energy electrical potential film conductance (alternate for h) frequency friction factor, Darcy-Weisbach formulation fF friction factor, Fanning formulation F force Fij angle factor (radiation) g gravitational acceleration G mass velocity h heat transfer coefficient h hydraulic head h specific enthalpy ha enthalpy of dry air hD mass transfer coefficient hs enthalpy of moist air at saturation H total enthalpy I electric current k thermal conductivity k (or ) ratio of specific heats, cp /cv K proportionality constant KD mass transfer coefficient l or L length Lp sound pressure Lw sound power m or M mass M molecular weight n or N number in general N rate of rotation p or P pressure pa partial pressure of dry air ps partial pressure of water vapor in moist air pw vapor pressure of water in saturated moist air P power q time rate of heat transfer Q total heat transfer Q volumetric flow rate r radius r or R thermal resistance R gas constant s specific entropy S total entropy t temperature tm or Tm mean temperature difference T absolute temperature u specific internal energy U total internal energy U overall heat transfer coefficient v specific volume V total volume a A b B c c cp cv C C C CL CP d d or D De or Dh Dv e E E f f fD
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Description of Item
Symbol Typical Units m/s m2 m kPa kg/m3 kJ/(kg·K) kJ/(kg·K) kJ/(kg·K) — W/K W/(m2 ·K) — — — m m mm2/s — kJ V W/(m2 ·K) Hz — — N — m/s2 kg/(s·m2) W/(m2 ·K) m kJ/kg kJ/kg m/s kJ/kg kJ A W/(m·K) — — kg/(h·m2) m dB dB kg kg/kg mol — kPa kPa kPa kPa kPa kW W kJ L/s m (m2 ·K)/W J/(kg·K) kJ/(kg·K) kJ/K °C K K kJ/kg kJ W/(m2 ·K) m3/kg m3
V w W W W Ws x x x,y,z Z (or k)
Description of Item linear velocity mass rate of flow weight humidity ratio of moist air (dry air basis) work humidity ratio of moist air at saturation (dry air basis) mole fraction quality, mass fraction of vapor lengths along principal coordinate axes figure of merit absolute Seebeck coefficient absorptivity, absorptance radiation linear coefficient of thermal expansion thermal diffusivity volume coefficient of thermal expansion ratio of specific heats, cp /cv specific weight difference between values emissivity, emittance (radiation) time efficiency or effectiveness wavelength degree of saturation dynamic viscosity kinematic viscosity density reflectivity, reflectance (radiation) volume resistivity Stefan-Boltzmann constant surface tension stress time transmissivity, transmittance (radiation) relative humidity
Typical Units
g/kg g/kg — — m — V/K — 1/K 1/K — N/m3 — — s, h — nm — mPa·s m2/s kg/m3 — ·m W/(m2 ·K4) N/m N/m2 s — —
DIMENSIONLESS NUMBERS Fo Gr Gz jD jH Le M Nu Pe Pr Re Sc Sh St Str
Fourier number Grashof number Graetz number Colburn mass transfer Colburn heat transfer Lewis number Mach number Nusselt number Peclet number Prandtl number Reynolds number Schmidt number Sherwood number Stanton number Strouhal number
/L2 L32g(t)/2 wcp /kL Sh/ReSc1/3 Nu/RePr1/3 /Dv V/a hD/k GDcp /k cp/k VD/ Dv hD L/Dv h/Gcp fd/V
MATHEMATICAL SYMBOLS equal to not equal to approximately equal to greater than less than greater than or equal to less than or equal to plus minus plus or minus a multiplied by b a divided by b
= < + ± ab, a·b, a b a --- , a/b, ab–1 b
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Abbreviations and Symbols ratio of circumference of a circle to its diameter a raised to the power n square root of a infinity percent summation of natural log logarithm to base 10
38.5 an a,a % ln log
0.5
SUBSCRIPTS
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These are to be affixed to the appropriate symbols. Several subscripts may be used together to denote combinations of various states, points, or paths. Often the subscript indicates that a particular property is to be kept constant in a process. a,b,... referring to different phases, states or physical conditions of a substance, or to different substances a air a ambient b barometric (pressure) c referring to critical state or critical value c convection db dry bulb dp dew point e base of natural logarithms f referring to saturated liquid f film fg referring to evaporation or condensation F friction g referring to saturated vapor h referring to change of phase in evaporation H water vapor i referring to saturated solid i internal if referring to change of phase in melting ig referring to change of phase in sublimation k kinetic L latent m mean value M molar basis p referring to constant pressure conditions or processes p potential r refrigerant r radiant or radiation s referring to moist air at saturation s sensible s referring to isentropic conditions or processes s static (pressure) s surface t total (pressure) T referring to isothermal conditions or processes v referring to constant volume conditions or processes v vapor v velocity (pressure) w wall w water wb wet bulb 0 referring to initial or standard states or conditions 1,2,... different points in a process, or different instants of time
GRAPHICAL SYMBOLS FOR DRAWINGS Graphical symbols have been extracted from ANSI/ASHRAE Standard 134-2005. Additional symbols are from current practice and extracted from ASME Standards Y32.2.3 and Y32.2.4.
Low-pressure steam condensate Boiler blowdown Pumped condensate Vacuum pump discharge Makeup water Atmospheric vent Fuel oil Low-temperature hot water supply Medium-temperature hot water supply High-temperature hot water supply Low-temperature hot water return Medium-temperature hot water return High-temperature hot water return Compressed air Vacuum (air) Existing piping Pipe to be removed Air Conditioning and Refrigeration Refrigerant discharge Refrigerant suction Brine supply Brine return Condenser water supply Condenser water return Chilled water supply Chilled water return Fill line Humidification line Drain Hot/chilled water supply Hot/chilled water return Refrigerant liquid Heat pump water supply Heat pump water return Plumbing Sanitary drain above floor or grade Sanitary drain below floor or grade Storm drain above floor or grade Storm drain below floor or grade Condensate drain above floor or grade Condensate drain below floor or grade Vent Cold water Hot water Hot water return Gas Acid waste Drinking water supply Chemical supply pipesa Floor drain
LPC BBD PC VPD MU ATV FO(NAME) HWS MTWS HTWS HWR MTWR HTWR A(NAME) VAC (NAME)E XX (NAME) XX RD RS B BR CWS CWR CHWS CHWR FILL H D H/C S H/C R RL HPWS HPWR SAN SAN ST ST CD CD –––––––––––
G
G AW DCW (NAME)
Funnel drain, open
Fire Safety Devicesb Signal Initiating Detectors Heat (thermal)
Gas
Smoke
Flame
Radiant Panels
Piping Heating High-pressure steam Medium-pressure steam Low-pressure steam High-pressure steam condensate Medium-pressure steam condensate
HPS MPS LPS HPC MPC
a See
section on Piping Identification in this chapter. to Standard for Fire Safety Symbols, 1999 edition (NFPA Standard 170).
b Refer
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38.6
2017 ASHRAE Handbook—Fundamentals (SI)
Radiant Ceiling Panels
Valves Actuators Manual Non-rising sun Outside stem & yoke
Embedded Above ceiling
Lever
Surface mounted
Gear
Suspended
Electric Motor
Radiant Floor Panels
Solenoid
Slab on grade
Pneumatic Motor
Above subfloor
Diaphragm
Below subfloor Slab above subfloor
Valves, Special Duty Check, swing gate
Radiant Wall Panels
Check, spring Control, electric-pneumatic Control, pneumatic-electric
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Hose end drain Embedded
Coils Cooling coil
Surface mounted
Decorative
Lock shield Needle Pressure-reducing regulator Quick-opening
Heating coil
Quick-closing Safety or relief
Electrical coil
Solenoid Square-head cock
Humidifier
Unclassified (number and specify)
Fittings Valves Valves for Selective Actuators Air line
The following fittings are shown without connection notations. This reflects current practice. The symbol for the body of a fitting is the same for all types of connections, unless otherwise specified. The types of connections are often specified for a range of pipe sizes, but are shown with the fitting symbol where required. For example, an elbow would be:
Ball Flanged
Threaded
Belt & Spigot
Weldeda
Soldered
Solvent Cement
Butterfly Diaphragm Gate
Fitting Bushing
Gate, angle
Cap
Globe
Cross
Globe, angle
Elbow, 90° Elbow, 45°
Plug valve Three way
Elbow, facing toward viewer a Includes
fusion; specify type.
Symbol
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Abbreviations and Symbols Elbow, facing away from viewer
38.7 Thermometer
Elbow, base-supported Lateral
Thermometer well, only
Reducer, concentric
Thermostat Traps, steam (indicate type)
Reducer, eccentric, flat on bottom
Unit heater (indicate type)
Reducer, eccentric, flat on top
Air-Moving Devices and Components
Tee
Fans (indicate use)
Tee, facing toward viewer
Centrifugal
Tee, facing away from viewer Union, screwed Union, flanged
Piping Specialties Air vent, automatic
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Air vent, manual Air separator Pipe guide Anchor, intermediate Anchor, main Ball joint
Propeller Roof ventilator, intake Roof ventilator, exhaust Roof ventilator, louvered Vaneaxial Ductworka,b Direction of flow Duct size, where first dimension is visible duct Duct section, supply
Expansion joint Expansion loop Flexible connector Flowmeter, orifice plate with flanges Flowmeter, venturi
Duct section, return
Duct section, exhaust
Flow switch
Change of elevation rise (R) drop (D)
Hanger rod
Access doors, vertical or horizontal
Hanger spring Heat exchanger, liquid
Cowl, (gooseneck) and flashing
Heat transfer surface (indicate type) Pitch of pipe, rise (R) drop (D)
Duct lining
Pressure gage and cock
Flexible connection
Pressure switch Pump (indicate use)
Flexible duct
Pump suction diffuser
Sound attenuator
Spool piece, flanged
Terminal unit, mixing
Strainer
Terminal unit, variable volume
Strainer, blow off Strainer, duplex Tank (indicate use)
a Units of measurement are not shown here, but should be shown on drawings. The first
dimension is visible duct dimension for duct size, top dimension for grilles, and horizontal dimension for registers. bShow volumetric flow rate at each device.
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38.8
2017 ASHRAE Handbook—Fundamentals (SI) Reciprocating
Transitiona
Rotary
Turning vanes
Rotary screw
Smoke detectors
Condensers
Dampers
Air cooled
Backdraft damper Evaporative Fire damper Water cooled, (specify type) Condensing Units
Manual volume
Air cooleda Smoke damper Water cooleda
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Diffusersb
Grilles, Register and Sidewall inlet, (exhaust) outlet, registers, and grilles Sidewall outlet, registers, and grilles Rectangular four-way outlet, supply Louver and screen
Condenser-Evaporator (Cascade System)
Cooling Towers Cooling tower Spray pond Evaporatorsb
Transfer grille or louver
Finned coil Forced convection
Door grille or louver Immersion cooling unit Undercut door
Ceiling diffuser, rectangular
Plate coil Pipe coilc Liquid Chillers (Chillers only)
Round outlet
Direct expansiond
Linear outlet
Floodedd Tank, closed
Light troffer outlet Tank, open
Refrigeration Compressors
Chilling Units Absorption
Centrifugal aL
aIndicate
flat on bottom or top (FOB or FOT), if applicable. bShow volumetric flow rate at each device.
= Liquid being cooled, RL = Refrigerant liquid, RS = Refrigerant suction. manifolding. c Frequently used diagrammatically as evaporator and/or condenser with label indicating name and type. d L = Liquid being cooled, RL = Refrigerant liquid, RS = Refrigerant suction. bSpecify
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Abbreviations and Symbols Centrifugal
38.9 Valve, condenser water regulating
Auxiliary Equipment Reciprocating
Rotary screw
Filter
Controls
Strainer
Refrigerant Controls
Filter and drier
Capillary tube Expansion valve, hand Expansion valve, automatic Expansion valve, thermostatic
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Refrigerant
Scale trap Drier Vibration absorber Heat exchanger
Float valve, high side, or liquid drain valve
Oil separator
Float valve, low side
Sight glass
Thermal bulb
Fusible plug
Solenoid valve
Constant pressure valve, suction Evaporator pressure regulating valve, thermostatic, throttling
Rupture disk Receiver, high-pressure, horizontal Receiver, high-pressure, vertical
Evaporator pressure regulating valve, thermostatic, snap-action
Receiver, low-pressure
Evaporator pressure regulating valve, throttling-type, evaporator side Compressor suction valve, pressure-limiting, throttlingtype, compressor side
Intercooler
Thermosuction valve
Intercooler/desuperheater
Snap-action valve
Energy Recovery Equipment Refrigerant reversing valve Temperature or Temperature-Actuated Electrical or Flow Controls Thermostat, self-contained
Condenser, double bundle Air to Air Energy Recovery Rotary heat wheel
Thermostat, remote bulb Coil loop Sensor, temperature Pressure-reducing regulator Pressure regulator
Heat pipe Fixed plate
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38.10
2017 ASHRAE Handbook—Fundamentals (SI) Table 2 Examples of Legends
Plate fin, crossflow
HOT WATER AIR 700 kPA H.P. RETURN STEAM 700 kPA
Power Sources Motor, electric (number for identification of description in specifications) Engine (indicate fuel)
Table 3 Classification of Hazardous Materials and Designation of Colorsa Classification
Gas turbine Steam turbine Steam turbine, condensing
Electrical
Equipmenta
Symbols for electrical equipment shown on mechanical drawings are usually geometric figures with an appropriate name or abbreviation, with details described in the specifications. The following are some common examples.b
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Motor control Disconnect switch, unfused Disconnect switch, fused Time clock Automatic filter panel Lighting panel Power panel a See
ARI Standard 130 for preferred symbols of common electrical parts. each symbol if more than one; see ASME Standard Y32.4.
b Number
PIPING SYSTEM IDENTIFICATION The material in piping systems is identified to promote greater safety and lessen the chances of error, confusion, or inaction in times of emergency. Primary identification should be by means of a lettered legend naming the material conveyed by the piping. In addition to, but not instead of, lettered identification, color can be used to identify the hazards or use of the material. The data in this section have been extracted from ASME Standard A13.1.
Definitions Piping Systems. Piping systems include pipes of any kind, fittings, valves, and pipe coverings. Supports, brackets, and other accessories are not included. Pipes are defined as conduits for the transport of gases, liquids, semiliquids, or fine particulate dust. Materials Inherently Hazardous to Life and Property. There are four categories of hazardous materials:
Color Field
Colors of Letters for Legend
Materials Inherently Hazardous Flammable or explosive Yellow Chemically active or toxic Yellow Extreme temperatures or pressures Yellow Purple Radioactiveb Materials of Inherently Low Hazard Green Liquid or liquid admixturec Gas or gaseous admixture Blue Fire-Quenching Materials Red Water, foam, CO2, Halon, etc.
Black White White
a When preceding color scheme is used, colors should be as recommended
in latest revision of NEMA Standard Z535.1. b Previously specified radioactive markers using yellow or purple are acceptable if already installed and/or until existing supplies are depleted, subject to applicable federal regulations. cMarkers with black letters on green field are acceptable if already installed and/or until existing supplies are depleted.
Fire-Quenching Materials. This classification includes sprinkler systems and other piped firefighting or fire protection equipment. This includes water (for firefighting), chemical foam, CO2, Halon, and so forth.
Method of Identification Legend. The legend is the primary and explicit identification of content. Positive identification of the content of the piping system is by lettered legend giving the name of the contents, in full or abbreviated form, as shown in Table 2. Arrows should be used to indicate the direction of flow. Use the legend to identify contents exactly and to provide temperature, pressure, and other details necessary to identify the hazard. The legend should be brief, informative, pointed, and simple. Legends should be applied close to valves and adjacent to changes in direction, branches, and where pipes pass through walls or floors, and as frequently as needed along straight runs to provide clear and positive identification. Identification may be applied by stenciling, tape, or markers (see Figure 1). The number and location of identification markers on a particular piping system is based on judgment. Color. Colors listed in Table 3 are used to identify the characteristic properties of the contents. Color can be shown on or contiguous to the piping by any physical means, but it should be used
• Flammable or explosive materials that are easily ignited, including materials known as fire producers or explosives • Chemically active or toxic materials that are corrosive or are in themselves toxic or productive of poisonous gases • Materials at extreme temperatures or pressures that, when released from the piping, cause a sudden outburst with the potential for inflicting injury or property damage by burns, impingement, or flashing to vapor state • Radioactive materials that emit ionizing radiation Materials of Inherently Low Hazard. These include all materials that are not hazardous by nature, and are near enough to ambient pressure and temperature that people working on systems carrying these materials run little risk through their release.
Black Black Black Yellow
Fig. 1 Visibility of Pipe Markings
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Abbreviations and Symbols
38.11
Table 4 Size of Legend Letters Outside Diameter of Pipe or Covering, mm
Length of Color Field A, mm
Size of Letters B, mm
20 to 32 40 to 50 65 to 150 200 to 250 over 250
200 200 300 600 800
13 20 32 65 90
CODES AND STANDARDS ASHRAE. 2005. Graphic symbols for heating, ventilating, air-conditioning, and refrigeration systems. ANSI/ASHRAE Standard 134-2005. ASME. 2007. Scheme for the identification of piping systems. ANSI/ASME Standard A13.1-2007. American Society of Mechanical Engineers, New York. ASME. 1998. Graphic symbols for heating, ventilating and air conditioning. Standard Y32.2.4-1949 (RA 1988). American Society of Mechanical Engineers, New York. IEEE. 2004. Standard letter symbols for units of measurement. Standard 260.1-2004. Institute of Electrical and Electronics Engineers, Piscataway, NJ. IEEE. 1996. Letter symbols and abbreviations for quantities used in acoustics. Standard 260.4-1996. Institute of Electrical and Electronics Engineers, Piscataway, NJ. NEMA. 2011. Safety colors. Standard Z535.1-2006 (RA 2011). National Electrical Manufacturers Association, Rosslyn, VA. NFPA. 2012. Standard for fire safety and emergency symbols, 2012 edition. Standard 170. National Fire Protection Association, Quincy, MA.
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in combination with a legend. Color can be used in continuous total length coverage or in intermittent displays. Visibility. Pipe markings should be highly visible. If pipe lines are above the normal line of vision, the lettering is placed below the horizontal centerline of the pipe (Figure 1). Type and Size of Letters. Provide the maximum contrast between color field and legend (Table 3). Table 4 shows the size of letters recommended. Use of standard size letters of 13 mm or larger is recommended. For identifying materials in pipes of less than 20 mm in diameter and for valve and fitting identification, use a permanently legible tag. Unusual or Extreme Situations. When the piping layout occurs in or creates an area of limited accessibility or is extremely complex,
other identification techniques may be required. Although a certain amount of imagination may be needed, the designer should always clearly identify the hazard and use the recommended color and legend guidelines.
Related Commercial Resources
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Related Commercial Resources CHAPTER 39
UNITS AND CONVERSIONS Table 1
Conversions to I-P and SI Units
(Multiply I-P values by conversion factors to obtain SI; divide SI values by conversion factors to obtain I-P)
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Multiply I-P acre (43 560 ft2).................................................. .................................................. atmosphere (standard) ........................................ bar....................................................................... barrel (42 U.S. gal, petroleum)........................... .......................... Btu (International Table).................................... Btu (thermochemical) ........................................ Btu/ft2 (International Table)............................... Btu/ft3 (International Table)............................... Btu/gal ................................................................ Btu·ft/h·ft2 · °F.................................................... Btu·in/h·ft2 · °F (thermal conductivity k) .......... . Btu/h ................................................................... Btu/h·ft2 ............................................................. Btu/h·ft2 · °F (overall heat transfer coefficient U) Btu/lb.................................................................. Btu/lb·°F (specific heat cp) ................................ bushel (dry, U.S.) ............................................... calorie (thermochemical).................................... centipoise (dynamic viscosity )........................ centistokes (kinematic viscosity ) .................... clo ....................................................................... dyne .................................................................... dyne/cm2 ............................................................. EDR hot water (150 Btu/h) ................................ EDR steam (240 Btu/h) ...................................... EER .................................................................... ft ......................................................................... .......................................................................... ft/min, fpm.......................................................... ft/s, fps................................................................ ft of water ........................................................... ft of water per 100 ft pipe................................... ft2 ........................................................................ ft2 ·h· °F/Btu (thermal resistance R).................... ft2/s (kinematic viscosity )................................ ft3 ........................................................................ ........................................................................ ft3/min, cfm ........................................................ ft3/s, cfs............................................................... ft·lbf (torque or moment) ................................... ft·lbf (work)........................................................ ft·lbf /lb (specific energy)................................... ft·lbf /min (power) .............................................. footcandle ........................................................... gallon (U.S., *231 in3)........................................ gph...................................................................... gpm..................................................................... gpm/ft2 ................................................................ gpm/ton refrigeration.......................................... grain (1/7000 lb)................................................. gr/gal................................................................... gr/lb .................................................................... horsepower (boiler) (33 470 Btu/h).................... horsepower (550 ft·lbf /s) ................................... inch ..................................................................... in. of mercury (60°F).......................................... in. of water (60°F) .............................................. in/100 ft, thermal expansion coefficient............. To Obtain I-P
By 0.4047 4046.873 *101.325 *100 159.0 0.1580987 1055.056 1054.350 11,356.53 37,258.951 278,717.1765 1.730735 0.1442279 0.2930711 3.154591 5.678263 *2.326 *4.1868 0.0352394 *4.184 *1.00 *1.00 0.155 1.0 10–5 *0.100 43.9606 70.33706 0.293 *0.3048 *304.8 *0.00508 *0.3048 2989 98.1 0.092903 0.176110 92,900 28.316846 0.02832 0.471947 28.316845 1.355818 1.356 2.99 0.0226 10.76391 3.785412 1.05 0.0631 0.6791 0.0179 0.0648 17.1 0.143 9.81 0.7457 *25.4 3.3864 248.84 0.833 By
To Obtain SI
Multiply I-P
ha m2 kPa kPa L m3 J J J/m2 J/m3 J/m3 W/(m·K) W/(m·K) W W/m2 W/(m2 ·K) kJ/kg kJ/(kg·K) m3 J mPa·s mm2/s (m2 ·K)/W N Pa W W COP m mm m/s m/s Pa Pa/m m2 (m2 ·K)/W mm2/s L m3 L/s L/s N·m J J/kg W lx L mL/s L/s L/(s·m2) mL/J g g/m3 g/kg kW kW mm kPa Pa mm/m
in·lbf (torque or moment) ................................. 113 in2...................................................................... 645.16 in3 (volume)...................................................... 16.3874 in3/min (SCIM)................................................. 0.273117 in3 (section modulus)........................................ 16387 in4 (section moment) ........................................ 416 231 kWh .................................................................. *3.60 kW/1000 cfm .................................................... 2.118880 kilopond (kg force) ........................................... 9.81 kip (1000 lbf) .................................................... 4.45 kip/in2 (ksi) ....................................................... 6.895 litre.................................................................... *0.001 met .................................................................... 58.15 micron (m) of mercury (60°F)........................ 133 mile ................................................................... 1.609 mile, nautical .................................................... *1.852 mile per hour (mph).......................................... 1.609344 ......................................... 0.447 millibar ............................................................. *0.100 mm of mercury (60°F)...................................... 0.133 mm of water (60°F) .......................................... 9.80 ounce (mass, avoirdupois) ................................ 28.35 ounce (force or thrust) ...................................... 0.278 ounce (liquid, U.S.) .......................................... 29.6 ounce inch (torque, moment)............................ 7.06 ounce (avoirdupois) per gallon ......................... 7.489152 perm (permeance at 32°F) ................................ 5.72135 10–11 perm inch (permeability at 32°F) ..................... 1.45362 10–12 pint (liquid, U.S.).............................................. 4.73176 10–4 pound lb (avoirdupois, mass) ...................................... 0.453592 ...................................... 453.592 lbf (force or thrust)............................................ 4.448222 lbf /ft (uniform load).......................................... 14.59390 lb/ft·h (dynamic viscosity )............................ 0.4134 lb/ft·s (dynamic viscosity ) ............................ 1490 lbf ·s/ft2 (dynamic viscosity ) ......................... 47.88026 lb/h .................................................................... 0.000126 lb/min................................................................ 0.007559 lb/h [steam at 212°F (100°C)] .......................... 0.2843 lbf /ft2 ................................................................. 47.9 lb/ft2 .................................................................. 4.88 lb/ft3 (density )................................................ 16.0 lb/gallon ............................................................ 120 ppm (by mass) .................................................. *1.00 psi ..................................................................... 6.895 quad (1015 Btu) ................................................. 1.055 quart (liquid, U.S.)............................................ 0.9463 square (100 ft2) ................................................. 9.2903 tablespoon (approximately) .............................. 15 teaspoon (approximately) ................................. 5 therm (U.S.) ...................................................... 105.5 ton, long (2240 lb) ............................................ 1.016046 ton, short (2000 lb) ........................................... 0.907184 ton, refrigeration (12 000 Btu/h) ...................... 3.517 torr (1 mm Hg at 0°C) ...................................... 133 watt per square foot .......................................... 10.76 yd ...................................................................... *0.9144 yd2 ..................................................................... 0.8361 yd3 ..................................................................... 0.7646
kg g N N/m mPa·s mPa·s Pa·s kg/s kg/s kW Pa kg/m2 kg/m3 kg/m3 mg/kg kPa EJ L m2 mL mL MJ Mg Mg; t (tonne) kW Pa W/m2 m m2 m3
Divide SI
To Obtain I-P
Divide SI
The preparation of this chapter is assigned to TC 1.6, Terminology.
By
To Obtain SI mN·m mm2 mL mL/s mm3 mm4 MJ kJ/m3 N kN MPa m3 W/m2 mPa km km km/h m/s kPa kPa Pa g N mL mN·m kg/m3 kg/(Pa·s·m2) kg/(Pa·s·m) m3
*Conversion factor is exact. Notes: 1. Units are U.S. values unless noted otherwise. 2. Litre is a special name for the cubic decimetre. 1 L = 1 dm3 and 1 mL = 1 cm3.
39.1 Copyright © 2017, ASHRAE
By
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39.2
2017 ASHRAE Handbook—Fundamentals (SI) Table 2 Conversion Factors
Pressure psi
in. of water (60°F)
in. Hg (32°F)
1 = 27.708 0.036091 1 0.491154 13.609 14.6960 407.19 0.0193368 0.53578 14.5038 401.86 14.223 394.1 1.45038 × 10–4 4.0186 × 10–3 Mass
= 2.0360 0.073483 1 29.921 0.03937 29.530 28.959 2.953 × 10–4
lb (avoir.) 1 1.4286 × 10–4 0.06250 2.20462
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Volume
cubic inch 1 1728* 231.0* 61.02374 6.102374 × 104
Energy
Btu
Note: MBtu, which is 1000 Btu, is confusing and should not be used.
1 1.2851 × 10–3 3.9683 × 10–3 9.4782 × 10–4 3.41214
Density
lb/ft3
ft3/lb
Specific Volume
grain
1 0.133680 0.016018 16.018463
= 0.068948 2.4884 × 10–3 0.033864 1.01325* 1.3332 × 10–3 1 0.980665* 10–5*
ounce (avoir.)
= 7000* 1 437.5* 1.5432 × 104
= 16* 2.2857 × 10–3 1 35.274
cubic foot 10–4
= 5.787 × 1 0.13368 0.035315 35.315
kgf/cm2
bar
= 0.068046 = 51.715 1.8665 2.4559 × 10-3 0.033421 25.400 1 760.0 1 1.31579 × 10–3 0.98692 750.062 0.96784 735.559 9.8692 × 10–6 7.50 × 10–3
kg
cubic metre (m3)
litre 10–3
= 4.329 × 7.48052 1 0.264173 264.173
joule (J) = watt-second (W·s)
= 778.17 1 3.08803 0.73756 2655.22
= 251.9958 0.32383 1 0.23885 859.85
= 1055.056 1.355818 4.1868* 1 3600*
lb/gal
g/cm3
kg/m3
= 0.133680 1 8.34538 0.008345
= 0.016018 0.119827 1 0.001*
= 16.018463 119.827 1000* 1
gal/lb
cm3/g = 62.4280 8.34538 1 1000*
= 1.63871 × 10–5 0.028317 0.0037854 0.001* 1
= 0.0163871 28.317 3.7854 1 1000*
calorie (cal)
= 7.48055 1 0.119827 119.827
pascal
= 0.07030696 = 6894.8 2.537 × 10–3 248.84 0.034532 3386.4 1.03323 1.01325 × 105* 1.3595 × 10–3 133.32 1.01972* 105* 1 9.80665 × 104* 1.01972 × 10–5* 1
= 0.45359 6.4800 × 10–5 0.028350 1
gallon
ft·lbf
1 7.48055 62.4280 0.0624280
mm Hg (32°F)
atmosphere
watt-hour (W·h) = 0.293071 3.76616 × 10–4 1.163 × 10–3* 2.7778 × 10–4 1
m3/kg = 0.0624280 0.008345 0.001* 1
1 poise = 1 dyne-sec/cm2 = 0.1 Pa·s = 1 g/(cm·s)
Viscosity (absolute)
lbf ·s/ft2
poise 1 478.8026 1.72369 × 106 10* 14.8819
lbf ·h/ft2
= 2.0885 × 10–3 1 3600* 0.020885 0.031081
Temperature
kg/(m·s) = N·s/m2
= 5.8014 × 10–7 2.7778 × 10–4 1 5.8014 × 10–6 8.6336 × 10–6
lbm/ft·s
= 0.1* 47.88026 1.72369 × 105 1 1.4882
= 0.0671955 32.17405 1.15827 × 105 0.0671955 1
Temperature
Scale
Temperature Interval
K
°C
°R
°F
Kelvin
xK=
x
x – 273.15
1.8x
1.8x – 459.67
1K=
Celsius
x°C =
x + 273.15
x
1.8x + 491.67
1.8x + 32
1°C =
Rankine
x°R =
x/1.8
(x – 491.67)/1.8
x
x – 459.67
1°R =
Fahrenheit
x°F =
(x + 459.67)/1.8
(x – 32)/1.8
x + 459.67
x
1°F =
5/9
Notes: Conversions with * are exact. The Btu and calorie are based on the International Table.
K
°C
°R
°F
1
1
9/5 = 1.8
9/5 = 1.8
1
1
9/5 = 1.8
9/5 = 1.8
5/9
5/9
1
1
5/9
1
1
All temperature conversions and factors are exact. The term centigrade is obsolete and should not be used.
When making conversions, remember that a converted value is no more precise than the original value. For many applications, rounding off the converted value to the same number of significant
figures as those in the original value provides acceptable accuracy. See ANSI Standard SI-10-1997 (available from ASTM or IEEE) for additional conversions.
Related Commercial Resources
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Related Commercial Resources CHAPTER 40
CODES AND STANDARDS HE codes and standards and other publications listed here represent practices, methods, or standard published by the organizations indicated. They are useful guides for the practicing engineer in determining test methods, ratings, performance requirements, and limits of HVAC&R equipment. Copies of these publications can be obtained from most of the organizations listed in the Publisher column, from Global Engineering Documents at global.ihs.com, or from Techstreet at techstreet.com. Addresses of the organizations are given at the end of the chapter. A comprehensive database with over 300,000 industry, government, and international standards is at www.nssn.org.
T
Licensed for single user. © 2017 ASHRAE, Inc.
Selected Codes and Standards Published by Various Societies and Associations Subject
Title
Publisher
Reference
Air Conditioners
Commercial Systems Overview Residential Equipment Selection, 2nd ed. HVAC Quality Installation Specification Technician’s Guide & Workbook for Quality Installations Laboratory Methods of Testing Air Terminal Units Non-Ducted Air Conditioners and Heat Pumps—Testing and Rating for Performance Ducted Air-Conditioners and Air-to-Air Heat Pumps—Testing and Rating for Performance Guidelines for Roof Mounted Outdoor Air-Conditioner Installations Heating and Cooling Equipment
ACCA ACCA ACCA ACCA ASHRAE ISO ISO SMACNA UL/CSA
Performance Standard for Split-System and Single-Package Central Air Conditioners and Heat Pumps Performance Standard for Rating Large and Single Packaged Air Conditioners and Heat Pumps Commercial Systems Overview Residential Equipment Selection, 2nd ed. Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Home Evaluation and Performance Improvement Gas-Fired, Heat Activated Air Conditioning and Heat Pump Appliances
CSA
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S®-2014 ANSI/ACCA 5 QI-2015 ACCA 2015 ANSI/ASHRAE 130-2016 ISO 5151:2010 ISO 13253:2011 SMACNA 1997 ANSI/UL 1995-2011/C22.2 No. 236-11 CAN/CSA-C656-14
CSA
CAN/CSA-C746-06 (R2012)
ACCA ACCA ACCA ACCA CSA
Gas-Fired
Packaged Terminal Room
Unitary
Gas-Fired Work Activated Air Conditioning and Heat Pump Appliances (Internal Combustion) Performance Testing and Rating of Gas-Fired Air Conditioning and Heat Pump Appliances Packaged Terminal Air-Conditioners and Heat Pumps
CSA
AHRI/CSA
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ANSI/ACCA 12 QH-2014 ACCA 2014 ANSI Z21.40.1-1996/CGA 2.91-M96 (R2012) ANSI Z21.40.2-1996/CGA 2.92-M96 (R2002) ANSI Z21.40.4-1996/CGA 2.94-M96 (R2002) AHRI 310/380-2014/CSA C744-14
Room Air Conditioners Method of Testing for Rating Room Air Conditioners and Packaged Terminal Air Conditioners Method of Testing for Rating Room Air Conditioner and Packaged Terminal Air Conditioner Heating Capacity Method of Testing for Rating Fan-Coil Conditioners Energy Performance of Room Air Conditioners Room Air Conditioners Room Air Conditioners, Ed. 7 Commercial Systems Overview Residential Equipment Selection, 2nd ed. Unitary Air-Conditioning and Air-Source Heat Pump Equipment Sound Rating of Outdoor Unitary Equipment Application of Sound Rating Levels of Outdoor Unitary Equipment Commercial and Industrial Unitary Air-Conditioning and Heat Pump Equipment Methods of Testing for Rating Electrically Driven Unitary Air-Conditioning and Heat Pump Equipment Methods of Testing for Rating Heat Operated Unitary Air-Conditioning and Heat-Pump Equipment Methods of Testing for Rating Seasonal Efficiency of Unitary Air Conditioners and Heat Pumps Method of Testing for Rating Computer and Data Processing Room Unitary Air Conditioners Method of Rating Unitary Spot Air Conditioners
AHAM ASHRAE
ANSI/AHAM RAC-1-2014 ANSI/ASHRAE 16-2016
ASHRAE
ANSI/ASHRAE 58-1986 (RA14)
ASHRAE CSA CSA UL ACCA ACCA AHRI AHRI AHRI AHRI ASHRAE
ANSI/ASHRAE 79-2015 CAN/CSA-C368.1-14 C22.2 No. 117-1970 (R2012) ANSI/UL 484-2014 ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ANSI/AHRI 210/240-2008 AHRI 270-2008 AHRI 275-2010 AHRI 340/360-2007 ANSI/ASHRAE 37-2009
ASHRAE
ANSI/ASHRAE 40-2014
ASHRAE
ANSI/ASHRAE 116-2010
ASHRAE
ANSI/ASHRAE 127-2012
ASHRAE
ANSI/ASHRAE 128-2011
40.1 Copyright © 2017, ASHRAE
CSA
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.2
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Title
Publisher
Reference
Specification for Mechanically Refrigerated Shipboard Air Conditioner Flashing and Stand Combination for Air Conditioning Units (Unit Curb)
ASTM IAPMO
ASTM F1433-97 (2010) IAPMO PS 120-2004
Commercial Systems Overview Heat Pump Systems Residential Load Calculation, 8th ed. Commercial Load Calculation, 5th ed. Comfort, Air Quality, and Efficiency by Design HVAC Quality Installation Specification Technician’s Guide & Workbook for Quality Installations Peak Cooling and Heating Load Calculations in Buildings Except Low-Rise Residential Buildings Environmental Systems Technology, 2nd ed. (1999) Installation of Air-Conditioning and Ventilating Systems Standard of Purity for Use in Mobile Air-Conditioning Systems HVAC Systems Applications, 2nd ed. HVAC Systems—Duct Design, 4th ed. Heating and Cooling Equipment
ACCA ACCA ACCA ACCA ACCA ACCA ACCA ASHRAE/ ACCA NEBB NFPA SAE SMACNA SMACNA UL/CSA
Air Conditioning of Aircraft Cargo Aircraft Fuel Weight Penalty Due to Air Conditioning Air Conditioning Systems for Subsonic Airplanes Environmental Control Systems Terminology Testing of Airplane Installed Environmental Control Systems (ECS) Guide for Qualification Testing of Aircraft Air Valves Control of Excess Humidity in Avionics Cooling Engine Bleed Air Systems for Aircraft Aircraft Ground Air Conditioning Service Connection Air Cycle Air Conditioning Systems for Military Air Vehicles Refrigerant 12 Automotive Air-Conditioning Hose Design Guidelines for Air Conditioning Systems for Off-Road Operator Enclosures Test Method for Measuring Power Consumption of Air Conditioning and Brake Compressors for Trucks and Buses Information Relating to Duty Cycles and Average Power Requirements of Truck and Bus Engine Accessories Rating Air-Conditioner Evaporator Air Delivery and Cooling Capacities Recovery and Recycle Equipment for Mobile Automotive Air-Conditioning Systems R134a Refrigerant Automotive Air-Conditioning Hose Service Hose for Automotive Air Conditioning Mechanical Refrigeration and Air-Conditioning Installations Aboard Ship Practice for Mechanical Symbols, Shipboard Heating, Ventilation, and Air Conditioning (HVAC)
SAE SAE SAE SAE SAE SAE SAE SAE SAE SAE SAE SAE SAE
ACCA Manual CS-1993 ACCA Manual H-1984 ANSI/ACCA 2 Manual J-2016 ACCA Manual N-2012 ACCA Manual RS-1997 ANSI/ACCA 5 QI-2015 ACCA 2015 ANSI/ASHRAE/ACCA 183-2007 (RA14) NEBB NFPA 90A-2015 SAE J1991-2011 SMACNA 2010 SMACNA 2006 ANSI/UL 1995-2011/C22.2 No. 236-11 SAE AIR806B-1997 (R2011) SAE AIR1168/8-2011 SAE ARP85F-2012 SAE ARP147E-2001 (R2012) SAE ARP217D-1999 (R2011) SAE ARP986D-2008 SAE ARP987B-2010 SAE ARP1796B-2012 SAE AS4262B-2012 SAE AS4073A-2013 SAE J51-2004 SAE J169-2013 SAE J1340-2011
SAE
SAE J1343-2000
SAE SAE SAE SAE ASHRAE ASTM
SAE J1487-2013 SAE J1990-2011 SAE J2064-2011 SAE J2196-2011 ANSI/ASHRAE 26-2010 ASTM F856-97 (2014)
Air Curtains
Laboratory Methods of Testing Air Curtains for Aerodynamic Performance Air Terminals Standard Methods for Laboratory Airflow Measurement Method of Testing the Performance of Air Outlets and Inlets Residential Mechanical Ventilating Systems Air Curtains for Entranceways in Food and Food Service Establishments
AMCA AHRI ASHRAE ASHRAE CSA NSF
AMCA 220-05 (R2012) AHRI 881-2011 ANSI/ASHRAE 41.2-1987 (RA92) ANSI/ASHRAE 70-2006 (RA11) CAN/CSA-F326-M91 (R2010) ANSI/NSF 37-2012
Air Diffusion
Air Distribution Basics for Residential and Small Commercial Buildings Balancing and Testing Air and Hydronic Systems Method of Testing the Performance of Air Outlets and Inlets Method of Testing for Room Air Diffusion
ACCA ACCA ASHRAE ASHRAE
ACCA Manual T-2001 ACCA Manual B-2009 ANSI/ASHRAE 70-2006 (RA11) ANSI/ASHRAE 113-2013
Air Filters
Comfort, Air Quality, and Efficiency by Design Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Home Evaluation and Performance Improvement Balancing and Testing Air and Hydronic Systems Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Air Cleaners—Portable Residential Air Filter Equipment Commercial and Industrial Air Filter Equipment Method of Testing General Ventilation Air-Cleaning Devices for Removal Efficiency by Particle Size
ACCA ACCA ACCA ACCA ACGIH AHAM AHRI AHRI ASHRAE
ACCA Manual RS-1997 ANSI/ACCA 12 QH-2014 ACCA 2015 ACCA Manual B-2009 ACGIH ANSI/AHAM AC-1-2013 ANSI/AHRI 681-2009 ANSI/AHRI 850-2013 ANSI/ASHRAE 52.2-2012
Ships Accessories Air Conditioning
Licensed for single user. © 2017 ASHRAE, Inc.
Aircraft
Automotive
Ships
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Codes and Standards
40.3
Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
Laboratory Test Method for Assessing the Performance of Gas-Phase Air Cleaning Systems: Loose Granular Media Laboratory Test Method for Assessing the Performance of Gas-Phase Air Cleaning Systems: Air Cleaning Devices Code on Nuclear Air and Gas Treatment Nuclear Power Plant Air-Cleaning Units and Components Testing of Nuclear Air-Treatment Systems Test Method for Air Cleaning Performance of a High-Efficiency Particulate Air Filter System Specification for Filters Used in Air or Nitrogen Systems Method for Sodium Flame Test for Air Filters Particulate Air Filters for General Ventilation: Determination of Filtration Performance Electrostatic Air Cleaners High-Efficiency, Particulate, Air Filter Units Air Filter Units Exhaust Hoods for Commercial Cooking Equipment Grease Filters for Exhaust Ducts
ASHRAE
ANSI/ASHRAE 145.1-2015
ASHRAE
ANSI/ASHRAE 145.2-2016
ASME ASME ASME ASTM
ASME AG-1-2012 ASME N509-2002 ASME N510-2007 ASTM F1471-09
ASTM BSI BSI
ASTM F1791-00 (2013) BS 3928:1969 BS EN 779:2002
UL UL UL UL UL ACCA ACCA AHRI CSA
ANSI/UL 867-2011 ANSI/UL 586-2009 ANSI/UL 900-2004 UL 710-2012 UL 1046-2010 ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ANSI/AHRI 430-2014 ANSI Z83.4-2013/CSA 3.7-2013
Air-Handling Units
Commercial Systems Overview Residential Equipment Selection Central Station Air-Handling Units Non-Recirculating Direct Gas-Fired Industrial Air Heaters
Air Leakage
Technician’s Guide & Workbook for Duct Diagnostics and Repair HVAC Quality Installation Specification Technician’s Guide & Workbook for Quality Installations Ventilation and Acceptable Indoor Air Quality in Low-Rise Residential Buildings Method of Test for Determining the Airtightness of HVAC Equipment Test Method for Determining Air Change in a Single Zone by Means of a Tracer Gas Dilution Test Method for Determining Air Leakage Rate by Fan Pressurization Test Method for Field Measurement of Air Leakage Through Installed Exterior Window and Doors Practices for Air Leakage Site Detection in Building Envelopes and Air Barrier Systems Test Method for Determining the Rate of Air Leakage Through Exterior Windows, Curtain Walls, and Doors Under Specified Pressure and Temperature Differences Across the Specimen Test Methods for Determining Airtightness of Buildings Using an Orifice Blower Door Practice for Determining the Effects of Temperature Cycling on Fenestration Products Test Method for Determining Air Flow Through the Face and Sides of Exterior Windows, Curtain Walls, and Doors Under Specified Pressure Differences Across the Specimen Test Method for Determining Air Leakage of Air Barrier Assemblies HVAC Air Duct Leakage Test Manual, 2nd ed.
ACCA ACCA ACCA ASHRAE ASHRAE ASTM
ACCA 2016 ANSI/ACCA 5 QI-2015 ACCA 2015 ANSI/ASHRAE 62.2-2016 ANSI/ASHRAE 193-2010 (RA14) ASTM E741-11
ASTM ASTM
ASTM E779-10 ASTM E783-02 (2010)
ASTM ASTM
ASTM E1186-03 (2009) ASTM E1424-91 (2008)
ASTM ASTM ASTM
ASTM E1827-11 ASTM E2264-05 (2013) ASTM E2319-04 (2011)
ASTM SMACNA
ASTM E2357-11 SMACNA 2012
Packaged Boiler Engineering Manual (1998) Selected Codes and Standards of the Boiler Industry (2001) Operation and Maintenance Safety Manual (1995) Fluidized Bed Combustion Guidelines (1995) Guide to Clean and Efficient Operation of Coal Stoker-Fired Boilers (2002) Guideline for Performance Evaluation of Heat Recovery Steam Generating Equipment (1995) Guidelines for Industrial Boiler Performance Improvement (1999) Measurement of Sound from Steam Generators (1995) Guideline for Gas and Oil Emission Factors for Industrial, Commercial, and Institutional Boilers (1997) Combustion Control Guidelines for Single Burner Firetube and Watertube Industrial/ Commercial/Institutional Boilers (1999) Combustion Control Guidelines for Multiple-Burner Boilers (2001) Boiler Water Quality Requirements and Associated Steam Quality for Industrial/ Commercial and Institutional Boilers (2005) HVAC Quality Installation Specification Technician’s Guide & Workbook for Quality Installations Residential Equipment Selection Commercial Systems Overview
ABMA ABMA ABMA ABMA ABMA ABMA
ABMA 100 ABMA 103 ABMA 106 ABMA 200 ABMA 203 ABMA 300
ABMA ABMA ABMA
ABMA 302 ABMA 304 ABMA 305
ABMA
ABMA 307
ABMA ABMA
ABMA 308 ABMA 402
ACCA ACCA ACCA ACCA
ANSI/ACCA 5 QI-2015 ACCA 2015 ANSI/ACCA 3 Manual S 2014 ACCA Manual CS-1993
Boilers
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40.4
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
ASHRAE
ANSI/ASHRAE 103-2007
ASME
BPVC-2015
ASME CSA HYDI
ASME PTC 4-2013 CSA B51-14 HYDI BTS-2000
HYDI NFPA NFPA UL NFPA ACCA ACCA ASME CSA CSA CSA UL UL HEI ABMA
IBR ANSI/NFPA 8501-97 ANSI /NFPA 8502-99 ANSI/UL 834-2004 NFPA 85-11 ANSI/ACCA 12QH-2014 ACCA 2015 ASME CSD-1-2012 ANSI Z21.13-2014/CSA 4.9-2014 CAN 1-3.1-77 (R2011) B140.7-05 (R2010) UL 726-1995 UL 795-2011 HEI 2954 ABMA 101
ASTM ASTM ASTM ASTM AWS BOCA ICBO ICC ICC ICC ICC ICC ICC NFPA NRCC SBCCI ASME CSA CSA IAPMO ICC ICC SBCCI
ASTM ASTM E2255M-13 ASTM E2270-14 ASTM E2813-12 AWS D1.1M/D1.1:2010 BNBC UBC V1, V2, V3 IBC ICC PC IEBC IECC IPMC IRC ANSI/NFPA 5000-2015 NRCC SBC ASME A17.1-2013 CAN/CSA-B149.1-10 CAN/CSA-B149.2-10 IAPMO IMC IFGC SBC
Domestic Gas Conversion Burners
CSA
Installation of Domestic Gas Conversion Burners Installation Code for Oil Burning Equipment Oil-Burning Equipment: General Requirements Vapourizing-Type Oil Burners Atomizing-Type Oil Burners Pressure Atomizing Oil Burner Nozzles Oil Burners Waste Oil-Burning Air-Heating Appliances Commercial-Industrial Gas Heating Equipment Commercial/Industrial Gas and/or Oil-Burning Assemblies with Emission Reduction Equipment
CSA CSA CSA CSA CSA CSA UL UL UL UL
ANSI Z21.17-1998/CSA 2.7-M98 (R2014) ANSI Z21.8-1994 (R2012) CAN/CSA-B139-09 (R2014) CAN/CSA-B140.0-03 (R2013) B140.1-1966 (R2011) CAN/CSA-B140.2.1-10 B140.2.2-1971 (R2011) ANSI/UL 296-2003 ANSI/UL 296A-1995 UL 795-2011 UL 2096-2006
Commercial Systems Overview Absorption Water Chilling and Water Heating Packages Water Chilling Packages Using the Vapor Compression Cycle
ACCA AHRI AHRI
ACCA Manual CS-1993 ANSI/AHRI 560-2000 ANSI/AHRI 551/591-2011
Method of Testing for Annual Fuel Utilization Efficiency of Residential Central Furnaces and Boilers Boiler and Pressure Vessel Code—Section I: Power Boilers; Section IV: Heating Boilers Fired Steam Generators Boiler, Pressure Vessel, and Pressure Piping Code Testing Standard, Method to Determine Efficiency of Commercial Heating Boilers, 2nd ed. (2007) Rating Procedure for Heating Boilers, 6th ed. (2005) Single Burner Boiler Operations Prevention of Furnace Explosions/Implosions in Multiple Burner Boilers Heating, Water Supply, and Power Boilers—Electric Boiler and Combustion Systems Hazards Code Gas or Oil Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Home Evaluation and Performance Improvement Controls and Safety Devices for Automatically Fired Boilers Gas-Fired Low-Pressure Steam and Hot Water Boilers Industrial and Commercial Gas-Fired Package Boilers Oil-Burning Equipment: Steam and Hot-Water Boilers Oil-Fired Boiler Assemblies Commercial-Industrial Gas Heating Equipment Standards and Typical Specifications for Tray Type Deaerators, 9th ed. (2011) Terminology Ultimate Boiler Industry Lexicon: Handbook of Power Utility and Boiler Terms and Phrases, 6th ed. (2001) Building Codes ASTM Standards Used in Building Codes Practice for Conducting Visual Assessments for Lead Hazards in Buildings Standard Practice for Periodic Inspection of Building Facades for Unsafe Conditions Standard Practice for Building Enclosure Commissioning Structural Welding Code—Steel BOCA National Building Code, 14th ed. (1999) Uniform Building Code, vol. 1, 2, and 3 (1997) International Building Code® (2015) International Code Council Performance Code® (2015) International Existing Building Code® (2015) International Energy Conservation Code® (2015) International Property Maintenance Code® (2015) International Residential Code® (2015) Building Construction and Safety Code® National Building Code of Canada (2010) Standard Building Code (1999) Mechanical Safety Code for Elevators and Escalators Natural Gas and Propane Installation Code Propane Storage and Handling Code Uniform Mechanical Code (2012) International Mechanical Code® (2015) International Fuel Gas Code® (2015) Standard Gas Code (1999) Burners
Chillers
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Codes and Standards
40.5
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Chimneys
Licensed for single user. © 2017 ASHRAE, Inc.
Cleanrooms
Title
Publisher
Reference
Method of Testing Liquid-Chilling Packages Method of Testing Absorption Water-Chilling and Water-Heating Packages Performance Standard for Rating Packaged Water Chillers
ASHRAE ASHRAE CSA
ANSI/ASHRAE 30-1995 ANSI/ASHRAE 182-2008 (RA13) CAN/CSA C743-09
Specification for Clay Flue Liners Specification for Industrial Chimney Lining Brick Practice for Installing Clay Flue Lining Guide for Design and Construction of Brick Liners for Industrial Chimneys Guide for Design, Fabrication, and Erection of Fiberglass Reinforced Plastic (FRP) Chimney Liners with Coal-Fired Units Chimneys, Fireplaces, Vents, and Solid Fuel-Burning Appliances Medium Heat Appliance Factory-Built Chimneys Factory-Built Chimneys for Residential Type and Building Heating Appliance
ASTM ASTM ASTM ASTM ASTM
ASTM C315-07 (2011) ASTM C980-13a ASTM C1283-11 ASTM C1298-95 (2013) ASTM D5364-14
NFPA UL UL
ANSI/NFPA 211-2013 ANSI/UL 959-2010 ANSI/UL 103-2010
ASTM ASTM
ASTM E2042/E2042M-09 ASTM E2217-12
ASTM ASTM
ASTM E2312-11 ASTM E2352-04 (2010)
ASTM
ASTM F25/F25M-09
ASTM
ASTM F50-12
NEBB
NEBB
Practice for Cleaning and Maintaining Controlled Areas and Clean Rooms Practice for Design and Construction of Aerospace Cleanrooms and Contamination Controlled Areas Practice for Tests of Cleanroom Materials Practice for Aerospace Cleanrooms and Associated Controlled Environments— Cleanroom Operations Test Method for Sizing and Counting Airborne Particulate Contamination in Clean Rooms and Other Dust-Controlled Areas Designed for Electronic and Similar Applications Practice for Continuous Sizing and Counting of Airborne Particles in Dust-Controlled Areas and Clean Rooms Using Instruments Capable of Detecting Single SubMicrometre and Larger Particles Procedural Standards for Certified Testing of Cleanrooms, 3rd ed. (2009)
Climate Data
Residential Load Calculations Commercial Load Calculations Climatic Data for Building Design Standards
ACCA ACCA ASHRAE
ANSI/ACCA 2 Manual J-2016 ACCA Manual N-2012 ANSI/ASHRAE 169-2013
Coils
Forced-Circulation Air-Cooling and Air-Heating Coils Methods of Testing Forced Circulation Air Cooling and Air Heating Coils
AHRI ASHRAE
AHRI 410-2001 ANSI/ASHRAE 33-2016
Comfort Conditions
Threshold Limit Values for Physical Agents (updated annually)
ACGIH
ACGIH
Comfort, Air Quality, and Efficiency by Design Thermal Environmental Conditions for Human Occupancy Classification for Serviceability of an Office Facility for Thermal Environment and Indoor Air Conditions Hot Environments—Estimation of the Heat Stress on Working Man, Based on the WBGT Index (Wet Bulb Globe Temperature) Ergonomics of the Thermal Environment—Analytical Determination and Interpretation of Thermal Comfort Using Calculation of the PMV and PPD Indices and Local Thermal Comfort Criteria Ergonomics of the Thermal Environment—Determination of Metabolic Rate Ergonomics of the Thermal Environment—Estimation of the Thermal Insulation and Water Vapour Resistance of a Clothing Ensemble
ACCA ASHRAE ASTM
ACCA Manual RS-1997 ANSI/ASHRAE 55-2013 ASTM E2320-04 (2012)
ISO
ISO 7243:1989
ISO
ISO 7730:2005
ISO ISO
ISO 8996:2004 ISO 9920:2007
ACCA ACCA ACCA ASHRAE ASHRAE ASHRAE ASTM SMACNA
ANSI/ACCA 5 QI-2015 ANSI/ACCA 9 QIvp-2016 ACCA 2015 ASHRAE Guideline 0-2013 ASHRAE Guideline 1.1-2007 ASHRAE 202-2013 ASTM E2813-12 SMACNA 2013
ASME ASME CAGI AHRI AHRI ASHRAE ASHRAE
ASME PTC 9-1970 ASME PTC 10-1997 (RA14) CAGI ANSI/AHRI 520-2004 ANSI/AHRI 540-2004 ANSI/ASHRAE 15-2016 ANSI/ASHRAE 23.1-2010
ISO ISO
ISO 917:1989 ISO 9309:1989
Commissioning HVAC Quality Installation Specification HVAC Quality Installation Verification Protocols Technician’s Guide & Workbook for Quality Installations The Commissioning Process HVAC&R Technical Requirements for the Commissioning Process Commissioning Process for Buildings and Systems Standard Practice for Building Enclosure Commissioning HVAC Systems—Commissioning Manual, 2nd ed. Compressors
Refrigerant
Displacement Compressors, Vacuum Pumps and Blowers Performance Test Code on Compressors and Exhausters Compressed Air and Gas Handbook, 6th ed. (2003) Positive Displacement Condensing Units Positive Displacement Refrigerant Compressors and Compressor Units Safety Standard for Refrigeration Systems Methods of Testing for Rating Positive Displacement Refrigerant Compressors and Condensing Units That Operate at Subcritical Temperatures of the Refrigerant Testing of Refrigerant Compressors Refrigerant Compressors—Presentation of Performance Data
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.6
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Title
Publisher
Reference
Hermetic Refrigerant Motor-Compressors
UL/CSA
UL 984-1996/C22.2 No.140.2-96 (R2011)
Method of Testing for Rating Computer and Data Processing Room Unitary Air Conditioners Method of Test for the Evaluation of Building Energy Analysis Computer Programs Fire Protection of Information Technology Equipment
ASHRAE
ANSI/ASHRAE 127-2012
ASHRAE
ANSI/ASHRAE 140-2011
NFPA
NFPA 75-2013
Commercial Systems Overview Residential Equipment Selection, 2nd ed. Water-Cooled Refrigerant Condensers, Remote Type Remote Mechanical-Draft Air-Cooled Refrigerant Condensers Remote Mechanical Draft Evaporative Refrigerant Condensers Safety Standard for Refrigeration Systems Methods of Testing for Rating Remote Mechanical-Draft Air-Cooled Refrigerant Condensers Methods of Testing for Rating Liquid Water-Cooled Refrigerant Condensers Methods of Testing for Rating Positive Displacement Refrigerant Compressors and Condensing Units That Operate at Subcritical Temperatures of the Refrigerant Methods of Laboratory Testing Remote Mechanical-Draft Evaporative Refrigerant Condensers Steam Surface Condensers Standards for Steam Surface Condensers, 11th ed. (2012) Standards for Direct Contact Barometric and Low Level Condensers, 8th ed. (2010) Refrigerant-Containing Components and Accessories, Nonelectrical
ACCA ACCA AHRI AHRI AHRI ASHRAE ASHRAE
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 AHRI 450-2007 ANSI/AHRI 460-2005 ANSI/AHRI 491-2011 ANSI/ASHRAE 15-2016 ANSI/ASHRAE 20-1997 (RA06)
ASHRAE ASHRAE
ANSI/ASHRAE 22-2014 ANSI/ASHRAE 23.1-2010
ASHRAE
ANSI/ASHRAE 64-2011
ASME HEI HEI UL
ASME PTC 12.2-2010 HEI 2629 HEI 2634 ANSI/UL 207-2009
Commercial Systems Overview Residential Equipment Selection, 2nd ed. Commercial and Industrial Unitary Air-Conditioning Condensing Units Methods of Testing for Rating Positive Displacement Refrigerant Compressors and Condensing Units Heating and Cooling Equipment
ACCA ACCA AHRI ASHRAE
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ANSI/AHRI 366-2009 ANSI/ASHRAE 23.1-2010
UL/CSA
ANSI/UL 1995-2011/C22.2 No. 236-11
Containers
Series 1 Freight Containers—Classifications, Dimensions, and Ratings Series 1 Freight Containers—Specifications and Testing; Part 2: Thermal Containers Animal Environment in Cargo Compartments
ISO ISO SAE
ISO 668:2013 ISO 1496-2:2008 SAE AIR1600B-2010
Controls
Temperature Control Systems (2002) Field Testing of HVAC Controls Components Specifying Direct Digital Control Systems BACnet™—A Data Communication Protocol for Building Automation and Control Networks Method of Test for Conformance to BACnet® Method of Test for Rating Air Terminal Unit Controls Temperature-Indicating and Regulating Equipment Performance Requirements for Thermostats Used with Individual Room Electric Space Heating Devices Solid-State Controls for Appliances Limit Controls Primary Safety Controls for Gas- and Oil-Fired Appliances Temperature-Indicating and -Regulating Equipment Tests for Safety-Related Controls Employing Solid-State Devices Automatic Electrical Controls for Household and Similar Use; Part 1: General Requirements Guidelines for Boiler Control Systems (Gas/Oil Fired Boilers) (1998) Guideline for the Integration of Boilers and Automated Control Systems in Heating Applications (1998) Industrial Control and Systems: General Requirements Preventive Maintenance of Industrial Control and Systems Equipment Industrial Control and Systems, Controllers, Contactors, and Overload Relays Rated Not More than 2000 Volts AC or 750 Volts DC Industrial Control and Systems: Instructions for the Handling, Installation, Operation and Maintenance of Motor Control Centers Rated Not More than 600 Volts Industrial Control Equipment Manually Operated Gas Valves for Appliances, Appliance Connector Valves and Hose End Valves Gas Appliance Pressure Regulators
AABC ASHRAE ASHRAE ASHRAE
National Standards, Ch. 12 ASHRAE Guideline 11-2009 ASHRAE Guideline 13-2015 ANSI/ASHRAE 135-2016
ASHRAE ASHRAE CSA CSA
ANSI/ASHRAE 135.1-2013 ASHRAE 195-2013 C22.2 No. 24-15 CAN/CSA C828-13
UL UL UL UL UL UL
UL 244A-2003 ANSI/UL 353-1994 ANSI/UL 372-1994 UL 873-2007 UL 991-2004 UL 60730-1-2009
ABMA ABMA
ABMA 301 ABMA 306
NEMA NEMA NEMA
NEMA ICS 1-2000 (R2008) NEMA ICS 1.3-1986 (R2009) NEMA ICS 2-2000 (R2005)
NEMA
NEMA ICS 2.3-1995 (R2008)
UL CSA
ANSI/UL 508-2007 ANSI Z21.15-2009/CSA 9.1-2009 (R2014) ANSI Z21.18-2007/CSA 6.3-2007 (R2012)
Computers
Licensed for single user. © 2017 ASHRAE, Inc.
Condensers
Condensing Units
Commercial and Industrial
Residential
CSA
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Codes and Standards
40.7
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Coolers
Licensed for single user. © 2017 ASHRAE, Inc.
Air Drinking Water
Evaporative Food and Beverage
Liquid
Title
Publisher
Reference
Automatic Gas Ignition Systems and Components Gas Appliance Thermostats Manually-Operated Piezo-Electric Spark Gas Ignition Systems and Components
CSA CSA CSA
Manually Operated Electric Gas Ignition Systems and Components
CSA
Residential Controls—Electrical Wall-Mounted Room Thermostats Residential Controls—Surface Type Controls for Electric Storage Water Heaters Residential Controls—Temperature Limit Controls for Electric Baseboard Heaters Residential Controls—Hot-Water Immersion Controls Line-Voltage Integrally Mounted Thermostats for Electric Heaters Residential Controls—Class 2 Transformers Safety Guidelines for the Application, Installation, and Maintenance of Solid State Controls Electrical Quick-Connect Terminals
NEMA NEMA NEMA NEMA NEMA NEMA NEMA UL
ANSI Z21.20-2005 ANSI Z21.23-2010 ANSI Z21.77-2005/CSA 6.23-2005 (R2011) ANSI Z21.92-2001/CSA 6.29-2001 (R2012) NEMA DC 3-2013 NEMA DC 5-1989 (R2008) NEMA DC 10-2009 NEMA DC 12-1985 (R2013) NEMA DC 13-1979 (R2013) NEMA DC 20-1992 (R2009) NEMA ICS 1.1-1984 (R2009) ANSI/UL 310-2014
Refrigeration Equipment CSA Unit Coolers for Refrigeration AHRI Quality Maintenance of Commercial Refrigeration Systems ACCA Refrigeration Unit Coolers UL Methods of Testing Forced Convection and Natural Convection Air Coolers for Refrigeration ASHRAE Methods of Testing for Rating Drinking-Water Coolers with Self-Contained ASHRAE Mechanical Refrigeration Drinking-Water Coolers UL Drinking Water System Components—Health Effects NSF Method of Testing Direct Evaporative Air Coolers ASHRAE Method of Test for Rating Indirect Evaporative Coolers ASHRAE Milking Machine Installations—Vocabulary ASABE Methods of Testing for Rating Vending Machines for Sealed Beverages Methods of Testing for Rating Pre-Mix and Post-Mix Beverage Dispensing Equipment Manual Food and Beverage Dispensing Equipment Commercial Bulk Milk Dispensing Equipment Refrigerated Vending Machines Refrigerant-Cooled Liquid Coolers, Remote Type Methods of Testing for Rating Liquid Coolers Liquid Cooling Systems
Cooling Towers Cooling Tower Testing (2002) Commercial Systems Overview Bioaerosols: Assessment and Control (1999) Atmospheric Water Cooling Equipment Water-Cooling Towers Acceptance Test Code for Water-Cooling Towers Code for Measurement of Sound from Water Cooling Towers Nomenclature for Industrial Water Cooling Towers Recommended Practice for Airflow Testing of Cooling Towers Fiberglass-Reinforced Plastic Panels Certification of Water Cooling Tower Thermal Performance Crop Drying
Density, Specific Gravity, and Mass-Moisture Relationships of Grain for Storage Thermal Properties of Grain and Grain Products Moisture Relationships of Plant-Based Agricultural Products Dielectric Properties of Grain and Seed Construction and Rating of Equipment for Drying Farm Crops Resistance to Airflow of Grains, Seeds, Other Agricultural Products, and Perforated Metal Sheets Shelled Corn Storage Time for 0.5% Dry Matter Loss Moisture Measurement—Unground Grain and Seeds Moisture Measurement—Meat and Meat Products Moisture Measurement—Forages Moisture Measurement—Peanuts Thin-Layer Drying of Agricultural Crops Moisture Measurement—Tobacco Energy Efficiency of Peanut Drying Systems Temperature Sensor Locations for Seed-Cotton Drying Systems
CAN/CSA-C22.2 No. 120-13 ANSI/AHRI 420-2008 ANSI/ACCA 14 QMref-2015 ANSI/UL 412-2011 ANSI/ASHRAE 25-2001 (RA06) ANSI/ASHRAE 18-2008 (RA13)
ASHRAE ASHRAE
ANSI/UL 399-2008 ANSI/NSF 61-2014 ANSI/ASHRAE 133-2015 ANSI/ASHRAE 143-2015 ANSI/ASABE AD3918-2007 (R2011) ANSI/ASHRAE 32.1-2010 ANSI/ASHRAE 32.2-2003 (RA11)
NSF NSF UL AHRI ASHRAE SAE
ANSI/NSF 18-2012 ANSI/NSF 20-2012 ANSI/UL 541-2011 AHRI 480-2007 ANSI/ASHRAE 24-2013 SAE AIR1811A-1997 (R2010)
AABC ACCA ACGIH ASME NFPA CTI CTI CTI CTI CTI CTI ASABE ASABE ASABE ASABE ASABE ASABE
National Standards, Ch 13 ACCA Manual CS-1993 ACGIH ASME PTC 23-2003 (RA14) NFPA 214-2011 CTI ATC-105 (2000) CTI ATC-128 (2014) CTI BUL-109 (1997) CTI PFM-143 (1994) CTI STD-131 (2009) CTI STD-201RS (2013) ANSI/ASAE D241.4-1992 (R2012) ASAE D243.4-2003 (R2012) ASAE D245.6-2007 (R2012) ASAE D293.4-2012 ASAE S248.3-1976 (R2010) ASAE D272.3-1996 (R2011)
ASABE ASABE ASABE ASABE ASABE ASABE ASABE ASABE ASABE
ASAE D535-2005 (R2010) ASAE S352.2-1998 (R2012) ASAE S353-1972 (R2012) ANSI/ASAE S358.3-2012 ASAE S410.2-2010 ANSI/ASAE S448.1-2001 (R2012) ASAE S487-1987 (R2012) ASAE S488.1-2013 ASAE S530.1-2007 (R2012)
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40.8
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Title
Publisher
Reference
Dampers
Laboratory Methods of Testing Dampers for Rating Selecting Outdoor, Return, and Relief Dampers for Air-Side Economizer Systems
AMCA ASHRAE
AMCA 500-D-07 ASHRAE Guideline 16-2014
ACCA ACGIH AHAM ASHRAE
ACCA Manual CS-1993 ACGIH ANSI/AHAM DH-1-2008 ANSI/ASHRAE 139-2015
ASHRAE ASHRAE ASHRAE ASHRAE
ANSI/ASHRAE 160-2016 ANSI/ASHRAE 174-2009 ANSI/ASHRAE 190-2013 ANSI/ASHRAE 198-2013
ASME CSA CSA UL
PTC 12.4-1992 (RA14) C22.2 No. 92-1971 (R2013) CAN/CSA-C749-07 (R2012) ANSI/UL 474-2009
ASHRAE
ANSI/ASHRAE 35-2014
AHRI ASHRAE UL SAE SMACNA UL ACGIH ASME AWS NAIMA NAIMA SMACNA SMACNA SMACNA SMACNA SMACNA ADC ACCA ACCA ACCA NFPA NFPA ASTM
ANSI/AHRI 711-2009 ANSI/ASHRAE 63.1-1995 (RA01) ANSI/UL 207-2009 SAE AS1501C-1994 (R2013) SMACNA 1994 ANSI/UL 181-2013 ACGIH ASME B32.100-2005 AWS D9.1M/D9.1:2012 NAIMA AH116 NAIMA AH119 SMACNA 1995 SMACNA 2002 SMACNA 2003 SMACNA 1999 SMACNA 2007 ADC-91 ANSI/ACCA 5 QI-2015 ANSI/ACCA 9 QIvp-2016 ACCA 2016 NFPA 90A-2015 NFPA 90B-2015 ASTM A480/A480M-14b
ASTM
ASTM A568/A568M-14
ASTM
ASTM A653/A653M-13
ASTM
ASTM A924/A924M-14
ASTM
ASTM A1008/A1008M-13
ASTM
ASTM A1011/A1011M-14
ASTM ACCA ACCA ACCA ACCA ASHRAE
ASTM A1030/A1030M-11 ACCA Manual 4-1990 ANSI/ACCA Manual D-2016 ACCA Manual Q-1990 ACCA Manual T-2001 ANSI/ASHRAE 152-2014
UL UL AABC
ANSI/UL 181A-2013 ANSI/UL 181B-2013 National Standards, Ch 5
Dehumidifiers Commercial Systems Overview Bioaerosols: Assessment and Control (1999) Dehumidifiers Method of Testing for Rating Desiccant Dehumidifiers Utilizing Heat for the Regeneration Process Criteria for Moisture-Control Design Analysis in Buildings Method of Test for Rating Desiccant-Based Dehumidification Equipment Method of Testing for Rating Indoor Pool Dehumidifiers Method of Testing for Rating DX-Dedicated Outdoor Air Systems for Moisture Removal Capacity and Moisture Removal Efficiency Moisture Separator Reheaters Dehumidifiers Performance of Dehumidifiers Dehumidifiers Desiccants
Licensed for single user. © 2017 ASHRAE, Inc.
Method of Testing Desiccants for Refrigerant Drying
Liquid-Line Driers Method of Testing Liquid Line Refrigerant Driers Refrigerant-Containing Components and Accessories, Nonelectrical Ducts and Hose, Air Duct, Flexible Nonmetallic, Aircraft Fittings Ducted Electric Heat Guide for Air Handling Systems, 2nd ed. Factory-Made Air Ducts and Air Connectors Construction Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Preferred Metric Sizes for Flat, Round, Square, Rectangular, and Hexagonal Metal Products Sheet Metal Welding Code Fibrous Glass Duct Construction Standards, 5th ed. Residential Fibrous Glass Duct Construction Standards, 3rd ed. Thermoplastic Duct (PVC) Construction Manual, 2nd ed. Accepted Industry Practices for Sheet Metal Lagging, 1st ed. Fibrous Glass Duct Construction Standards, 7th ed. Round Industrial Duct Construction Standards, 2nd ed. Rectangular Industrial Duct Construction Standards, 2nd ed. Installation Flexible Duct Performance and Installation Standards, 5th ed. HVAC Quality Installation Specification HVAC Quality Installation Verification Protocols Technician’s Guide & Workbook for Duct Diagnostics and Repair Installation of Air-Conditioning and Ventilating Systems Installation of Warm Air Heating and Air-Conditioning Systems Material Specification for General Requirements for Flat-Rolled Stainless and Heat-Resisting Specifica- Steel Plate, Sheet and Strip tions Specification for Steel, Sheet, Carbon, Structural, and High-Strength, Low-Alloy, HotRolled and Cold-Rolled, General Requirements for Specification for Steel Sheet, Zinc-Coated (Galvanized) or Zinc-Iron Alloy-Coated (Galvannealed) by the Hot-Dipped Process Specification for General Requirements for Steel Sheet, Metallic-Coated by the Hot-Dip Process Specification for Steel, Sheet, Cold-Rolled, Carbon, Structural, High-Strength LowAlloy, High-Strength Low-Alloy with Improved Formability, Solution Hardened, and Bake Hardenable Specification for Steel, Sheet and Strip, Hot-Rolled, Carbon, Structural, High-Strength Low-Alloy, High-Strength Low-Alloy with Improved Formability, and Ultra-High Strength Practice for Measuring Flatness Characteristics of Coated Sheet Products System Installation Techniques for Perimeter Heating and Cooling Design Residential Duct Systems Commercial Low Pressure, Low Velocity Duct System Design Air Distribution Basics for Residential and Small Commercial Buildings Method of Test for Determining the Design and Seasonal Efficiencies of Residential Thermal Distribution Systems Closure Systems for Use with Rigid Air Ducts Closure Systems for Use with Flexible Air Ducts and Air Connectors Testing Duct Leakage Testing (2002) Driers
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Codes and Standards
40.9
Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
Balancing and Testing Air and Hydronic Systems HVAC Quality Installation Specification HVAC Quality Installation Verification Protocols Technician’s Guide & Workbook for Quality Installations Technician’s Guide & Workbook for Duct Diagnostics and Repair Flexible Air Duct Test Code, 3rd ed. Test Method for Measuring Acoustical and Airflow Performance of Duct Liner Materials and Prefabricated Silencers Method of Testing to Determine Flow Resistance of HVAC Ducts and Fittings Method of Testing HVAC Air Ducts HVAC Air Duct Leakage Test Manual, 2nd ed. HVAC Duct Systems Inspection Guide, 3rd ed.
ACCA ACCA ACCA ACCA ACCA ADC ASTM
ACCA Manual B 2009 ANSI/ACCA 5 QI-2015 ANSI/ACCA 9 QIvp-2016 ACCA 2015 ACCA 2016 ADC FD 72-R1 ASTM E477-13
ASHRAE ASHRAE SMACNA SMACNA
ANSI/ASHRAE 120-2008 ANSI/ASHRAE/SMACNA 126-2016 SMACNA 2012 SMACNA 2006
Economizers
Selecting Outdoor, Return, and Relief Dampers for Air-Side Economizer Systems
ASHRAE
ASHRAE Guideline 16-2014
Electrical
Electrical Power Systems and Equipment—Voltage Ratings Test Method for Bond Strength of Electrical Insulating Varnishes by the Helical Coil Test Standard Specification for Shelter, Electrical Equipment, Lightweight Canadian Electrical Code, Part I (22nd ed.), Safety Standard for Electrical Installations Part II—General Requirements ICC Electrical Code, Administrative Provisions (2006) Low Voltage Cartridge Fuses Industrial Control and Systems: Terminal Blocks Industrial Control and Systems: Enclosures Application Guide for Ground Fault Protective Devices for Equipment General Color Requirements for Wiring Devices Wiring Devices—Dimensional Specifications National Electrical Code® National Fire Alarm and Signaling Code Compatibility of Electrical Connectors and Wiring Molded-Case Circuit Breakers, Molded-Case Switches, and Circuit-Breaker Enclosures
ANSI ASTM
ANSI C84.1-2011 ASTM D2519-07 (2012)
ASTM CSA CSA ICC NEMA NEMA NEMA NEMA NEMA NEMA NFPA NFPA SAE UL
ASTM E2377-10 CSA C22.1-12 CAN/CSA-C22.2 No. 0-10 ICCEC NEMA FU 1-2012 NEMA ICS 4-2010 ANSI/NEMA ICS 6-1993 (R2006) ANSI/NEMA PB 2.2-2014 NEMA WD 1-1999 (R2010) ANSI/NEMA WD 6-2012 NFPA 70-2014 NFPA 72-2013 SAE AIR1329-2014 ANSI/UL 489-2013
AHRI ACCA ACCA ASHRAE ASHRAE ASHRAE ASHRAE ASHRAE
ANSI/AHRI 110-2012 ACCA Manual RS-1997 ACCA 2005 ASHRAE Guideline 14-2014 ANSI/ASHRAE/IES 90.1-2016 ANSI/ASHRAE/IES 90.2-2007 ANSI/ASHRAE/IES 100-2015 ANSI/ASHRAE 105-2014
ASHRAE ASHRAE
ANSI/ASHRAE 140-2011 ANSI/ASHRAE 152-2014
ASHRAE/ USGBC ICC/ ASHRAE ASME ICC ICC IAPMO NEMA
ANSI/ASHRAE/USGBC/IES 189.12014 ICC/ASHRAE 700-2015
NEMA SMACNA SMACNA SMACNA UL
NEMA MG 11-1977 (R2012) SMACNA 2013 SMACNA 2011 SMACNA 2014 UL 916-2007
AABC ACCA ACGIH AIHA
National Standards, Ch 10 ACCA Manual CS-1993 ACGIH ANSI/AIHA Z9.2-2012
Energy
Air-Conditioning and Refrigerating Equipment Nameplate Voltages Comfort, Air Quality, and Efficiency by Design Thermal Energy Storage Measurement of Energy and Demand Savings Energy Standard for Buildings Except Low-Rise Residential Buildings Energy-Efficient Design of Low-Rise Residential Buildings Energy Conservation in Existing Buildings Methods of Determining, Expressing, and Comparing Building Energy Performance and Greenhouse Gas Emissions Method of Test for the Evaluation of Building Energy Analysis Computer Programs Method of Test for Determining the Design and Seasonal Efficiencies of Residential Thermal Distribution Systems Standard for the Design of High-Performance, Green Buildings Except Low-Rise Residential Buildings National Green Building Standard Fuel Cell Power Systems Performance International Energy Conservation Code® (2015) International Green Construction Code™ (2012) Uniform Solar Energy Code (2012) Energy Management Guide for Selection and Use of Fixed Frequency Medium AC Squirrel-Cage Polyphase Induction Motors Energy Management Guide for Selection and Use of Single-Phase Motors HVAC Systems—Commissioning Manual, 2nd ed. Building Systems Analysis and Retrofit Manual, 2nd ed. Energy Systems Analysis and Management, 2nd ed. Energy Management Equipment
Exhaust Systems
Fan Systems: Supply/Return/Relief/Exhaust (2002) Commercial Systems Overview Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Fundamentals Governing the Design and Operation of Local Exhaust Ventilation Systems
PTC 50-2002 (RA14) IECC IGCC IAPMO NEMA MG 10-2013
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.10
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Expansion Valves
Title
Publisher
Reference
Spray Finishing Operations: Safety Code for Design, Construction, and Ventilation Laboratory Ventilation Recirculation of Air from Industrial Process Exhaust Systems Method of Testing Performance of Laboratory Fume Hoods Ventilation for Commercial Cooking Operations Performance Test Code on Compressors and Exhausters Flue and Exhaust Gas Analyses Mechanical Flue-Gas Exhausters Exhaust Systems for Air Conveying of Vapors, Gases, Mists, and Particulate Solids Draft Equipment
AIHA AIHA AIHA ASHRAE ASHRAE ASME ASME CSA NFPA UL
ANSI/AIHA Z9.3-2007 ANSI/AIHA Z9.5-2012 ANSI/AIHA Z9.7-2007 ANSI/ASHRAE 110-2016 ANSI/ASHRAE 154-2016 PTC 10-1997 (RA14) PTC 19.10-1981 CAN B255-M81 (R2005) ANSI/NFPA 91-2015 UL 378-2006
Thermostatic Refrigerant Expansion Valves Method of Testing Capacity of Thermostatic Refrigerant Expansion Valves
AHRI ASHRAE
ANSI/AHRI 750-2007 ANSI/ASHRAE 17-2015
ACGIH AHRI ASHRAE UL/CSA
ACGIH ANSI/AHRI 440-2008 ANSI/ASHRAE 79-2015 ANSI/UL 1995-2011/C22.2 No. 236-11
Residential Duct Systems Commercial Low Pressure, Low Velocity Duct System Design Balancing and Testing Air and Hydronic Systems HVAC Quality Installation Specifications Technician’s Guide & Workbook for Quality Installations Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Standards Handbook Air Systems Fans and Systems Troubleshooting Field Performance Measurement of Fan Systems Balance Quality and Vibration Levels for Fans Laboratory Methods of Testing Air Circulator Fans for Rating and Certification Laboratory Method of Testing Positive Pressure Ventilators for Rating Reverberant Room Method for Sound Testing of Fans Methods for Calculating Fan Sound Ratings from Laboratory Test Data Application of Sone Ratings for Non-Ducted Air Moving Devices Application of Sound Power Level Ratings for Fans Recommended Safety Practices for Users and Installers of Industrial and Commercial Fans Industrial Process/Power Generation Fans: Site Performance Test Standard Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Home Evaluation and Performance Improvement Mechanical Balance of Fans and Blowers Acoustics—Measurement of Airborne Noise Emitted and Structure-Borne Vibration Induced by Small Air-Moving Devices—Part 1: Airborne Noise Measurement Part 2: Structure-Borne Vibration
ACCA ACCA ACCA ACCA ACCA ACGIH AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA AMCA ACCA ACCA AHRI ASA
Laboratory Methods of Testing Fans for Aerodynamic Performance Rating
ASHRAE/ AMCA ASHRAE
ANSI/ACCA 1 Manual D-2016 ACCA Manual Q-2003 ACCA Manual B-2009 ANSI/ACCA 5 QI-2015 ACCA 2015 ACGIH AMCA 99-10 AMCA 200-95 (R2011) AMCA 201-02 (R2011) AMCA 202-98 (R2011) AMCA 203-90 (R2011) ANSI/AMCA 204-05 (R2012) ANSI/AMCA 230-12 ANSI/AMCA 240-06 AMCA 300-08 AMCA 301-90 AMCA 302-73 (R2012) AMCA 303-79 (R2012) AMCA 410-96 AMCA 803-02 (R2008) ANSI/ACCA 12 QH-2014 ACCA 2015 AHRI Guideline G-2011 ANSI/ASA S12.11-2013-Part 1/ISO 103021:2013 ANSI/ASA S12.11-2013-Part 2/ISO 10302-2:2013 ANSI/ASHRAE 51-2016 ANSI/AMCA 210-16 ANSI/ASHRAE 149-2013
ASHRAE ASME CSA CSA CSA UL UL
ANSI/ASHRAE 154-2016 ANSI/ASME PTC 11-2008 C22.2 No. 113-12 CAN/CSA-C814-10 CAN/CSA-F326-M91 (R2010) ANSI/UL 507-1999 ANSI/UL 705-2004
ASTM
ASTM E971-11
ASTM ASTM ASTM ASTM ASTM ASTM ASTM
ASTM E972-96 (2013) ASTM E1084-86 (2009) ASTM E1300-12ae1 ASTM E2112-07 ASTM E2188-10 ASTM E2189-10e1 ASTM E2190-10
Fan-Coil Units Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Room Fan-Coils Methods of Testing for Rating Fan-Coil Conditioners Heating and Cooling Equipment
Licensed for single user. © 2017 ASHRAE, Inc.
Fans
Laboratory Methods of Testing Fans Used to Exhaust Smoke in Smoke Management Systems Ventilation for Commercial Cooking Operations Fans Fans and Ventilators Energy Performance of Ceiling Fans Residential Mechanical Ventilating Systems Electric Fans Power Ventilators Fenestration
Practice for Calculation of Photometric Transmittance and Reflectance of Materials to Solar Radiation Test Method for Solar Photometric Transmittance of Sheet Materials Using Sunlight Test Method for Solar Transmittance (Terrestrial) of Sheet Materials Using Sunlight Practice for Determining Load Resistance of Glass in Buildings Practice for Installation of Exterior Windows, Doors and Skylights Test Method for Insulating Glass Unit Performance Test Method for Testing Resistance to Fogging Insulating Glass Units Specification for Insulating Glass Unit Performance and Evaluation
ASA
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.11
Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
Guide for Assessing the Durability of Absorptive Electrochemical Coatings within Sealed Insulating Glass Units Tables for Reference Solar Spectral Irradiance: Direct Normal and Hemispherical on 37° Tilted Surface Windows Fenestration Energy Performance Window, Door, and Skylight Installation
ASTM
ASTM E2354-10
ASTM
ASTM G173-03 (2012)
CSA CSA CSA
CSA A440-00 (R2005) CSA A440.2-14 CSA A440.4-07 (R2012)
Filter-Driers
Flow-Capacity Rating of Suction-Line Filters and Suction-Line Filter-Driers Method of Testing Liquid Line Filter-Drier Filtration Capability Method of Testing Flow Capacity of Suction Line Filters and Filter-Driers
AHRI ASHRAE ASHRAE
AHRI 731-2013 ANSI/ASHRAE 63.2-1996 (RA10) ANSI/ASHRAE 78-1985 (RA07)
Fireplaces
Factory-Built Fireplaces Fireplace Stoves
UL ANSI/UL 127-2011 UL ANSI/UL 737-2011 ASTM/NFPA ASTM E84-14 ASTM ASTM E119-14 ASTM ASTM E2257-13a ASTM ASTM E2307-15
Fire Protection Test Method for Surface Burning Characteristics of Building Materials Test Methods for Fire Test of Building Construction and Materials Test Method for Room Fire Test of Wall and Ceiling Materials and Assemblies Test Method for Determining Fire Resistance of Perimeter Fire Barriers Using Intermediate-Scale Multi-Story Test Apparatus Guide for Laboratory Monitors Test Method for Fire Resistance Grease Duct Enclosure Systems Practice for Specimen Preparation and Mounting of Paper or Vinyl Wall Coverings to Assess Surface Burning Characteristics BOCA National Fire Prevention Code, 11th ed. (1999) Uniform Fire Code International Fire Code® (2015) International Mechanical Code® (2015) International Urban-Wildland Interface Code® (2012) Fire-Resistance Tests—Elements of Building Construction; Part 1: Gen. Requirements Fire-Resistance Tests—Door and Shutter Assemblies Reaction to Fire Tests—Ignitability of Building Products Using a Radiant Heat Source Fire Containment—Elements of Building Construction—Part 1: Ventilation Ducts Fire Service Annunciator and Interface Fire Protection Handbook (2008) National Fire Codes® (issued annually) Fire Protection Guide to Hazardous Materials Fire Code Installation of Sprinkler Systems Flammable and Combustible Liquids Code Fire Protection for Laboratories Using Chemicals National Fire Alarm Code Fire Doors and Other Opening Protectives Health Care Facilities Life Safety Code® Methods of Fire Tests of Door Assemblies Fire, Smoke and Radiation Damper Installation Guide for HVAC Systems, 5th ed. Fire Tests of Door Assemblies Heat Responsive Links for Fire-Protection Service Fire Tests of Building Construction and Materials Fire Dampers Fire Tests of Through-Penetration Firestops Smoke The Commissioning Process for Smoke Control Systems ManageLaboratory Methods of Testing Fans Used to Exhaust Smoke in Smoke Management ment Systems Smoke-Control Systems Utilizing Barriers and Pressure Differences Smoke Management Systems in Malls, Atria, and Large Spaces Ceiling Dampers Smoke Dampers Freezers
Commercial
Quality Maintenance of Commercial Refrigeration Systems Energy Performance and Capacity of Household Refrigerators, Refrigerator-Freezers, Freezers, and Wine Chillers Energy Performance of Food Service Refrigerators and Freezers Refrigeration Equipment Dispensing Freezers
ASTM ASTM ASTM
ASTM E2335-12 ASTM E2336-14 ASTM E2404-13e1
BOCA IFCI ICC ICC ICC ISO ISO ISO ISO NEMA NFPA NFPA NFPA NFPA NFPA NFPA NFPA NFPA NFPA NFPA NFPA NFPA SMACNA UL UL UL UL UL ASHRAE ASHRAE
BNFPC UFC 1997 IFC IMC IUWIC ISO 834-1:1999 ISO 3008:2007 ISO 5657:1997 ISO 6944-1:2008 NEMA SB 30-2005 NFPA NFPA NFPA HAZ-2010 NFPA 1-2015 NFPA 13-2013 NFPA 30-2015 NFPA 45-2015 NFPA 72-2013 NFPA 80-2013 NFPA 99-2015 NFPA 101-2015 NFPA 252-2012 SMACNA 2002 ANSI/UL 10B-2008 ANSI/UL 33-2010 ANSI/UL 263-2011 ANSI/UL 555-2006 ANSI/UL 1479-2003 ASHRAE Guideline 1.5-2012 ANSI/ASHRAE 149-2013
NFPA NFPA UL UL
NFPA 92A-2009 NFPA 92B-2009 ANSI/UL 555C-2014 ANSI/UL 555S-2014
ACCA CSA
ANSI/ACCA 14 QMref-2015
CSA CSA NSF
C827-10 CAN/CSA-C22.2 No. 120-13 ANSI/NSF 6-2012
C300-12
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.12
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Household
Licensed for single user. © 2017 ASHRAE, Inc.
Fuels
Furnaces
Title
Publisher
Reference
Commercial Refrigerators and Freezers Commercial Refrigerators and Freezers Ice Makers Ice Cream Makers Refrigerators, Refrigerator-Freezers and Freezers Household Refrigerators and Freezers
NSF UL UL UL AHAM UL/CSA
ANSI/NSF 7-2014 ANSI/UL 471-2010 ANSI/UL 563-2009 ANSI/UL 621-2010 ANSI/AHAM HRF-1-2008 ANSI/UL 250-1993/C22.2 No. 63-93 (R2013)
Threshold Limit Values for Chemical Substances (updated annually) National Gas Fuel Code Reporting of Fuel Properties when Testing Diesel Engines with Alternative Fuels Derived from Plant Oils and Animal Fats Coal Pulverizers Classification of Coals by Rank Specification for Fuel Oils Specification for Diesel Fuel Oils Specification for Gas Turbine Fuel Oils Specification for Kerosine Practice for Receipt, Storage and Handling of Fuels Test Method for Determination of Yield Stress and Apparent Viscosity of Used Engine Oils at Low Temperature Test Method for Total Sulfur in Naphthas, Distillates, Reformulated Gasolines, Diesels, Biodiesels, and Motor Fuels by Oxidative Combustion and Electrochemical Detection Test Method for Determination of Homogeneity and Miscibility in Automotive Engine Oils Test Method for Measurement of Hindered Phenolic and Aromatic Amine Antioxidant Content in Non-Zinc Turbine Oils by Linear Sweep Voltammetry Practice for Enumeration of Viable Bacteria and Fungi in Liquid Fuels—Filtration and Culture Procedures Test Method for Evaluation of Automotive Engine Oils in the Sequence IIIF, SparkIgnition Engine Test Method for Determination of Ignition Delay and Derived Cetane Number (DCN) of Diesel Fuel Oils by Combustion in a Constant Volume Chamber Test Method for Determination of Total Sulfur in Light Hydrocarbon, Motor Fuels, and Oils by Online Gas Chromatography with Flame Photometric Detection Test Method for Sulfur in Gasoline, Diesel Fuel, Jet Fuel, Kerosine, Biodiesel, Biodiesel Blends, and Gasoline-Ethanol Blends by Monochromatic Wavelength Dispersive X-Ray Fluorescence Spectrometry Test Method for Flash Point by Modified Continuously Closed Cup (MCCCFP) Tester Test Method for Determining the Viscosity-Temperature Relationship of Used and Soot-Containing Engine Oils at Low Temperatures Test Method for Determination of Trace Elements in Middle Distillate Fuels by Inductively Coupled Plasma Atomic Emission Spectrometry (ICP-AES) Test Method for Determining Stability and Compatibility of Heavy Fuel Oils and Crude Oils by Heavy Fuel Oil Stability Analyzer (Optical Detection) Test Method for Determination of Intrinsic Stability of Asphaltene-Containing Residues, Heavy Fuel Oils, and Crude Oils (n-Heptane Phase Separation; Optical Detection) Test Method for Hydrogen Content of Middle Distillate Petroleum Products by LowResolution Pulsed Nuclear Magnetic Resonance Spectroscopy Gas-Fired Central Furnaces Gas Unit Heaters, Gas Packaged Heaters, Gas Utility Heaters and Gas-Fired Duct Furnaces Industrial and Commercial Gas-Fired Package Furnaces Uniform Mechanical Code (2012) Uniform Plumbing Code (2012) International Fuel Gas Code® (2015) Standard Gas Code (1999) Commercial-Industrial Gas Heating Equipment (2011)
ACGIH NFPA ASABE
ACGIH ANSI Z223.1/NPFA 54-2015 ASABE EP552.1-2009
ASME ASTM ASTM ASTM ASTM ASTM ASTM ASTM
PTC 4.2 1969 (RA09) ASTM D388-12 ASTM D396-15 ASTM D975-14b ASTM D2880-14a ASTM D3699-13be1 ASTM D4418-00 (2011) ASTM D6896-14
ASTM
ASTM D6920-13
ASTM
ASTM D6922-13
ASTM
ASTM D6971-09 (2014)
ASTM
ASTM D6974-09 (2013)e2
ASTM
ASTM D6984-14a
ASTM
ASTM D6890-13be1
ASTM
ASTM D7041-04 (2010)e1
ASTM
ASTM D7039-13
ASTM ASTM
ASTM D7094-12e1 ASTM D7110-14
ASTM
ASTM D7111-14
ASTM
ASTM D7112-12
ASTM
ASTM D7157-12
ASTM
ASTM D7171-05 (2011)
CSA CSA
ANSI Z21.47-2012/CSA 2.3-2012 ANSI Z83.8-2013/CSA-2.6-2013
CSA IAPMO IAPMO ICC SBCCI UL
CGA 3.2-1976 (R2009) Chapter 13 Chapter 12 IFGC SGC UL 795-2011
Commercial Systems Overview Residential Equipment Selection, 2nd ed. HVAC Quality Installation Specification Technician’s Guide & Workbook for Quality Installations Method of Testing for Annual Fuel Utilization Efficiency of Residential Central Furnaces and Boilers
ACCA ACCA ACCA ACCA ASHRAE
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ANSI/ACCA 5 QI-2015 ACCA 2015 ANSI/ASHRAE 103-2007
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.13
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Gas
Licensed for single user. © 2017 ASHRAE, Inc.
Oil
Solid Fuel
Green Buildings
Heaters
Title
Publisher
Reference
Prevention of Furnace Explosions/Implosions in Multiple Burner Boilers Residential Gas Detectors Heating and Cooling Equipment Single and Multiple Station Carbon Monoxide Alarms National Fuel Gas Code Gas-Fired Central Furnaces Gas Unit Heaters, Gas Packaged Heaters, Gas Utility Heaters and Gas-Fired Duct Furnaces Industrial and Commercial Gas-Fired Package Furnaces International Fuel Gas Code® (2015) Standard Gas Code (1999) Commercial-Industrial Gas Heating Equipment Specification for Fuel Oils Specification for Diesel Fuel Oils Test Method for Smoke Density in Flue Gases from Burning Distillate Fuels Standard Test Method for Vapor Pressure of Liquefied Petroleum Gases (LPG) (Expansion Method) Oil Burning Stoves and Water Heaters Oil-Fired Warm Air Furnaces Installation of Oil-Burning Equipment Oil-Fired Central Furnaces Oil-Fired Floor Furnaces Oil-Fired Wall Furnaces Installation Code for Solid-Fuel-Burning Appliances and Equipment Solid-Fuel-Fired Central Heating Appliances Solid-Fuel and Combination-Fuel Central and Supplementary Furnaces
NFPA UL UL/CSA UL AGA/NFPA CSA CSA
NFPA 8502-99 ANSI/UL 1484-2000 ANSI/UL 1995-2011/C22.2 No. 236-11 ANSI/UL 2034-2008 ANSI Z223.1/NFPA 54-2015 ANSI Z21.47-2012/CSA 2.3-2012 ANSI Z83.8-2013/CSA-2.6-2013
CSA ICC SBCCI UL ASTM ASTM ASTM ASTM
CGA 3.2-1976 (R2009) IFGC SGC UL 795-2011 ASTM D396-15 ASTM D975-14b ASTM D2156-09 (2013) ASTM D6897-09
CSA CSA NFPA UL UL UL CSA CSA UL
B140.3-1962 (R2011) B140.4-04 (R2009) NFPA 31-2011 UL 727-2006 ANSI/UL 729-2003 ANSI/UL 730-2003 CSA B365-10 CAN/CSA-B366.1-11 ANSI/UL 391-2010 ANSI/ASHRAE/USGBC/IES 189.1-2014 ICC/ASHRAE 700-2015
International Green Construction Code™ (2012) Building Systems Analysis and Retrofit Manual, 2nd ed.
ASHRAE/ USGBC ICC/ ASHRAE ICC SMACNA
Gas-Fired High-Intensity Infrared Heaters Gas-Fired Low-Intensity Infrared Heaters
CSA CSA ACGIH ACGIH ASABE ASME ASTM CSA CSA CSA HEI UL UL UL UL UL SAE SAE
ANSI Z83.19-2009/CSA 2.35-2009 ANSI Z83.20-2008/CSA 2.34-2008 (R2013) ACGIH ACGIH ANSI/ASAE S423-1991 (R2012) ASME PTC 4.3-1968 (RA91) ASTM E1602-03 (2010)e1 ANSI Z83.4-2013/CSA 3.7-2013 C22.2 No. 155-M1986 (R2013) CAN3-B140.9.3-M86 (R2011) HEI 2622 ANSI/UL 499-2014 ANSI/UL 574-2003 UL 733-1993 ANSI/UL 875-2009 ANSI/UL 896-1993 SAE J1310-2011 SAE J1350-2011
SAE ASABE CSA CSA CSA SAE UL UL ACCA ASHRAE
SAE J1422-2011 ASAE EP258.4-2014 ANSI Z83.7-2011/CSA 2.14-2011 ANSI Z83.18-2015 B140.8-1967 (R2011) SAE J1024-2011 UL 795-2011 ANSI/UL 823-2006 ANSI/ACCA 10 SPS-2010 ANSI/ASHRAE 146-2011
Standard for the Design of High-Performance, Green Buildings Except Low-Rise Residential Buildings National Green Building Standard
Threshold Limit Values for Chemical Substances (updated annually) Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Thermal Performance Testing of Solar Ambient Air Heaters Air Heaters Guide for Construction of Solid Fuel Burning Masonry Heaters Non-Recirculating Direct Gas-Fired Industrial Air Heaters Electric Duct Heaters Portable Kerosene-Fired Heaters Standards for Closed Feedwater Heaters, 8th ed. (2009) Electric Heating Appliances Electric Oil Heaters Oil-Fired Air Heaters and Direct-Fired Heaters Electric Dry Bath Heaters Oil-Burning Stoves Engine Electric Engine Preheaters and Battery Warmers for Diesel Engines Selection and Application Guidelines for Diesel, Gasoline, and Propane Fired Liquid Cooled Engine Pre-Heaters Fuel Warmer—Diesel Engines Nonresidential Installation of Electric Infrared Brooding Equipment Gas-Fired Construction Heaters Recirculating Direct Gas-Fired Industrial Air Heaters Portable Industrial Oil-Fired Heaters Fuel-Fired Heaters—Air Heating—for Construction and Industrial Machinery Commercial-Industrial Gas Heating Equipment Electric Heaters for Use in Hazardous (Classified) Locations Pool HVAC Design for Swimming Pools and Spas Methods of Testing and Rating Pool Heaters
IGCC SMACNA 2011
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.14
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Room
Transport
Unit
Licensed for single user. © 2017 ASHRAE, Inc.
Heat Exchangers
Heating
Heat Pumps
Title
Publisher
Reference
Gas-Fired Pool Heaters Oil-Fired Service Water Heaters and Swimming Pool Heaters Specification for Room Heaters, Pellet Fuel-Burning Type Gas-Fired Room Heaters, Vol. II, Unvented Room Heaters Gas-Fired Unvented Catalytic Room Heaters for Use with Liquefied Petroleum (LP) Gases Vented Gas-Fired Space Heating Appliances
CSA CSA ASTM CSA CSA CSA
Vented Gas Fireplace Heaters Unvented Kerosene-Fired Room Heaters and Portable Heaters Movable and Wall- or Ceiling-Hung Electric Room Heaters Fixed and Location-Dedicated Electric Room Heaters Solid Fuel-Type Room Heaters Heater, Airplane, Engine Exhaust Gas to Air Heat Exchanger Type Heater, Aircraft Internal Combustion Heat Exchanger Type Motor Vehicle Heater Test Procedure Gas Unit Heaters, Gas Packaged Heaters, Gas Utility Heaters and Gas-Fired Duct Furnaces Oil-Fired Unit Heaters
CSA UL UL UL UL SAE SAE SAE CSA
ANSI Z21.56-2014/CSA 4.7-2014 B140.12-03 (R2013) ASTM E1509-12 ANSI Z21.11.2-2013 ANSI Z21.76-1994 (R2006) ANSI Z21.86-2008/CSA 2.32-2008 (R2014) ANSI Z21.88-2014/CSA 2.33-2014 UL 647-1993 UL 1278-2014 UL 2021-2013 ANSI/UL 1482-2011 SAE ARP86-2011 SAE AS8040B-2013 SAE J638-2011 ANSI Z83.8-2013/CSA-2.6-2013
UL
ANSI/UL 731-1995
Remote Mechanical-Draft Evaporative Refrigerant Condensers Method of Testing Air-to-Air Heat/Energy Exchangers Boiler and Pressure Vessel Code—Section VIII, Division 1: Pressure Vessels Single Phase Heat Exchangers Air Cooled Heat Exchangers Laboratory Methods of Test for Rating the Performance of Heat/Energy-Recovery Ventilators Standards for Gasketed Plate Heat Exchangers, 1st ed. Standards for Shell and Tube Heat Exchangers, 5th ed. (2013) Standards of Tubular Exchanger Manufacturers Association, 9th ed. (2007) Refrigerant-Containing Components and Accessories, Nonelectrical
AHRI ASHRAE ASME ASME ASME CSA
ANSI/AHRI 491-2011 ANSI/ASHRAE 84-2013 ASME BPVC-2015 ASME PTC 12.5-2000 (RA05) ASME PTC 30-1991 (RA11) C439-09 (R2014)
HEI HEI TEMA UL
HEI HEI 2623 TEMA ANSI/UL 207-2009
Commercial Systems Overview HVAC Quality Installation Specification Technician’s Guide & Workbook for Quality Installations Residential Load Calculations Comfort, Air Quality, and Efficiency by Design Residential Equipment Selection, 2nd ed. Heating, Ventilating and Cooling Greenhouses Peak Cooling and Heating Load Calculations in Buildings Except Low-Rise Residential Buildings Heater Elements Determining the Required Capacity of Residential Space Heating and Cooling Appliances Heat Loss Calculation Guide (2001) Residential Hydronic Heating Installation Design Guide Radiant Floor Heating (1995) Advanced Installation Guide (Commercial) for Hot Water Heating Systems (2001) Environmental Systems Technology, 2nd ed. (1999) Pulverized Fuel Systems Aircraft Electrical Heating Systems Heating Value of Fuels Performance Test for Air-Conditioned, Heated, and Ventilated Off-Road SelfPropelled Work Machines HVAC Systems Applications, 2nd ed. Electric Baseboard Heating Equipment Electric Duct Heaters Heating and Cooling Equipment
ACCA ACCA ACCA ACCA ACCA ACCA ASABE ASHRAE/ ACCA CSA CSA
ACCA Manual CS-1993 ANSI/ACCA 5 QI-2015 ACCA 2015 ANSI/ACCA 2 Manual J-2016 ACCA Manual RS-1997 ANSI/ACCA 3 Manual S-2014 ANSI/ASAE EP406.4-2003 (R2008) ANSI/ASHRAE/ACCA 183-2007 (RA14) C22.2 No. 72-10 (R2014) CAN/CSA-F280-12
HYDI HYDI HYDI HYDI NEBB NFPA SAE SAE SAE
HYDI H-22 IBR Guide HYDI 004 HYDI 250 NEBB NFPA 8503-97 SAE AIR860B-2011 SAE J1498-2011 SAE J1503-2004
SMACNA UL UL UL/CSA
SMACNA 2010 ANSI/UL 1042-2009 ANSI/UL 1996-2009 ANSI/UL 1995-2011/C22.2 No. 236-11
Commercial Systems Overview Geothermal Heat Pump Training Certification Program Heat Pumps Systems, Principles and Applications, 2nd ed. Residential Equipment Selection, 2nd ed. Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Performance Rating of Unitary Air-Conditioning and Air-Source Heat Pump Equipment Commercial and Industrial Unitary Air-Conditioning and Heat Pump Equipment
ACCA ACCA ACCA ACCA ACGIH AHRI
ACCA Manual CS-1993 ACCA Training Manual ACCA Manual H-1984 ANSI/ACCA 3 Manual S-2014 ACGIH ANSI/AHRI 210/240-2008
AHRI
ANSI/AHRI 340/360-2007
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.15
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Licensed for single user. © 2017 ASHRAE, Inc.
Gas-Fired
Title
Publisher
Reference
Single Package Vertical Air-Conditioner and Heat Humps Direct Geoexchange Heat Pumps Variable Refrigerant Flow (VRF) Multi-Split Air-Conditioning and Heat Pump Equipment Methods of Testing for Rating Electrically Driven Unitary Air-Conditioning and Heat Pump Equipment Methods of Testing for Rating Seasonal Efficiency of Unitary Air-Conditioners and Heat Pumps Method of Test for Direct-Expansion Ground Source Heat Pumps Method of Testing for Rating of Multi-Purpose Heat Pumps for Residential Space Conditioning and Water Heating Performance Standard for Split-System and Single-Package Central Air Conditioners and Heat Pumps Installation of Air-Source Heat Pumps and Air Conditioners Performance of Direct-Expansion (DX) Ground-Source Heat Pumps Water-Source Heat Pumps—Testing and Rating for Performance, Part 1: Water-to-Air and Brine-to-Air Heat Pumps Part 2: Water-to-Water and Brine-to-Water Heat Pumps Heating and Cooling Equipment (2011) Gas-Fired, Heat Activated Air Conditioning and Heat Pump Appliances
AHRI AHRI AHRI
ANSI/AHRI 390-2003 ANSI/AHRI 870-2005 ANSI/AHRI 1230-2010
ASHRAE
ANSI/ASHRAE 37-2009
ASHRAE
ANSI/ASHRAE 116-2010
ASHRAE ASHRAE
ANSI/ASHRAE 194-2012 ANSI/ASHRAE 206-2013
CSA
CAN/CSA-C656-14
CSA CSA CSA
CAN/CSA-C273.5-11 C748-13 CAN/CSA-C13256-1-01 (R2011)
CSA UL/CSA CSA
CAN/CSA-C13256-2-01 (R2010) ANSI/UL 1995/C22.2 No. 236-11 ANSI Z21.40.1-1996/CGA 2.91-M96 (R2012) ANSI Z21.40.2-1996/CGA 2.92-M96 (R2002) ANSI Z21.40.4-1996/CGA 2.94-M96 (R2002)
Gas-Fired, Work Activated Air Conditioning and Heat Pump Appliances (Internal Combustion) Performance Testing and Rating of Gas-Fired Air Conditioning and Heat Pump Appliances Heat Recovery Gas Turbine Heat Recovery Steam Generators Water Heaters, Hot Water Supply Boilers, and Heat Recovery Equipment
CSA CSA ASME NSF
ANSI/ASME PTC 4.4-2008 ANSI/NSF 5-2012
HighPerformance Buildings
Sustainable, High Performance Operations and Maintenance
ASHRAE
ASHRAE Guideline 32-2012
Standard for the Design of High-Performance, Green Buildings Except Low-Rise Residential Buildings International Green Construction Code™ (2012)
ASHRAE/ USGBC ICC
ANSI/ASHRAE/USGBC/IES 189.12014 IGCC
Humidifiers
Commercial Systems Overview Residential Equipment Selection Comfort, Air Quality, and Efficiency by Design Bioaerosols: Assessment and Control (1999) Humidifiers Central System Humidifiers for Residential Applications Self-Contained Humidifiers for Residential Applications Commercial and Industrial Humidifiers Method of Test for Residential Central-System Humidifiers Humidifiers
ACCA ACCA ACCA ACGIH AHAM AHRI AHRI AHRI ASHRAE UL/CSA
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ACCA Manual RS-1997 ACGIH ANSI/AHAM HU-1-2006 (R2011) ANSI/AHRI 610-2014 ANSI/AHRI 620-2014 ANSI/AHRI 640-2005 ANSI/ASHRAE 164.1-2012 ANSI/UL 998-2011/C22.2 No. 104-11
HVAC Load Calculations
Residential Load Calculations
ACCA
ANSI/ACCA 2 Manual J-2016
Commercial Load Calculations Peak Cooling and Heating Load Calculations in Buildings Except Low-Rise Residential Buildings
ACCA ASHRAE/ ACCA
ACCA Manual N-2012 ANSI/ASHRAE/ACCA 183-2007 (RA14)
Ice Makers
Quality Maintenance of Commercial Refrigeration Systems Performance Rating of Automatic Commercial Ice Makers Ice Storage Bins Methods of Testing Automatic Ice Makers Refrigeration Equipment Energy Performance of Automatic Icemakers and Ice Storage Bins Automatic Ice Making Equipment Ice Makers
ACCA AHRI AHRI ASHRAE CSA CSA NSF UL
ANSI/ACCA 14 QMref-2015 ANSI/AHRI 811-2012 ANSI/AHRI 821-2012 ANSI/ASHRAE 29-2015 C22.2 No. 120-13 CAN/CSA-C742-08 (R2014) ANSI/NSF 12-2012 ANSI/UL 563-2009
Incinerators
Incinerators and Waste and Linen Handling Systems and Equipment Residential Incinerators
NFPA UL
NFPA 82-2014 UL 791-2006
Indoor Air Quality
Good HVAC Practices for Residential and Commercial Buildings (2003) Comfort, Air Quality, and Efficiency by Design Bioaerosols: Assessment and Control (1999) Interactions Affecting the Achievement of Acceptable Indoor Environments Ventilation for Acceptable Indoor Air Quality Ventilation and Acceptable Indoor Air Quality in Low-Rise Residential Buildings
ACCA ACCA ACGIH ASHRAE ASHRAE ASHRAE
ACCA ACCA Manual RS-1997 ACGIH ASHRAE Guideline 10-2016 ANSI/ASHRAE 62.1-2016 ANSI/ASHRAE 62.2-2016
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.16
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Aircraft
Modeling Insulation
Title
Publisher
Reference
Test Method for Determination of Volatile Organic Chemicals in Atmospheres (Canister Sampling Methodology) Guide for Using Probability Sampling Methods in Studies of Indoor Air Quality in Buildings Guide for Using Indoor Carbon Dioxide Concentrations to Evaluate Indoor Air Quality and Ventilation Guide for Placement and Use of Diffusion Controlled Passive Monitors for Gaseous Pollutants in Indoor Air Test Method for Determination of Metals and Metalloids Airborne Particulate Matter by Inductively Coupled Plasma Atomic Emissions Spectrometry (ICP-AES) Test Method for Metal Removal Fluid Aerosol in Workplace Atmospheres Practice for Emission Cells for the Determination of Volatile Organic Emissions from Materials/Products Practice for Collection of Surface Dust by Micro-Vacuum Sampling for Subsequent Metals Determination Test Method for Determination of Beryllium in the Workplace Using Field-Based Extraction and Optical Fluorescence Detection Practice for Referencing Suprathreshold Odor Intensity Guide for Specifying and Evaluating Performance of a Single Family Attached and Detached Dwelling—Indoor Air Quality Classification for Serviceability of an Office Facility for Thermal Environment and Indoor Air Conditions Practice for Continuous Sizing and Counting of Airborne Particles in Dust-Controlled Areas and Clean Rooms Using Instruments Capable of Detecting Single SubMicrometre and Larger Particles Ambient Air—Determination of Mass Concentration of Nitrogen Dioxide—Modified Griess-Saltzman Method Air Quality—Exchange of Data—Part 1: General Data Format Part 2: Condensed Data Format Environmental Tobacco Smoke—Estimation of Its Contribution to Respirable Suspended Particles—Determination of Particulate Matter by Ultraviolet Absorptance and by Fluorescence Indoor Air—Part 3: Determination of Formaldehyde and Other Carbonyl Compounds in Indoor Air and Test Chamber Air—Active Sampling Method Workplace Air Quality—Sampling and Analysis of Volatile Organic Compounds by Solvent Desorption/Gas Chromatography—Part 1: Pumped Sampling Method Part 2: Diffusive Sampling Method Workplace Air Quality—Determination of Total Organic Isocyanate Groups in Air Using 1-(2-Methoxyphenyl) Piperazine and Liquid Chromatography Installation of Carbon Monoxide (CO) Detection and Warning Equipment Indoor Air Quality—A Systems Approach, 3rd ed. IAQ Guidelines for Occupied Buildings Under Construction, 2nd ed. Single and Multiple Station Carbon Monoxide Alarms Air Quality Within Commercial Aircraft Air Quality Within Commercial Aircraft Guide for Selecting Instruments and Methods for Measuring Air Quality in Aircraft Cabins Guide for Deriving Acceptable Levels of Airborne Chemical Contaminants in Aircraft Cabins Based on Health and Comfort Considerations Guideline for Documenting Indoor Airflow and Contaminant Transport Modeling
ASTM
ASTM D5466-01 (2007)
ASTM
ASTM D5791-95 (2012)e1
ASTM
ASTM D6245-12
ASTM
ASTM D6306-10
ASTM
ASTM D7035-10
ASTM ASTM
ASTM D7049-04 (2010) ASTM D7143-11
ASTM
ASTM D7144-05a (2011)
ASTM
ASTM D7202-14e1
ASTM ASTM
ASTM E544-10 ASTM E2267-04 (2013)
ASTM
ASTM E2320-04 (2012)
ASTM
ASTM F50-12
ISO
ISO 6768:1998
ISO ISO ISO
ISO 7168-1:1999 ISO 7168-2:1999 ISO 15593:2001
ISO
ISO 16000-3:2011
ISO
ISO 16200-1:2001
ISO ISO
ISO 16200-2:2000 ISO 16702:2007
NFPA SMACNA SMACNA UL ASHRAE ASHRAE ASTM ASTM
NFPA 720-2015 SMACNA 1998 ANSI/SMACNA 008-2007 ANSI/UL 2034-2008 ASHRAE Guideline 28-2016 ANSI/ASHRAE 161-2013 ASTM D6399-10 ASTM D7034-11
Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Home Evaluation and Performance Improvement Guidelines for Use of Thermal Insulation in Agricultural Buildings Terminology Relating to Thermal Insulating Materials Test Method for Steady-State Heat Flux Measurements and Thermal Transmission Properties by Means of the Guarded-Hot-Plate Apparatus Test Method for Steady-State Heat Transfer Properties of Pipe Insulations Practice for Fabrication of Thermal Insulating Fitting Covers for NPS Piping and Vessel Lagging Test Method for Steady-State Thermal Transmission Properties by Means of the Heat Flow Meter Apparatus Specification for Preformed Flexible Elastometric Cellular Thermal Insulation in Sheet and Tubular Form Specification for Cellular Glass Thermal Insulation Specification for Rigid, Cellular Polystyrene Thermal Insulation
ASHRAE
ASHRAE Guideline 33-2013
ACCA ACCA ASABE ASTM ASTM
ANSI/ACCA 12 QH-2014 ACCA 2014 ANSI/ASAE S401.2-1993 (R2012) ASTM C168-13 ASTM C177-13
ASTM ASTM
ASTM C335/C335M-10e1 ASTM C450-08 (2014)
ASTM
ASTM C518-10
ASTM
ASTM C534/C534M-14
ASTM ASTM
ASTM C552-14 ASTM C578-14a
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.17
Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
Practice for Inner and Outer Diameters of Thermal Insulation for Nominal Sizes of Pipe and Tubing Specification for Unfaced Preformed Rigid Cellular Polyisocyanurate Thermal Insulation Practice for Estimate of the Heat Gain or Loss and the Surface Temperatures of Insulated Flat, Cylindrical, and Spherical Systems by Use of Computer Programs Specification for Adhesives for Duct Thermal Insulation Classification of Potential Health and Safety Concerns Associated with Thermal Insulation Materials and Accessories Practice for Thermographic Inspection of Insulation Installations in Envelope Cavities of Frame Buildings Specification for Fibrous Glass Duct Lining Insulation (Thermal and Sound Absorbing Material) Specification for Faced or Unfaced Rigid Cellular Phenolic Thermal Insulation Test Method for Thermal Performance of Building Materials and Envelope Assemblies by Means of a Hot Box Apparatus Specification for Perpendicularly Oriented Mineral Fiber Roll and Sheet Thermal Insulation for Pipes and Tanks Guide for Measuring and Estimating Quantities of Insulated Piping and Components Specification for Cellular Melamine Thermal and Sound-Absorbing Insulation Guide for Selecting Jacketing Materials for Thermal Insulation Specification for Preformed Flexible Cellular Polyolefin Thermal Insulation in Sheet and Tubular Form Specification for Polyimide Flexible Cellular Thermal and Sound Absorbing Insulation Specification for Cellulosic Fiber Stabilized Thermal Insulation Test Method for Characterizing the Effect of Exposure to Environmental Cycling on Thermal Performance of Insulation Products Specification for Flexible Polymeric Foam Sheet Insulation Used as a Thermal and Sound Absorbing Liner for Duct Systems Standard Guide for Development of Standard Data Records for Computerization of Thermal Transmission Test Data for Thermal Insulation Guide for Determining Blown Density of Pneumatically Applied Loose Fill Mineral Fiber Thermal Insulation Test Method for Determining the Moisture Content of Organic and Inorganic Insulation Materials by Weight Classification for Rating Sound Insulation Test Method for Determining the Drainage Efficiency of Exterior Insulation and Finish Systems (EIFS) Clad Wall Assemblies Practice for Use of Test Methods E96/E96M for Determining the Water Vapor Transmission (WVT) of Exterior Insulation and Finish Systems Thermal Insulation—Vocabulary National Commercial and Industrial Insulation Standards, 7th ed. Accepted Industry Practices for Sheet Metal Lagging, 1st ed.
ASTM
ASTM C585-10
ASTM
ASTM C591-13
ASTM
ASTM C680-14
ASTM ASTM
ASTM C916-14 ASTM C930-12
ASTM
ASTM C1060-11a
ASTM
ASTM C1071-12
ASTM ASTM
ASTM C1126-14 ASTM C1363-11
ASTM
ASTM C1393-14
ASTM ASTM ASTM ASTM
ASTM C1409-12 ASTM C1410-14 ASTM C1423-14 ASTM C1427-13
ASTM ASTM ASTM
ASTM C1482-12 ASTM C1497-12 ASTM C1512-10
ASTM
ASTM C1534-14
ASTM
ASTM C1558-12
ASTM
ASTM C1574-04 (2013)
ASTM
ASTM C1616-07 (2012)
ASTM ASTM
ASTM E413-10 ASTM E2273-03 (2011)
ASTM
ASTM E2321-03 (2011)
ISO MICA SMACNA
ISO 9229:2007 MICA SMACNA 2002
Legionellosis
Minimizing the Risk of Legionellosis Associated with Building Water Systems Legionellosis: Risk Management for Building Water Systems
ASHRAE ASHRAE
ASHRAE Guideline 12-2000 ANSI/ASHRAE 188-2015
Louvers
Laboratory Methods of Testing Dampers for Rating Laboratory Methods of Testing Louvers for Rating
AMCA AMCA
AMCA 500-D-12 AMCA 500-L-12
Lubricants
Methods of Testing the Floc Point of Refrigeration Grade Oils Refrigeration Oil Description Test Method for Pour Point of Petroleum Products Classification of Industrial Fluid Lubricants by Viscosity System Test Method for Relative Molecular Weight (Relative Molecular Mass) of Hydrocarbons by Thermoelectric Measurement of Vapor Pressure Test Method for Determination of Moderately High Temperature Piston Deposits by Thermo-Oxidation Engine Oil Simulation Test—TEOST MHT Petroleum Products—Corrosiveness to Copper—Copper Strip Test
ASHRAE ASHRAE ASTM ASTM ASTM
ANSI/ASHRAE 86-2013 ANSI/ASHRAE 99-2006 ASTM D97-12 ASTM D2422-97 (2013) ASTM D2503-92 (2012)
ASTM
ASTM D7097-09
ISO
ISO 2160:1998
Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Engineering Analysis of Experimental Data Instrumentation for Monitoring Central Chilled Water Plant Efficiency Standard Method for Measuring the Proportion of Lubricant in Liquid Refrigerant Standard Method for Measurement of Moist Air Properties Methods of Determining, Expressing, and Comparing Building Energy Performance and Greenhouse Gas Emissions
ACGIH ASHRAE ASHRAE ASHRAE ASHRAE ASHRAE
ACGIH ASHRAE Guideline 2-2010 (RA14) ASHRAE Guideline 22-2012 ANSI/ASHRAE 41.4-2015 ANSI/ASHRAE 41.6-2014 ANSI/ASHRAE 105-2014
Measurement
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.18
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Method for Establishing Installation Effects on Flowmeters Test Uncertainty Measurement of Industrial Sound Test Methods for Water Vapor Transmission of Materials Specification and Temperature-Electromotive Force (EMF) Tables for Standardized Thermocouples Practice for Continuous Sizing and Counting of Airborne Particles in Dust-Controlled Areas and Clean Rooms Using Instruments Capable of Detecting Single Sub-Micrometre and Larger Particles Ergonomics of the Thermal Environment—Instruments for Measuring Physical Quantities Ergonomics of the Thermal Environment—Determination of Metabolic Rate Ergonomics of the Thermal Environment—Estimation of the Thermal Insulation and Water Vapour Resistance of a Clothing Ensemble Energy Measurement of Energy and Demand Savings Fluid Flow Standard Methods for Liquid in Flow Measurement Methods for Volatile-Refrigerant Mass Flow Measurements Using Calorimeters Flow Measurement Glossary of Terms Used in the Measurement of Fluid Flow in Pipes Measurement Uncertainty for Fluid Flow in Closed Conduits Measurement of Fluid Flow in Pipes Using Orifice, Nozzle, and Venturi Measurement of Fluid Flow in Pipes Using Vortex Flowmeters Fluid Flow in Closed Conduits: Connections for Pressure Signal Transmissions Between Primary and Secondary Devices Measurement of Liquid Flow in Closed Conduits by Weighing Method Measurement of Fluid Flow Using Small Bore Precision Orifice Meters Measurement of Fluid Flow in Closed Conduits by Means of Electromagnetic Flowmeters Measurement of Fluid Flow Using Variable Area Meters Test Method for Determining the Moisture Content of Inorganic Insulation Materials by Weight Test Method for Indicating Wear Characteristics of Petroleum Hydraulic Fluids in a High Pressure Constant Volume Vane Pump Test Method for Dynamic Viscosity and Density of Liquids by Stabinger Viscometer (and the Calculation of Kinematic Viscosity) Test Method for Indicating Wear Characteristics of Non-Petroleum and Petroleum Hydraulic Fluids in a Constant Volume Vane Pump Practice for Calculating Viscosity of a Blend of Petroleum Products Test Method for Same-Different Test Practice for Field Use of Pyranometers, Pyrheliometers, and UV Radiometers Gas Flow Standard Methods for Laboratory Airflow Measurement Method of Test for Measurement of Flow of Gas Measurement of Gas Flow by Turbine Meters Measurement of Gas Flow by Means of Critical Flow Venturi Nozzles Pressure Standard Method for Pressure Measurement Pressure Gauges and Gauge Attachments Pressure Measurement Temperature Standard Method for Temperature Measurement Thermometers, Direct Reading and Remote Reading Temperature Measurement Total Temperature Measuring Instruments (Turbine Powered Subsonic Aircraft) Thermal Method of Testing Thermal Energy Meters for Liquid Streams in HVAC Systems Test Method for Steady-State Heat Flux Measurements and Thermal Transmission Properties by Means of the Guarded-Hot-Plate Apparatus Test Method for Steady-State Heat Flux Measurements Thermal Transmission Properties by Means of the Heat Flow Meter Apparatus Practice for In-Situ Measurement of Heat Flux and Temperature on Building Envelope Components Practice for Determining Thermal Resistance of Building Envelope Components from In-Situ Data Test Method for Thermal Performance of Building Materials and Envelope Assemblies by Means of a Hot Box Apparatus
Publisher
Reference
ASME ASME ASME ASTM ASTM
ASME MFC-10M-2000 (RA11) ASME PTC 19.1-2013 ANSI/ASME PTC 36-2004 (RA13) ASTM E96/E96M-14 ASTM E230/E230M-12
ASTM
ASTM F50-12
ISO
ISO 7726:1998
ISO ISO
ISO 8996:2004 ISO 9920:2007
ASHRAE ASHRAE ASHRAE ASME ASME ASME ASME ASME ASME
ASHRAE Guideline 14-2014 ANSI/ASHRAE 41.8-2016 ANSI/ASHRAE 41.9-2011 ASME PTC 19.5-2004 (RA13) ASME MFC-1M-2003 (RA08) ANSI/ASME MFC-2M-1983 (RA13) ASME MFC-3M-2004 ASME MFC-6M-1998 (RA05) ASME MFC-8M-2001 (RA11)
ASME ASME ASME ASME ASTM
ASME MFC-9M-1988 (RA11) ASME MFC-14M-2003 (RA08) ASME MFC-16-2014 ASME MFC-18M-2001 (RA11) ASTM C1616-07 (2012)
ASTM
ASTM D6973-14
ASTM
ASTM D7042-14
ASTM
ASTM D7043-12
ASTM ASTM ASTM ASHRAE ASHRAE ASME ASME ASHRAE ASME ASME ASHRAE ASME ASME SAE ASHRAE ASTM
ASTM D7152-11 ASTM E2139-05 (2011) ASTM G183-15 ANSI/ASHRAE 41.2-1987 (RA92) ANSI/ASHRAE 41.7-2015 ANSI/ASME MFC-4M-1986 (RA08) ANSI/ASME MFC-7M-1987 (RA14) ANSI/ASHRAE 41.3-2014 ASME B40.100-2013 ANSI/ASME PTC 19.2-2010 ANSI/ASHRAE 41.1-2013 ASME B40.200-2008 (RA13) ASME PTC 19.3-1974 (RA04) SAE AS793A-2001 (R2008) ANSI/ASHRAE 125-2016 ASTM C177-13
ASTM
ASTM C518-10
ASTM
ASTM C1046-95 (2013)
ASTM
ASTM C1155-95 (2013)
ASTM
ASTM C1363-11
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.19
Licensed for single user. © 2017 ASHRAE, Inc.
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Title
Publisher
Reference
Mobile Homes and Recreational Vehicles
Residential Load Calculation, 8th ed. Recreational Vehicle Cooking Gas Appliances Oil-Fired Warm Air Heating Appliances for Mobile Housing and Recreational Vehicles Manufactured Homes Recreational Vehicles Gas Supply Connectors for Manufactured Homes Fuel Supply: Manufactured/Mobile Home Parks & Recreational Vehicle Parks Manufactured Housing Construction and Safety Standards Manufactured Housing Recreational Vehicles Plumbing System Components for Recreational Vehicles Low Voltage Lighting Fixtures for Use in Recreational Vehicles Liquid Fuel-Burning Heating Appliances for Manufactured Homes and Recreational Vehicles Gas-Burning Heating Appliances for Manufactured Homes and Recreational Vehicles
ACCA CSA CSA CSA CSA IAPMO IAPMO ICC/ANSI NFPA NFPA NSF UL UL
ANSI/ACCA Manual J-2016 ANSI Z21.57-2010 B140.10-06 (R2011) CAN/CSA-Z240 MH Series-09 (R2014) CAN/CSA-Z240 RV Series-08 (R2013) IAPMO TS 9-2003 Chapter 13, Part II HUD 24 CFR Part 3280 (2008) NFPA 501-2013 NFPA 1192-2015 ANSI/NSF 24-2010 ANSI/UL 234-2005 ANSI/UL 307A-2009
UL
UL 307B-2006
Motors and Generators
Installation and Maintenance of Farm Standby Electric Power Nuclear Power Plant Air-Cleaning Units and Components Testing of Nuclear Air Treatment Systems Fired Steam Generators Gas Turbine Heat Recovery Steam Generators Test Methods for Film-Insulated Magnet Wire Test Method for Evaluation of Engine Oils in a High Speed, Single-Cylinder Diesel Engine—Caterpillar 1R Test Procedure Test Method for Evaluation of Diesel Engine Oils in the T-11 Exhaust Gas Recirculation Diesel Engine Test Methods, Marking Requirements, and Energy Efficiency Levels for Three-Phase Induction Motors Motors and Generators Emergency Electrical Power Supply for Buildings Energy Efficiency Test Methods for Small Motors Standard Test Procedure for Polyphase Induction Motors and Generators Motors and Generators Energy Management Guide for Selection and Use of Fixed Frequency Medium AC Squirrel-Cage Polyphase Industrial Motors Energy Management Guide for Selection and Use of Single-Phase Motors Magnet Wire Motion/Position Control Motors, Controls, and Feedback Devices Rotating Electrical Machines—General Requirements Electric Motors and Generators for Use in Hazardous (Classified) Locations Overheating Protection for Motors
ASABE ASME ASME ASME ASME ASTM ASTM
ANSI/ASABE EP364.3-2006 ASME N509-2002 (RA08) ASME N510-2007 ASME PTC 4-2013 ASME PTC 4.4-2008 (RA13) ASTM D1676-03 (2011) ASTM D6923-14
ASTM
ASTM D7156-13
CSA
CSA C390-10
CSA CSA CSA IEEE NEMA NEMA
C22.2 No. 100-14 CSA C282-09 CAN/CSA C747-09 (R2014) IEEE 112-1996 NEMA MG 1-2014 NEMA MG 10-2013
NEMA NEMA NEMA UL UL UL
NEMA MG 11-1977 (R2012) ANSI/NEMA MW 1000-2014 NEMA ICS 16-2001 UL 1004-1-2012 ANSI/UL 674-2011 ANSI/UL 2111-1997
ACCA ASHRAE ASHRAE ASHRAE/ ACCA ICC
ANSI/ACCA 4 QM-2013 ASHRAE Guideline 4-2008 (RA13) ASHRAE Guideline 32-2012 ANSI/ASHRAE/ACCA 180-2012
Panel Heating Method of Testing for Rating Ceiling Panels for Sensible Heating and Cooling and Cooling
ASHRAE
ANSI/ASHRAE 138-2013
Pipe, Tubing, Scheme for the Identification of Piping Systems and Fittings Pipe Threads, General Purpose (Inch) Wrought Copper and Copper Alloy Braze-Joint Pressure Fittings Power Piping Process Piping Refrigeration Piping and Heat Transfer Components Building Services Piping Practice for Obtaining Hydrostatic or Pressure Design Basis for “Fiberglass” (GlassFiber-Reinforced Thermosetting-Resin) Pipe and Fittings Specification for Welding of Austenitic Stainless Steel Tube and Piping Systems in Sanitary Applications Standards of the Expansion Joint Manufacturers Association, 9th ed. Pipe Hangers and Supports—Materials, Design and Manufacture Pipe Hangers and Supports—Selection and Application
ASME ASME ASME ASME ASME ASME ASME ASTM
ASME A13.1-2007 (RA13) ANSI/ASME B1.20.1-2013 ASME B16.50-2013 ASME B31.1-2014 ASME B31.3-2014 ASME B31.5-2013 ASME B31.9-2014 ASTM D2992-12
AWS
AWS D18.1:2009
EJMA MSS MSS
EJMA ANSI/MSS SP-58-2009 ANSI/MSS SP-69-2003
Operation and Maintenance of Residential HVAC Systems Maintenance Preparation of Operating and Maintenance Documentation for Building Systems Sustainable, High Performance Operations and Maintenance Inspection and Maintenance of Commercial-Building HVAC Systems International Property Maintenance Code® (2015)
IPMC
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40.20
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Licensed for single user. © 2017 ASHRAE, Inc.
Plastic
Metal
Plumbing
Title
Publisher
Reference
General Welding Guidelines (2009) National Fuel Gas Code Refrigeration Tube Fittings—General Specifications Seismic Restraint Manual—Guidelines for Mechanical Systems, 3rd ed. Tube Fittings for Flammable and Combustible Fluids, Refrigeration Service, and Marine Use Specification for Acrylonitrile-Butadiene-Styrene (ABS) Plastic Pipe, Schedules 40 and 80 Specification for Poly (Vinyl Chloride) (PVC) Plastic Pipe, Schedules 40, 80, and 120 Test Method for Obtaining Hydrostatic Design Basis for Thermoplastic Pipe Materials or Pressure Design Basis for Thermoplastic Pipe Products Specification for Perfluoroalkoxy (PFA)-Fluoropolymer Tubing Specification for Polyethylene Stay in Place Form System for End Walls for Drainage Pipe Specification for Chlorinated Poly (Vinyl Chloride) (CPVC) Plastic Pipe, Schedules 40 and 80 Specification for Crosslinked Polyethylene/Aluminum/Crosslinked Polyethylene Tubing OD Controlled SDR9 Test Method for Evaluating the Oxidative Resistance of Polyethylene (PE) Pipe to Chlorinated Water Specification for 12 to 60 in. [300 to 1500 mm] Annular Corrugated Profile-Wall Polyethylene (PE) Pipe and Fittings for Gravity-Flow Storm Sewer and Subsurface Drainage Applications Test Method for Determining Chemical Compatibility of Substances in Contact with Thermoplastic Pipe and Fittings Materials Test Method for Determining Thermoplastic Pipe Wall Stiffness Specification for Steel Reinforced Polyethylene (PE) Corrugated Pipe Electrical Polyvinyl Chloride (PVC) Tubing and Conduit PVC Plastic Utilities Duct for Underground Installation Smooth Wall Coilable Polyethylene Electrical Plastic Duct Fittings for PVC Plastic Utilities Duct for Underground Installation Electrical Nonmetallic Tubing (ENT) Plastics Piping System Components and Related Materials Rubber Gasketed Fittings Welded and Seamless Wrought Steel Pipe Stainless Steel Pipe Specification for Pipe, Steel, Black and Hot-Dipped, Zinc-Coated, Welded and Seamless Specification for Seamless Carbon Steel Pipe for High-Temperature Service Specification for Steel Line Pipe, Black, Furnace-Butt-Welded Specification for Composite Corrugated Steel Pipe for Sewers and Drains Specification for Seamless Copper Pipe, Standard Sizes Specification for Seamless Copper Tube Specification for Seamless Copper Water Tube Specification for Seamless Copper Tube for Air Conditioning and Refrigeration Field Service Specification for Hand-Drawn Copper Capillary Tube for Restrictor Applications Specification for Welded Copper Tube for Air Conditioning and Refrigeration Service Specification for Copper-Beryllium Seamless Tube UNS Nos. C17500 and C17510 Test Method for Rapid Determination of Corrosiveness to Copper from Petroleum Products Using a Disposable Copper Foil Strip Thickness Design of Ductile-Iron Pipe Fittings, Cast Metal Boxes, and Conduit Bodies for Conduit and Cable Assemblies Polyvinyl-Chloride (PVC) Externally Coated Galvanized Rigid Steel Conduit and Intermediate Metal Conduit
NCPWB AGA/NFPA SAE SMACNA UL
NCPWB ANSI Z223.1/NFPA 54-2015 SAE J513-1999 ANSI/SMACNA 001-2008 ANSI/UL 109-1997
ASTM
ASTM D1527-99 (2005)
ASTM ASTM
ASTM D1785-12 ASTM D2837-13e1
ASTM ASTM
ASTM D6867-03 (2014) ASTM D7082-04 (2010)
ASTM
ASTM F441/F441M-13e1
ASTM
ASTM F2262-09
ASTM
ASTM F2263-14
ASTM
ASTM F2306/F2306M-14
ASTM
ASTM F2331-11
ASTM ASTM NEMA NEMA NEMA NEMA NEMA NSF UL ASME ASME ASTM ASTM ASTM ASTM ASTM ASTM ASTM ASTM
ASTM F2433-05 (2013) ASTM F2435-12 NEMA TC 2-2013 NEMA TC 6 and 8-2013 NEMA TC 7-2013 NEMA TC 9-2004 (R2012) NEMA TC 13-2005 ANSI/NSF 14-2014 ANSI/UL 213-2008 ASME B36.10M-2004 (RA10) ASME B36.19M-2004 (RA10) ASTM A53/53M-12 ASTM A106/A106M-14 ASTM A1037/A1037M-05 (2012) ASTM A1042/A1042M-04 (2009) ASTM B42-10 ASTM B75/B75M-11 ASTM B88-14 ASTM B280-13
ASTM ASTM ASTM ASTM
ASTM B360-09 ASTM B640-12a ASTM B937-04 (2010) ASTM D7095-04 (2014)
AWWA NEMA NEMA
ANSI/AWWA C150/A21.50-14 NEMA FB 1-2012 NEMA RN 1-2005 (R2013)
ASME ASME ASME ASME
ASME A112.14.1-2003 (RA12) ASME A112.18.1-2012 ASME A112.18.2-2011 ASME A112.18.3-2002 (RA12)
IAPMO ICC ICC PHCC PHCC
IAPMO IPC IPSDC NSPC 2012 PHCC 2012
Backwater Valves Plumbing Supply Fittings Plumbing Waste Fittings Performance Requirements for Backflow Protection Devices and Systems in Plumbing Fixture Fittings Uniform Plumbing Code (2012) (with IAPMO Installation Standards) International Plumbing Code® (2015) International Private Sewage Disposal Code® (2015) 2012 National Standard Plumbing Code (NSPC) 2012 National Standard Plumbing Code—Illustrated
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.21
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Title
Publisher
Standard Plumbing Code (1997)
SBCCI
SPC
Balancing and Testing Air and Hydronic Systems Centrifugal Pumps Specification for Horizontal End Suction Centrifugal Pumps for Chemical Process Specification for Vertical-in-Line Centrifugal Pumps for Chemical Process Specification for Sealless Horizontal End Suction Metallic Centrifugal Pumps for Chemical Process Specification for Thermoplastic and Thermoset Polymer Material Horizontal End Suction Centrifugal Pumps for Chemical Process Liquid Pumps Energy Efficiency Test Methods for Small Pumps Performance Standard for Liquid Ring Vacuum Pumps, 4th ed. (2011) Rotodynamic (Centrifugal) Pumps for Nomenclature and Definitions Rotodynamic (Centrifugal) Pumps for Design and Application Rotodynamic (Centrifugal) Pumps for Manuals Describing Installation, Operation and Maintenance Rotodynamic (Vertical) Nomenclature Rotodynamic (Vertical) Application Rotodynamic (Vertical) Operations Rotary Pumps for Nomenclature, Definitions, Application, and Operation Sealless, Magnetically Driven Rotary Pumps for Nomenclature, Definitions, Application, Operation, and Test Sealless Rotodynamic Pumps for Nomenclature, Definitions, Application, Operation, and Test Reciprocating Pumps for Nomenclature, Definitions, Application, and Operation Direct Acting (Steam) Pumps for Nomenclature, Definitions, Application, and Operation Pumps—General Guidelines for Types, Definitions, Application, Sound Measurement and Decontamination Rotodynamic Pumps for Assessment of Allowable Nozzle Loads Centrifugal and Vertical Pumps for Allowable Operating Region Rotodynamic Pumps for Vibration Measurements and Allowable Values Rotodynamic (Centrifugal and Vertical) Pumps Guideline for Condition Monitoring Intake Design for Rotodynamic Pumps Engineering Data Book, 2nd ed. Equipment for Swimming Pools, Spas, Hot Tubs and Other Recreational Water Facilities Pumps for Oil-Burning Appliances Motor-Operated Water Pumps Swimming Pool Pumps, Filters, and Chlorinators
ACCA ASME ASME ASME ASME
ACCA Manual B-2009 ASME PTC 8.2-1990 ASME B73.1-2012 ASME B73.2-2003 (RA08) ASME B73.3-2003 (RA08)
ASME
ASME B73.5M-1995 (RA07)
CSA CSA HEI HI HI HI
CAN/CSA-C22.2 No. 108-14 CAN/CSA C820-02 (R2012) HEI 2854 ANSI/HI 1.1-1.2-2014 ANSI/HI 1.3-2013 ANSI/HI 1.4-2014
HI HI HI HI HI
ANSI/HI 2.1-2.2-2014 ANSI/HI 2.3-2013 ANSI/HI 2.4-2014 ANSI/HI 3.1-3.5-2008 ANSI/HI 4.1-4.6-2010
HI
ANSI/HI 5.1-5.6-2010
HI HI HI
ANSI/HI 6.1-6.5-2000 ANSI/HI 8.1-8.5-2000 ANSI/HI 9.1-9.5-2000
HI HI HI HI HI HI NSF
ANSI/HI 9.6.2-2011 ANSI/HI 9.6.3-2012 ANSI/HI 9.6.4-2009 ANSI/HI 9.6.5-2009 ANSI/HI 9.8-2012 HI (1990) ANSI/NSF 50-2013
UL UL UL
UL 343-2008 ANSI/UL 778-2010 ANSI/UL 1081-2008
Radiators
Balancing and Testing Air and Hydronic Systems Testing and Rating Standard for Baseboard Radiation, 8th ed. (2005) Testing and Rating Standard for Finned Tube (Commercial) Radiation, 6th ed. (2005)
ACCA HYDI HYDI
ACCA Manual B-2009 IBR IBR
Receivers
Refrigerant Liquid Receivers Refrigerant-Containing Components and Accessories, Nonelectrical
AHRI UL
ANSI/AHRI 495-2005 ANSI/UL 207-2009
Refrigerants
Threshold Limit Values for Chemical Substances (updated annually) Specifications for Refrigerants Refrigerant Recovery/Recycling Equipment Refrigerant Information Recommended for Product Development and Standards Method of Testing Flow Capacity of Refrigerant Capillary Tubes Designation and Safety Classification of Refrigerants Sealed Glass Tube Method to Test the Chemical Stability of Materials for Use Within Refrigerant Systems Refrigeration Oil Description Reducing the Release of Halogenated Refrigerants from Refrigerating and AirConditioning Equipment and Systems Methods of Testing Capacity for Refrigerant Pressure Regulators Method of Test to Determine the Performance of Halocarbon Refrigerant Leak Detectors Test Method for Acid Number of Petroleum Products by Potentiometric Titration Test Method for Concentration Limits of Flammability of Chemical (Vapors and Gases) Refrigerant-Containing Components for Use in Electrical Equipment Refrigerants—Designation System Procedure Retrofitting CFC-12 (R-12) Mobile Air-Conditioning Systems to HFC-134a (R-134a)
ACGIH AHRI AHRI ASHRAE ASHRAE ASHRAE ASHRAE
ACGIH AHRI 700-2014 AHRI 740-1998 ASHRAE Guideline 6-2015 ANSI/ASHRAE 28-1996 (RA10) ANSI/ASHRAE 34-2016 ANSI/ASHRAE 97-2007
ASHRAE ASHRAE
ANSI/ASHRAE 99-2006 ANSI/ASHRAE 147-2013
ASHRAE ASHRAE
ANSI/ASHRAE 158.2-2011 ANSI/ASHRAE 173-2012
ASTM ASTM CSA ISO SAE
ASTM D664-11a ASTM E681-09 C22.2 No. 140.3-09 (R2014) ISO 817:2014 SAE J1661-2011
Licensed for single user. © 2017 ASHRAE, Inc.
Pumps
Reference
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40.22
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
Recommended Service Procedure for the Containment of CFC-12 (R-12) Standard of Purity for Recycled R-134a (HFC-134a) and R-1234yf (HFO-1234yf) for Use in Mobile Air-Conditioning Systems HFC-134a (R-134a) Service Hose Fittings for Automotive Air-Conditioning Service Equipment CFC-12 (R-12) Refrigerant Recovery Equipment for Mobile Automotive AirConditioning Systems Recommended Service Procedure for the Containment of HFC-134a Refrigerant-Containing Components and Accessories, Nonelectrical Refrigerant Recovery/Recycling Equipment Refrigerants
SAE SAE
SAE J1989-2011 SAE J2099-2012
SAE
SAE J2197-2011
SAE
SAE J2209-2011
SAE UL UL UL
SAE J2211-2011 ANSI/UL 207-2009 ANSI/UL 1963-2011 ANSI/UL 2182-2006
Refrigeration
Quality Maintenance of Commercial Refrigeration Systems Safety Standard for Refrigeration Systems Mechanical Refrigeration Code Refrigeration Equipment Equipment, Design and Installation of Ammonia Mechanical Refrigerating Systems Refrigerated Medical Equipment
ACCA ASHRAE CSA CSA IIAR UL
ANSI/ACCA 14 QMref-2015 ANSI/ASHRAE 15-2016 B52-13 CAN/CSA-C22.2 No. 120-13 ANSI/IIAR 2-2008 ANSI/UL 416-1993
Refrigeration Systems
Ejectors Safety Standard for Refrigeration Systems Designation and Safety Classification of Refrigerants Reducing the Release of Halogenated Refrigerants from Refrigerating and AirConditioning Equipment and Systems Testing of Refrigerating Systems Standards for Steam Jet Vacuum Systems, 7th ed. (2012) Mechanical Transport Refrigeration Units Mechanical Refrigeration and Air-Conditioning Installations Aboard Ship General Requirements for Application of Vapor Cycle Refrigeration Systems for Aircraft Safety Standard for Motor Vehicle Refrigerant Vapor Compression Systems
ASME ASHRAE ASHRAE ASHRAE
ASME PTC 24-1976 (RA82) ANSI/ASHRAE 15-2016 ANSI/ASHRAE 34-2016 ANSI/ASHRAE 147-2013
ISO HEI AHRI ASHRAE SAE SAE
ISO 916-1968 HEI 2866-1 ANSI/AHRI 1111-2013 ANSI/ASHRAE 26-2010 SAE ARP731C-2003 (R2010) SAE J639-2011
Quality Maintenance of Commercial Refrigeration Systems Method of Testing Commercial Refrigerators and Freezers Energy Performance Standard for Commercial Refrigerated Display Cabinets and Merchandise Energy Performance of Food Service Refrigerators and Freezers Gas Food Service Equipment
ACCA ASHRAE CSA
ANSI/ACCA 14 QMref-2015 ANSI/ASHRAE 72-2014 CAN/CSA-C657-12
CSA CSA
Food Equipment Commercial Refrigerators and Freezers Mobile Food Carts Refrigeration Unit Coolers Refrigerating Units Commercial Refrigerators and Freezers Refrigerators, Refrigerator-Freezers and Freezers Refrigerators Using Gas Fuel Energy Performance and Capacity of Household Refrigerators, Refrigerator-Freezers, Freezers, and Wine Chillers Household Refrigerators and Freezers
NSF NSF NSF UL UL UL AHAM CSA CSA
C827-10 ANSI Z83.11-2006/CSA 1.8A-2006 (R2011) ANSI/NSF 2-2014 ANSI/NSF 7-2014 ANSI/NSF 59-2012 ANSI/UL 412-2011 ANSI/UL 427-2011 ANSI/UL 471-2010 ANSI/AHAM HRF-1-2008 ANSI Z21.19-2014/CSA1.4-2014 CAN/CSA C300-12
UL/CSA
ANSI/UL 250-1993/C22.2 No. 63-93 (R2013)
Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Home Evaluation and Performance Improvement Technician’s Guide & Workbook for Duct Diagnostics and Repair Bob’s House Good HVAC Practices for Residential and Commercial Buildings Building Systems Analysis and Retrofit Manual, 2nd ed. Procedure for Retrofitting CFC-12 (R-12) Mobile Air Conditioning Systems to HFC134a (R-134a)
ACCA ACCA ACCA ACCA ACCA SMACNA SAE
ANSI/ACCA 12 QH-2014 ACCA 2015 ACCA 2016 ACCA 2008 ACCA 2003 SMACNA 2011 SAE J1661-2011
Commercial Low Pressure, Low Velocity Duct System Design Power Ventilators
ACCA UL
ACCA Manual Q-1990 ANSI/UL 705-2004
Transport
Refrigerators Commercial
Household
Retrofitting Building
Refrigerant Roof Ventilators Safety
Guideline for Risk Management of Public Health and Safety in Buildings
ASHRAE
ASHRAE Guideline 29-2009
Seismic Restraint
Method of Test of Seismic Restraint Devices for HVAC&R Equipment Seismic Restraint Manual—Guidelines for Mechanical Systems, 3rd ed.
ASHRAE SMACNA
ANSI/ASHRAE 171-2008 ANSI/SMACNA 001-2008
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.23
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Title
Publisher
Reference
Smoke Control Systems
The Commissioning Process for Smoke Management Systems Laboratory Methods of Testing Fans Used to Exhaust Smoke in Smoke Management Systems Smoke-Control Systems Utilizing Barriers and Pressure Differences Smoke Management Systems in Malls, Atria, and Large Spaces Ceiling Dampers Smoke Dampers
ASHRAE ASHRAE
ASHRAE Guideline 1.5-2012 ANSI/ASHRAE 149-2013
NFPA NFPA UL UL
NFPA 92A-2009 NFPA 92B-2009 ANSI/UL 555C-2014 ANSI/UL 555S-2014
Thermal Performance Testing of Solar Ambient Air Heaters Testing and Reporting Solar Cooker Performance Method of Measuring Solar-Optical Properties of Materials Methods of Testing to Determine the Thermal Performance of Solar Collectors Methods of Testing to Determine the Thermal Performance of Solar Domestic Water Heating Systems Methods of Testing to Determine the Thermal Performance of Unglazed Flat-Plate Liquid-Type Solar Collectors Practice for Installation and Service of Solar Space Heating Systems for One and Two Family Dwellings Practice for Evaluating Thermal Insulation Materials for Use in Solar Collectors Practice for Installation and Service of Solar Domestic Water Heating Systems for One and Two Family Dwellings Reference Solar Spectral Irradiance at the Ground at Different Receiving Conditions— Part 1: Direct Normal and Hemispherical Solar Irradiance for Air Mass 1.5 Solar Collectors Packaged Solar Domestic Hot Water Systems (Liquid to Liquid Heat Transfer) Installation Code for Solar Domestic Hot Water Systems Solar Heating—Domestic Water Heating Systems—Part 2: Outdoor Test Methods for System Performance Characterization and Yearly Performance Prediction of Solar-Only Systems Solar Energy—Solar Thermal Collectors—Test Methods Solar Water Heaters—Elastomeric Materials for Absorbers, Connecting Pipes and Fittings—Method of Assessment Solar Energy—Calibration of a Pyranometer Using a Pyrheliometer
ASABE ASABE ASHRAE ASHRAE ASHRAE
ANSI/ASAE S423-1991 (R2012) ASAE S580.1-2013 ASHRAE 74-1988 ANSI/ASHRAE 93-2010 (RA14) ANSI/ASHRAE 95-1981 (RA87)
ASHRAE
ANSI/ASHRAE 96-1980 (RA89)
ASTM
ASTM E683-91 (2013)
ASTM ASTM
ASTM E861-13 ASTM E1056-13
ISO
ISO 9845:1992
CSA CSA CSA ISO
CAN/CSA F378 Series-11 CAN/CSA F379 Series-09 (R2013) CAN/CSA F383-08 (R2013) ISO 9459-2:1995
ISO ISO
ISO 9806:2013 ISO 9808:1990
Licensed for single user. © 2017 ASHRAE, Inc.
Solar Equipment
ISO
ISO 9846:1993
Solenoid Valves Solenoid Valves for Use with Volatile Refrigerants Methods of Testing Capacity of Refrigerant Solenoid Valves Electrically Operated Valves
AHRI ASHRAE UL
ANSI/AHRI 760-2014 ANSI/ASHRAE 158.1-2012 UL 429-2013
Sound Threshold Limit Values for Physical Agents (updated annually) Measurement Specification for Sound Level Meters Specification for Octave-Band and Fractional-Octave-Band Analog and Digital Filters Microphones, Part 1: Specifications for Laboratory Standard Microphones Part 2: Primary Method for Pressure Calibration of Laboratory Standard Microphones by the Reciprocity Technique Specification for Acoustical Calibrators Measurement of Industrial Sound Test Method for Measuring Acoustical and Airflow Performance of Duct Liner Materials and Prefabricated Silencers Test Method for Determination of Decay Rates for Use in Sound Insulation Test Methods Procedural Standards for the Measurement of Sound and Vibration, 2nd ed. (2006) HVAC Systems Sound and Vibration Procedural Guide, 1st ed. Fans Reverberant Room Method for Sound Testing of Fans Methods for Calculating Fan Sound Ratings from Laboratory Test Data Application of Sone Ratings for Non-Ducted Air Moving Devices Application of Sound Power Level Ratings for Fans Acoustics—Measurement of Airborne Noise Emitted and Structure-Borne Vibration Induced by Small Air-Moving Devices—Part 1: Airborne Noise Measurement Part 2: Structure-Borne Vibration
ACGIH ASA ASA ASA ASA
ACGIH ANSI S1.4-2014 ANSI S1.11-2014 ANSI S1.15-1997/Part 1 (R2011) ANSI S1.15-2005/Part 2 (R2010)
ASA ASME ASTM
ANSI S1.40-2006 (R2011) ASME PTC 36-2004 (RA13) ASTM E477-13
ASTM
ASTM E2235-04 (2012)
NEBB SMACNA AMCA AMCA AMCA AMCA ASA
NEBB SMACNA 2013 AMCA 300-08 AMCA 301-90 AMCA 302-73 (R2008) AMCA 303-79 (R2008) ANSI/ASA S12.11-1-2013/ISO 10302-1:2013 ANSI/ASA S12.11-2-2013/ISO 10302-2:2013 ANSI/AHRI 270-2008 ANSI/AHRI 275-2010 ANSI/AHRI 300-2008 ANSI/AHRI 350-2008 ANSI/AHRI 370-2011
Other Sound Rating of Outdoor Unitary Equipment Equipment Application of Sound Rating Levels of Outdoor Unitary Equipment Sound Rating and Sound Transmission Loss of Packaged Terminal Equipment Sound Rating of Non-Ducted Indoor Air-Conditioning Equipment Sound Rating of Large Air-Cooled Outdoor Refrigerating and Air-Conditioning Equipment
ASA AHRI AHRI AHRI AHRI AHRI
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.24
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject
Licensed for single user. © 2017 ASHRAE, Inc.
Techniques
Title
Publisher
Reference
Method of Rating Sound and Vibration of Refrigerant Compressors Method of Measuring Machinery Sound Within an Equipment Space Statistical Methods for Determining and Verifying Stated Noise Emission Values of Machinery and Equipment Sound Level Prediction for Installed Rotating Electrical Machines Preferred Frequencies, Frequency Levels, and Band Numbers for Acoustical Measurements Reference Quantities for Acoustical Levels Measurement of Sound Pressure Levels in Air Procedure for the Computation of Loudness of Steady Sound Criteria for Evaluating Room Noise Methods for Determining the Insertion Loss of Outdoor Noise Barriers Engineering Method for the Determination of Sound Power Levels of Noise Sources Using Sound Intensity Procedures for Outdoor Measurement of Sound Pressure Level Methods for Measurement of Sound Emitted by Machinery and Equipment at Workstations and Other Specified Positions Methods for Calculation of Sound Emitted by Machinery and Equipment at Workstations and Other Specified Positions from Sound Power Level Acoustics—Determination of Sound Power Levels and Sound Energy Levels of Noise Sources Using Sound Pressure—Precision Method for Reverberation Test Rooms Acoustics—Determination of Sound Power Levels of Noise Sources Using Sound Pressure—Engineering Methods for Small, Movable Sources in Reverberant Fields— Part 1: Comparison Method for Hard-Walled Test Rooms Part 2: Methods for Special Reverberation Test Rooms
AHRI AHRI ASA
ANSI/AHRI 530-2011 ANSI/AHRI 575-2008 ANSI S12.3-1985 (R2011)
NEMA ASA ASA ASA ASA ASA ASA ASA
NEMA MG 3-1974 (R2012) ANSI/ASA S1.6-1984 (R2011) ANSI/ASA S1.8-1989 (R2011) ANSI/ASA S1.13-2005 (R2010) ANSI/ASA S3.4-2007 (R2012) ANSI/ASA S12.2-2008 ANSI/ASA S12.8-1998 (R2013) ANSI/ASA S12.12-1992 (R2012)
ASA ASA
ANSI/ASA S12.18-1994 (R2009) ANSI/ASA S12.43-1997 (R2012)
ASA
ANSI/ASA S12.44-1997 (R2012)
ASA
ANSI/ASA S12.51-2012/ISO 3741:2012 ANSI/ASA S12.53/Part 1-2011/ISO 3743-1:2011
Acoustics—Determination of Sound Power Levels and Sound Energy Levels of Noise Sources Using Sound Pressure—Engineering Methods for an Essentially Free Field over a Reflecting Plane Acoustics—Determination of Sound Power Levels and Sound Energy Levels of Noise Sources Using Sound Pressure—Survey Method Using an Enveloping Measurement Surface over a Reflecting Plane Test Method for Impedance and Absorption of Acoustical Materials by the Impedance Tube Method Test Method for Sound Absorption and Sound Absorption Coefficients by the Reverberation Room Method Test Method for Measurement of Airborne Sound Attenuation Between Rooms in Buildings Test Method for Impedance and Absorption of Acoustical Materials Using a Tube, Two Microphones and a Digital Frequency Analysis System Test Method for Evaluating Masking Sound in Open Offices Using A-Weighted and One-Third Octave Band Sound Pressure Levels Test Method for Measurement of Sound in Residential Spaces Acoustics—Method for Calculating Loudness Level Acoustics—Determination of Sound Power Levels of Noise Sources Using Sound Intensity; Part 1: Measurement at Discrete Points Part 2: Measurement by Scanning Procedural Standards for the Measurement of Sound and Vibration, 2nd ed. (2006) Terminology Acoustical Terminology Terminology Relating to Building and Environmental Acoustics Space Heaters Methods of Testing for Rating Combination Space-Heating and Water-Heating Appliances Gas-Fired Room Heaters, Vol. II, Unvented Room Heaters Vented Gas-Fired Space Heating Appliances
Sustainability
Symbols
ASA ASA ASA
ANSI/ASA S12.53/Part 2-1999 (R2015)/ISO 3743-2:1999 (R2015) ANSI/ASA S12.54-2011/ISO 3744:2011
ASA
ANSI/ASA S12.56-2011/ISO 3746:2011
ASTM
ASTM C384-04 (2011)
ASTM
ASTM C423-09a
ASTM
ASTM E336-14
ASTM
ASTM E1050-12
ASTM
ASTM E1573-09
ASTM ISO ISO
ASTM E1574-98 (2014) ISO 532:1975 ISO 9614-1:1993
ISO NEBB ASA ASTM
ISO 9614-2:1996 NEBB ANSI S1.1-1994 (R2004) ASTM C634-13
ASHRAE CSA CSA
ANSI/ASHRAE 124-2007 ANSI Z21.11.2-2013 ANSI Z21.86-2008/CSA 2.32-2008 (R2014) UL 1278-2014 UL 2021-2013
Movable and Wall- or Ceiling-Hung Electric Room Heaters Fixed and Location-Dedicated Electric Room Heaters
UL UL
Sustainable, High Performance Operations and Maintenance Standard for the Design of High-Performance, Green Buildings Except Low-Rise Residential Buildings International Green Construction Code™ (2012)
ASHRAE ASHRAE/ USGBC ICC
ASHRAE Guideline 32-2012
Graphic Symbols for Heating, Ventilating, Air-Conditioning, and Refrigerating Systems Graphical Symbols for Plumbing Fixtures for Diagrams Used in Architecture and Building Construction Symbols for Mechanical and Acoustical Elements as Used in Schematic Diagrams
ASHRAE ASME
ANSI/ASHRAE 134-2005 (RA14) ANSI/ASME Y32.4-1977 (RA04)
ASME
ANSI/ASME Y32.18-1972 (RA13)
ANSI/ASHRAE/USGBC/IES 189.12014 IGCC
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.25
Selected Codes and Standards Published by Various Societies and Associations (Continued)
Licensed for single user. © 2017 ASHRAE, Inc.
Subject
Title
Publisher
Reference
Practice for Mechanical Symbols, Shipboard Heating, Ventilation, and Air Conditioning (HVAC) Standard Symbols for Welding, Brazing, and Nondestructive Examination Graphic Symbols for Electrical and Electronics Diagrams Standard for Logic Circuit Diagrams American National Standard for Metric Practice Abbreviations and Acronyms for Use on Drawings and Related Documents Engineering Drawing Practices American National Standard for Safety Colors
ASTM
ASTM F856-97 (2014)
AWS IEEE IEEE IEEE/ASTM ASME ASME NEMA
AWS A2.4:2012 ANSI/CSA/IEEE 315-1975 (R1993) IEEE 991-1986 (R1994) IEEE/ASTM-SI10-10 ASME Y14.38-2007 (RA13) ASME Y14.100-2013 ANSI/NEMA Z535.1-2006 (R2011)
Terminals, Wiring
Electrical Quick-Connect Terminals Wire Connectors Splicing Wire Connectors Equipment Wiring Terminals for Use with Aluminum and/or Copper Conductors
UL UL UL UL
ANSI/UL 310-2014 ANSI/UL 486A-486B-2013 ANSI/UL 486C-2013 ANSI/UL 486E-2009
Testing and Balancing
AABC National Standards for Total System Balance (2002) Balancing and Testing Air and Hydronic Systems Industrial Process/Power Generation Fans: Site Performance Test Standard Guidelines for Measuring and Reporting Environmental Parameters for Plant Experiments in Growth Chambers HVAC&R Technical Requirements for the Commissioning Process Measurement, Testing, Adjusting, and Balancing of Building HVAC Systems Rotary Pump Tests Air-Operated Pump Tests Pumps—General Guidelines for Types, Definitions, Application, Sound Measurement and Decontamination Rotodynamic Submersible Pump Tests Procedural Standards for Certified Testing of Cleanrooms, 3rd ed. (2009) Procedural Standards for TAB Environmental Systems, 8th ed. (2015) HVAC Systems Testing, Adjusting and Balancing, 3rd ed.
AABC ACCA AMCA ASABE
AABC ACCA Manual B-2009 AMCA 803-02 (R2008) ANSI/ASAE EP411.5-2012
ASHRAE ASHRAE HI HI HI
ASHRAE Guideline 1.1-2007 ANSI/ASHRAE 111-2008 ANSI/HI 3.6-2010 ANSI/HI 10.6-2010 HI 9.1-9.5-2000
HI NEBB NEBB SMACNA
ANSI/HI 11.6-2012 NEBB NEBB SMACNA 2002
Thermal Energy Storage: A Guide for Commercial HVAC Contractors Method of Testing Thermal Storage Devices with Electrical Input and Thermal Output Based on Thermal Performance Measurement, Testing, Adjusting, and Balancing of Building HVAC Systems Method of Testing the Performance of Cool Storage Systems
ACCA ASHRAE
ACCA 2005 ANSI/ASHRAE 94.2-2010
ASHRAE ASHRAE
ANSI/ASHRAE 111-2008 ANSI/ASHRAE 150-2000 (RA14)
Transformers
Minimum Efficiency Values for Liquid-Filled Distribution Transformers Minimum Efficiency Values for Dry-Type Transformers Maximum Losses For Power Transformers Guide for Determining Energy Efficiency of Distribution Transformers
CSA CSA CSA NEMA
CAN/CSA-C802.1-13 CAN/CSA-C802.2-12 CAN/CSA-C802.3-01 (R2012) NEMA TP-1-2002
Turbines
Steam Turbines Steam Turbines in Combined Cycle Hydraulic Turbines and Pump-Turbines Gas Turbines Wind Turbines Specification for Stainless Steel Bars for Compressor and Turbine Airfoils Specification for Gas Turbine Fuel Oils Steam Turbines for Mechanical Drive Service Land Based Steam Turbine Generator Sets, 0 to 33 000 kW
ASME ASME ASME ASME ASME ASTM ASTM NEMA NEMA
ASME PTC 6-2004 (RA14) ASME PTC 6.2-2011 ASME PTC 18-2011 ASME PTC 22-2014 ASME PTC 42-1988 (RA04) ASTM A1028-03 (2009) ASTM D2880-14a NEMA SM 23-1991 (R2002) NEMA SM 24-1991 (R2002)
Valves
Face-to-Face and End-to-End Dimensions of Valves Valves—Flanged, Threaded, and Welding End Manually Operated Metallic Gas Valves for Use in Aboveground Piping Systems up to 5 psi Pressure Relief Devices Methods of Testing Capacity of Refrigerant Solenoid Valves Relief Valves for Hot Water Supply
ASME ASME ASME ASME ASHRAE CSA
Control Valve Capacity Test Procedures Metal Valves for Use in Flanged Pipe Systems—Face-to-Face and Centre-to-Face Dimensions Safety Valves for Protection Against Excessive Pressure, Part 1: Safety Valves Oxygen System Fill/Check Valve Flow Control Valves for Anhydrous Ammonia and LP-Gas Safety Relief Valves for Anhydrous Ammonia and LP-Gas LP-Gas Regulators Electrically Operated Valves
ISA ISO ISO SAE UL UL UL UL
ASME B16.10-2009 ASME B16.34-2013 ASME B16.44-2012 ASME PTC 25-2014 ANSI/ASHRAE 158.1-2012 ANSI Z21.22-1999/CSA 4.4-1999 (R2014) ANSI/ISA-S75.02.01-2008 ISO 5752:1982 ISO 4126-1:2013 SAE AS1225A-1997 (R2012) ANSI/UL 125-2014 ANSI/UL 132-2007 ANSI/UL 144-2012 UL 429-2013
Thermal Storage
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40.26
2017 ASHRAE Handbook—Fundamentals (SI) Selected Codes and Standards Published by Various Societies and Associations (Continued)
Subject Gas
Publisher
Reference
Valves for Flammable Fluids Manually Operated Metallic Gas Valves for Use in Gas Piping Systems up to 125 psig (Sizes NPS 1/2 through 2) Large Metallic Valves for Gas Distribution (Manually Operated, NPS-2 1/2 to 12, 125 psig Maximum) Manually Operated Thermoplastic Gas Shutoffs and Valves in Gas Distribution Systems Manually Operated Gas Valves for Appliances, Appliance Connection Valves, and Hose End Valves Automatic Valves for Gas Appliances Combination Gas Controls for Gas Appliances Convenience Gas Outlets and Optional Enclosures
UL ASME
ANSI/UL 842-2007 ASME B16.33-2012
ASME
ANSI/ASME B16.38-2012
ASME CSA
Thermostatic Refrigerant Expansion Valves Solenoid Valves for Use with Volatile Refrigerants Refrigerant Pressure Regulating Valves Method of Testing Thermostatic Refrigerant Expansion Valves Methods of Testing Capacity of Refrigerant Solenoid Valves
AHRI AHRI AHRI ASHRAE ASHRAE
ASME B16.40-2013 ANSI Z21.15-2009/CGA 9.1-2009 (R2014) ANSI Z21.21-2012/CGA 6.5-2012 ANSI Z21.78-2010/CGA 6.20-2010 ANSI Z21.90-2001/CSA 6.24-2001 (R2011) ANSI/AHRI 750-2007 ANSI/AHRI 760-2014 ANSI/AHRI 770-2014 ANSI/ASHRAE 17-2008 ANSI/ASHRAE 158.1-2012
Vapor Retarders
Practice for Selection of Water Vapor Retarders for Thermal Insulation Practice for Determining the Properties of Jacketing Materials for Thermal Insulation Specification for Flexible, Low Permeance Vapor Retarders for Thermal Insulation
ASTM ASTM ASTM
ASTM C755-10 ASTM C921-10 ASTM C1136-12
Vending Machines
Methods of Testing for Rating Vending Machines for Sealed Beverages Methods of Testing for Rating Pre-Mix and Post-Mix Beverage Dispensing Equipment Vending Machines Energy Performance of Vending Machines Vending Machines for Food and Beverages Refrigerated Vending Machines Vending Machines
ASHRAE ASHRAE CSA CSA NSF UL UL
ANSI/ASHRAE 32.1-2010 ANSI/ASHRAE 32.2-2003 (RA11) C22.2 No. 128-95 (R2013) CAN/CSA C804-09 (R2014) ANSI/NSF 25-2012 ANSI/UL 541-2011 ANSI/UL 751-2012
CSA
ANSI Z21.66-1996/CSA 6.14-M96 (R2011) ANSI/UL 17-2008
Refrigerant
Licensed for single user. © 2017 ASHRAE, Inc.
Title
Vent Dampers Automatic Vent Damper Devices for Use with Gas-Fired Appliances
Ventilation
CSA CSA CSA
Vent or Chimney Connector Dampers for Oil-Fired Appliances
UL
Commercial Systems Overview Residential Equipment Selection
ACCA
Residential Duct Systems Commercial Low Pressure, Low Velocity Duct System Design Comfort, Air Quality, and Efficiency by Design Guide for Testing Ventilation Systems (1991) Industrial Ventilation: A Manual of Recommended Practice, 29th ed. (2016) Design of Ventilation Systems for Poultry and Livestock Shelters Design Values for Emergency Ventilation and Care of Livestock and Poultry Heating, Ventilating and Cooling Greenhouses Guidelines for Selection of Energy Efficient Agricultural Ventilation Fans Uniform Terminology for Livestock Production Facilities Agricultural Ventilation Constant Speed Fan Test Standard Guide for the Ventilation and Thermal Management of Batteries for Stationary Applications Ventilation for Acceptable Indoor Air Quality Ventilation and Acceptable Indoor Air Quality in Low-Rise Residential Buildings Method of Testing for Room Air Diffusion Measuring Air Change Effectiveness Ventilation for Commercial Cooking Operations Ventilation of Health Care Facilities Residential Mechanical Ventilation Systems Parking Structures Installation of Air-Conditioning and Ventilating Systems Ventilation Control and Fire Protection of Commercial Cooking Operations Food Equipment Biosafety Cabinetry: Design, Construction, Performance, and Field Certification Aerothermodynamic Systems Engineering and Design Heater, Airplane, Engine Exhaust Gas to Air Heat Exchanger Type Test Procedure for Battery Flame Retardant Venting Systems
ACCA ACCA ACCA ACCA ACGIH ACGIH ASABE ASABE ASABE ASABE ASABE ASABE ASHRAE/ IEEE ASHRAE ASHRAE ASHRAE ASHRAE ASHRAE ASHRAE/ ASHE CSA NFPA NFPA NFPA NSF NSF SAE SAE SAE
ACCA Manual CS-1993 ANSI/ACCA 3 Manual S-2014 ANSI/ACCA 1 Manual D-2016 ACCA Manual Q-1990 ACCA Manual RS-1997 ACGIH ACGIH ASAE EP270.5-1986 (R2012) ANSI/ASAE EP282.2-1993 (R2013) ANSI/ASAE EP406.4-2003 (R2008) ASAE EP566.2-2012 ASAE S501-1990 (R2011) ASABE S565-2005 (R2011) ASHRAE Guideline 21-2012/IEEE 1635-2012 ANSI/ASHRAE 62.1-2016 ANSI/ASHRAE 62.2-2016 ANSI/ASHRAE 113-2013 ANSI/ASHRAE 129-1997 (RA02) ANSI/ASHRAE 154-2016 ANSI/ASHRAE/ASHE 170-2013 CAN/CSA F326-M91 (R2010) NFPA 88A-2015 NFPA 90A-2015 NFPA 96-2014 ANSI/NSF 2-2014 ANSI/NSF 49-2012 SAE AIR1168/3-1989 (R2011) SAE ARP86-2011 SAE J1495-2013
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
Codes and Standards
40.27
Selected Codes and Standards Published by Various Societies and Associations (Continued) Subject
Commercial Systems Overview Draft Hoods National Fuel Gas Code Explosion Prevention Systems Smoke and Heat Venting Chimneys, Fireplaces, Vents and Solid Fuel-Burning Appliances Guide for Free Standing Steel Stack Construction, 2nd ed. Guyed Steel Stacks Draft Equipment Gas Vents Type L Low-Temperature Venting Systems Vibration Balance Quality and Vibration Levels for Fans Techniques of Machinery Vibration Measurement Mechanical Vibration and Shock—Resilient Mounting Systems—Part 1: Technical Information to Be Exchanged for the Application of Isolation Systems Evaluation of Human Exposure to Whole-Body Vibration—Part 2: Vibration in Buildings (1 Hz to 80 Hz) Guidelines for the Evaluation of the Response of Occupants of Fixed Structures, Especially Buildings and Off-Shore Structures, to Low-Frequency Horizontal Motion (0.063 to 1 Hz) Procedural Standards for the Measurement of Sound and Vibration, 2nd ed. (2006) HVAC Systems Sound and Vibration Procedural Guide, 1st ed. Water Heaters Desuperheater/Water Heaters Safety for Electrically Heated Livestock Waterers Methods of Testing to Determine the Thermal Performance of Solar Domestic Water Heating Systems Method of Testing for Rating Commercial Gas, Electric, and Oil Service Water Heating Equipment Method of Testing for Rating Residential Water Heaters Methods of Testing for Rating Combination Space-Heating and Water-Heating Appliances Methods of Testing for Efficiency of Space-Conditioning/Water-Heating Appliances That Include a Desuperheater Water Heater Method of Testing Absorption Water-Chilling and Water-Heating Packages Legionellosis: Risk Management for Building Water Systems Boiler and Pressure Vessel Code—Section IV: Heating Boilers Section VI: Recommended Rules for the Care and Operation of Heating Boilers Gas Water Heaters—Vol. I: Storage Water Heaters with Input Ratings of 75,000 Btu per Hour or Less Vol. III: Storage Water Heaters with Input Ratings Above 75,000 Btu per Hour, Circulating and Instantaneous Oil Burning Stoves and Water Heaters Oil-Fired Service Water Heaters and Swimming Pool Heaters Construction and Test of Electric Storage-Tank Water Heaters Performance of Electric Storage Tank Water Heaters for Domestic Hot Water Service Energy Efficiency of Electric Storage Tank Water Heaters and Heat Pump Water Heaters Water Heaters, Hot Water Supply Boilers, and Heat Recovery Equipment Household Electric Storage Tank Water Heaters Oil-Fired Storage Tank Water Heaters Commercial-Industrial Gas Heating Equipment Electric Booster and Commercial Storage Tank Water Heaters Welding and Boiler and Pressure Vessel Code—Section IX: Welding and Brazing Qualifications Brazing Structural Welding Code—Steel Specification for Welding of Austenitic Stainless Steel Tube and Piping Systems in Sanitary Applications Wood-Burning Threshold Limit Values for Chemical Substances (updated annually) Appliances Specification for Room Heaters, Pellet Fuel Burning Type Guide for Construction of Solid Fuel Burning Masonry Heaters Installation Code for Solid-Fuel-Burning Appliances and Equipment Solid-Fuel-Fired Central Heating Appliances Chimneys, Fireplaces, Vents, and Solid Fuel-Burning Appliances Commercial Cooking, Rethermalization and Powered Hot Food Holding and Transport Equipment Venting
Licensed for single user. © 2017 ASHRAE, Inc.
Title
Publisher
Reference
ACCA CSA AGA/NFPA NFPA NFPA NFPA SMACNA SMACNA UL UL UL AMCA ASA ISO
ACCA Manual CS-1993 ANSI Z21.12-1990 (R2000) ANSI Z223.1/NFPA 54-2015 NFPA 69-2014 NFPA 204-2015 NFPA 211-2013 SMACNA 2011 SMACNA 2011 UL 378-2006 ANSI/UL 441-2010 ANSI/UL 641-2010 ANSI/AMCA 204-05 (R2012) ANSI/ASA S2.17-1980 (R2004) ISO 2017-1:2005
ISO
ISO 2631-2:2003
ISO
ISO 6897:1984
NEBB SMACNA AHRI ASABE ASHRAE
NEBB SMACNA 2013 ANSI/AHRI 470-2006 ASAE EP342.3-2010 ANSI/ASHRAE 95-1981 (RA87)
ASHRAE
ANSI/ASHRAE 118.1-2012
ASHRAE ASHRAE ASHRAE
ANSI/ASHRAE 118.2-2006 (RA15) ANSI/ASHRAE 124-2007 ANSI/ASHRAE 137-2013
ASHRAE ASHRAE ASME ASME CSA
ANSI/ASHRAE 182-2008 (RA13) ANSI/ASHRAE 188-2015 BPVC-2015 BPVC-2015 ANSI Z21.10.1-2014/CSA 4.1-2014
CSA
ANSI Z21.10.3-2014/CSA 4.3-2014
CSA CSA CSA CSA CSA
B140.3-1962 (R2011) B140.12-03 (R2013) CAN/CSA-C22.2 No. 110-94 (R2014) CSA C191-13 CSA C745-03 (R2014)
NSF UL UL UL UL ASME AWS AWS
ANSI/NSF 5-2012 ANSI/UL 174-2004 ANSI/UL 732-1995 UL 795-2011 ANSI/UL 1453-2004 BPVC-2015 AWS D1.1M/D1.1:2010 AWS D18.1:2009
ACGIH ASTM ASTM CSA CSA NFPA NSF
ACGIH ASTM E1509-12 ASTM E1602-03 (2010)e1 CAN/CSA-B365-10 CAN/CSA-B366.1-11 ANSI/NFPA 211-2013 ANSI/NSF 4-2014
This file is licensed to John Murray ([email protected]). Publication Date: 6/1/2017
40.28
2017 ASHRAE Handbook—Fundamentals (SI)
Licensed for single user. © 2017 ASHRAE, Inc.
ORGANIZATIONS Abbrev.
Organization
Address
Telephone
http://www.
AABC
Associated Air Balance Council
1518 K Street NW, Suite 503 Washington, D.C. 20005
(202) 737-0202
aabc.com
ABMA
American Boiler Manufacturers Association
8221 Old Courthouse Road, Suite 202 Vienna, VA 22182
(703) 356-7172
abma.com
ACCA
Air Conditioning Contractors of America
2800 S. Shirlington Road, Suite 300 Arlington, VA 22206
(703) 575-4477
acca.org
ACGIH
American Conference of Governmental Industrial Hygienists
1330 Kemper Meadow Drive Cincinnati, OH 45240
(513) 742-2020
acgih.org
ADC
Air Diffusion Council
1901 N. Roselle Road, Suite 800 Schaumburg, IL 60195
(847) 706-6750
flexibleduct.org
AGA
American Gas Association
400 N. Capitol Street NW Washington, D.C. 20001
(202) 824-7000
aga.org
AHAM
Association of Home Appliance Manufacturers
1111 19th Street NW, Suite 402 Washington, D.C. 20036
(202) 872-5955
aham.org
AIHA
American Industrial Hygiene Association
3141 Fairview Park Dr, Suite 777 Falls Church, VA 22042
(703) 207-3561
aiha.org
AMCA
Air Movement and Control Association International
30 West University Drive Arlington Heights, IL 60004
(847) 394-0150
amca.org
ANSI
American National Standards Institute
1899 L Street NW, 11th Floor Washington, D.C. 20036
(202) 293-8020
ansi.org
AHRI
Air-Conditioning, Heating, and Refrigeration Institute
2111 Wilson Boulevard, Suite 500 Arlington, VA 22201
(703) 524-8800
ahrinet.org
ASA
Acoustical Society of America
1305 Walt Whitman Road, Suite 300 Melville, NY 11747-4502
(516) 576-2360
acousticalsociety.org
ASABE
American Society of Agricultural and Biological Engineers
2950 Niles Road St. Joseph, MI 49085
(269) 429-0300
asabe.org
ASHRAE
American Society of Heating, Refrigerating and Air-Conditioning Engineers
1791 Tullie Circle, NE Atlanta, GA 30329
(404) 636-8400
ashrae.org
ASME
American Society of Mechanical Engineers
Two Park Avenue New York, NY 10016-5990
(973) 882-1170
asme.org
ASTM
ASTM International
100 Barr Harbor Drive West Conshohocken, PA 19428-2959
(610) 832-9585
astm.org
AWS
American Welding Society
8669 NW 36 Street, #130 Miami, FL 33166
(305) 443-9353
aws.org
AWWA
American Water Works Association
6666 W. Quincy Avenue Denver, CO 80235
(303) 794-7711
awwa.org
BOCA
Building Officials and Code Administrators International
(see ICC)
BSI
British Standards Institution
389 Chiswick High Road London W4 4AL, UK
CAGI
Compressed Air and Gas Institute
1300 Sumner Avenue Cleveland, OH 44115
(216) 241-7333
cagi.org
CSA
Canadian Standards Association International
178 Rexdale Boulevard Toronto, ON M9W 1R3, Canada
(416) 747-4000
csagroup.org
CTI
Cooling Technology Institute
P.O. Box 681807 Houston, TX 77268
(281) 583-4087
cti.org
EJMA
Expansion Joint Manufacturers Association
25 North Broadway Tarrytown, NY 10591
(914) 332-0040
ejma.org
HEI
Heat Exchange Institute
1300 Sumner Avenue Cleveland, OH 44115
(216) 241-7333
heatexchange.org
HI
Hydraulic Institute
6 Campus Drive, First Floor North Parsippany, NJ 07054-4406
(973) 267-9700
pumps.org
HYDI
Hydronics Institute Division of AHRI
(see AHRI)
IAPMO
International Association of Plumbing and Mechanical Officials
4755 E. Philadelphia Street Ontario, CA 91761
(909) 472-4100
iapmo.org
ICBO
International Conference of Building Officials
(see ICC)
44 020 8996 9000
bsigroup.com
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Codes and Standards
40.29
Licensed for single user. © 2017 ASHRAE, Inc.
ORGANIZATIONS (Continued) Abbrev.
Organization
Address
Telephone
http://www.
ICC
International Code Council
500 New Jersey Ave NW, 6th Floor Washington, D.C. 20001
(888) 422-7233
iccsafe.org
IEEE
Institute of Electrical and Electronics Engineers
3 Park Avenue, 17th Floor New York, NY 10016-5997
(732) 981-0060
ieee.org
IES
Illuminating Engineering Society
120 Wall Street, Floor 17 New York, NY 10005-4001
(212) 248-5000
ies.org
IFCI
International Fire Code Institute
(see ICC)
IIAR
International Institute of Ammonia Refrigeration
1001 N. Fairfax St, Suite 503 Alexandria, VA 22314
(703) 312-4200
iiar.org
ISA
International Society of Automation
67 T.W. Alexander Drive, P.O. Box 12277 Research Triangle Park, NC 27709
(919) 549-8411
isa.org
ISO
International Organization for Standardization
Chemin de Blandonnet 8, CP 401 1214 Vernier, Geneva, Switzerland
41 22 749 01 11
iso.org
MCAA
Mechanical Contractors Association of America
1385 Piccard Drive Rockville, MD 20850
(301) 869-5800
mcaa.org
MICA
Midwest Insulation Contractors Association
16712 Elm Circle Omaha, NE 68130
(402) 342-3463
micainsulation.org
MSS
Manufacturers Standardization Society of the Valve and Fittings Industry
127 Park Street NE Vienna, VA 22180-4602
(703) 281-6613
mss-hq.com
NAIMA
North American Insulation Manufacturers Association
11 Canal Center Plaza, Suite 103 Alexandria, VA 22314
(703) 684-0084
naima.org
NCPWB
National Certified Pipe Welding Bureau
1385 Piccard Drive Rockville, MD 20850
(301) 869-5800
mcaa.org/ncpwb
NEBB
National Environmental Balancing Bureau
8575 Grovemont Circle Gaithersburg, MD 20877
(301) 977-3698
nebb.org
NEMA
Association of Electrical and Medical Imaging Equipment Manufacturers
1300 North 17th Street, Suite 900 Arlington, VA 22209
(703) 841-3200
nema.org
NFPA
National Fire Protection Association
1 Batterymarch Park Quincy, MA 02169-7471
(617) 770-3000
nfpa.org
NRCC
National Research Council of Canada, Institute for Research in Construction
1200 Montreal Road, Bldg M-58 Ottawa, ON K1A 0R6, Canada
(613) 993-9101
nrc-cnrc.ca
NSF
NSF International
P.O. Box 130140, 789 N. Dixboro Road Ann Arbor, MI 48105
(734) 769-8010
nsf.org
PHCC
Plumbing-Heating-Cooling Contractors Association
180 S. Washington Street, Suite 100 Falls Church, VA 22046
(703) 237-8100
phccweb.org
SAE
SAE International
400 Commonwealth Drive Warrendale, PA 15096
(724) 776-4841
sae.org
SBCCI
Southern Building Code Congress International
(see ICC)
SMACNA Sheet Metal and Air Conditioning Contractors’ National Association
4201 Lafayette Center Drive Chantilly, VA 20151-1219
(703) 803-2980
smacna.org
TEMA
Tubular Exchanger Manufacturers Association
25 North Broadway Tarrytown, NY 10591
(914) 332-0040
tema.org
UL
Underwriters Laboratories
333 Pfingsten Road Northbrook, IL 60062-2096
(847) 272-8800
ul.com
USGBC
U.S. Green Building Council
2101 L Street NW, Suite 500 Washington, D.C. 20037
(202) 742-3792
usgbc.org
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