The group of models described in this section is used to
account for non-uniform room air temperatures that may occur
within the interior air volume of a zone. These models are
accessed using the RoomAirModelType
input object. RoomAir modeling was added to EnergyPlus
starting with Version
1.2. Although there are many types of analyses (comfort,
indoor air quality, etc) that might benefit from localized
modeling of how room air varies across space, most of the room
air models in EnergyPlus only address the distribution of
temperature within the zone. This allows surface heat
transfer and air system heat balance calculations to be made
taking into account natural thermal stratification of air and
different types of intentional air distribution designs such
as under-floor and side-wall displacement ventilation that
purport to extract room air at higher-than-mean temperatures.
The exception is the RoomAirflowNetwork model, which
integrates the AirflowNetwork model and applies the nodal
airflow model within zones. Note that EnergyPlus does
not have completely general methods of
modeling room air that are applicable to every conceivable
type of airflow that might occur in a zone. Such models
(e.g. RANS-CFD) are too computationally expensive to use with
EnergyPlus for the foreseeable future. The models that are
available in EnergyPlus offer only limited modeling
capabilities for select room airflow configurations. Also note
that because the complete mixing model for room air has long
been the standard in building energy simulation, there is not
currently a consensus on how to best model non-uniform air
temperatures in buildings. Therefore, it is up to the user to
have a good understanding of when, where, and how to apply the
room air models available in EnergyPlus. The rest of this
section provides some guidance in the way of examples and
further discussion of the models available in EnergyPlus.
EnergyPlus offers the different types of air models listed
in the table below along with the input objects associated
with the use of that model.
Summary of room air models available in
EnergyPlus
Air model name
Applicability
Input Objects Required
Well-Mixed
All zones
None, default
User Defined
Any zone where the user has
prior knowledge of the temperature pattern
The room air models are coupled to the heat balance
routines using the framework described by Griffith and Chen
(2004). Their framework was modified to include features
needed for a comprehensive program for annual energy modeling
rather than one for hourly load calculations. The formulation
is largely shifted from being based on the setpoint
temperature to one based on the current mean air temperature.
This is necessary to allow for floating temperatures and dual
setpoint control where there may be times that the mean zone
temperatures are inside the dead band. The coupling framework
was also extended to allow for exhaust air flows
(e.g. bathroom exhaust fans) in addition to air system return
flows.
The inside face temperature calculation is modified by
rewriting the zone air temperature, T\(_{a}\), with an additional
subscript, i, for the surface index (\({T_{{a_j}}} \to {T_{{a_{i,j}}}}\)
or \({T_{{a_{}}}} \to
{T_{{a_i}}}\) ). The inside face heat balance is
solved for its surface temperature using,
The input object RoomAir:TemperaturePattern:UserDefined
provides a capabity for users to define the sort of air
temperature pattern he or she expects in the zone. With these
models, the pattern is generally set beforehand and does not
respond to conditions that evolve during the simulation.
(Exception: the pattern available through the RoomAir:TemperaturePattern:TwoGradient
object will switch between two different pre-defined vertical
gradients depending on the current value of certain
temperatures or thermal loads. )
The user-defined patterns obtain the mean air temperature,
\({T_{MAT}}\), from the heat
balance domain and then produce modified values for:
\({T_{ai}}\) is the
adjacent air temperature which is then used in the calculation
of inside face surface temperature during the heat balance
calculations.
\({T_{leaving}}\) is
the temperature of air leaving the zone and entering the air
system returns.
\({T_{exhaust}}\) is
the temperature of air leaving the zone and entering the
exhaust.
\({T_{stat}}\) is the
temperature of air “sensed” at the thermostat (not currently
used in air system control because air system flows use
load-based control).
The user defined room air models used indirect coupling so
that the patterns provide values for, or ways to calculate,
how specific temperatures differ from \({T_{MAT}}\). The various \(\Delta T\) values determined from
the model are applied to \({T_{MAT}}\) as follows:
\[{T_{ai}} = \Delta {T_{ai}} +
{T_{MAT}}\]
(where “i’s” represents each surface in the zone
that is affected by the model)
The patterns defined by the object ‘RoomAir:TemperaturePattern:SurfaceMapping’
are fairly straightforward. The user directly inputs values
for \(\Delta {T_{ai}}\) for
each surface. The pattern “maps” specific surfaces,
identified by name, to \(\Delta
{T_{ai}}\) values. This provides completely general
control (but in practice may be cumbersome to use). The other
patterns focus on temperature changes in the vertical
direction. Surfaces do not need to be identified, but all the
surfaces with the same height will be assigned the same \(\Delta {T_{ai}}\) values.
The patterns defined by the object
‘RoomAir:TemperaturePattern:NondimensonalHeight’ apply a
temperature profile based on a non-dimensionalized height,
\({\rm Z}\). The height of
each surface is defined to be the z-coordinate of the
surface’s centroid relative to the average z-coordinate of the
floor surfaces. The zone ceiling height is used as the length
scale to non-dimensionalize each surface’s height so that,
(where “i’s” represents each surface in the zone
that is affected by the model)
The values for \({Z_i}\)
are constrained to be between 0.01 and 0.99 because the value
is meant to describe the air layer near the surface (say
approximate 0.1 m from the surface) rather than the surface
itself.
The user-defined profile is treated as a look up table or
piecewise linear model. The values for \(\Delta {T_{ai}}\) are determined
by searching the \({\rm Z}\)
values in the user-defined profile and performing linear
interpolation on the associated \(\Delta {T_a}\) values.
The patterns defined by the object ‘RoomAir:TemperaturePattern:ConstantGradient’
apply a constant temperature gradient in the vertical
direction. The model assumes that \({T_{MAT}}\) occurs at the
mid-plane so that \({{\rm
Z}_{{T_{MAT}}}} = 0.5\) (by definition). The surface
\({Z_i}\) values are compared
to \({Z_{{T_{MAT}}}}\) and
then scaled with zone ceiling height to obtain values for the
change in height (in units of meters), \(\Delta z\). The user defined
gradient, \(grad\), (units of
\(^{\circ}\)C/m) is then used
to determine \(\Delta
{T_{ai}}\) values using:
\[\Delta {T_{ai}} = \Delta z *
grad\]
The patterns defined by the object ‘RoomAir:TemperaturePattern:TwoGradient’
are very similar to the constant gradient pattern above but
the value of \(grad\) used at
any given time is selected by interpolating between two
user-defined values for \(grad\). Five options are
available, three based on temperatures and two based on
thermal loads – see the Input Output Reference. The user
provides upper and lower bounding values. If the current
value of the “sensing” variable lies between the upper and
lower bounds, then \(grad\)
is determined using linear interpolation. If the designated
value is above the upper bound then the upper value for \(grad\) is used (no
extrapolation). Similarly, if the designated value is below
the lower bound, then the lower value for \(grad\) is used. Note that
“upper” and “lower” indicate the temperature and heat rate
bounds and that the values for \(grad\) do not have to follow in
the same way; the \(grad\)
value for the lower bound could be higher than the \(grad\) value for the upper bound
(providing a something of a reverse control scheme). Rather
than directly using \(\Delta
T\) values from the user, the temperatures for the
return air, exhaust and thermostat are determined based on
user-entered heights (in units of meters from the floor) and
applying the current value for \(grad\).
The input object RoomAirSettings:OneNodeDisplacementVentilation
provides a simple model for displacement ventilation. Mundt
(1996) points out that a floor air heat balance provides a
simple and reasonably accurate method of modeling the
temperature near the floor surface. The slope of a linear
temperature gradient can then be obtained by adding a second
upper air temperature value that comes from the usual overall
air system cooling load heat balance. The figure below
diagrams the temperature distribution versus height being
calculated by the model. Mundt’s floor air heat balance is
extended to include convection heat gain from equipment and by
ventilation or infiltration that may be introduced near the
floor in order to maintain all the terms in the air heat
balance of the Heat Balance Model. This yields the following
heat balance for a floor air node,
c\(_{p}\) is the
air specific heat at constant pressure
\(\mathop V\limits^\cdot\)
is the air system flow rate
T\(_{supply}\) is
the air system’s supply air drybulb temperature
h\(_{cFloor}\)
is the convection heat transfer coefficient for the floor
A\(_{floor}\) is
the surface area of the floor
T\(_{floor}\) is
the surface temperature of the floor
Q\(_{convSourceFloor}\) is the
convection from internal sources near the floor (< 0.2
m)
Q\(_{InfilFloor}\) is the heat
gain (or loss) from infiltration or ventilation near the
floor.
“Floor splits” are the fraction of total convective or
infiltration loads that are dispersed so as to add heat to the
air located near the floor. The user prescribes values for
floor splits as input. No guidance is known to be available
to use in recommending floor splits, but the user could for
example account for equipment known to be near the floor, such
as tower computer cases, or supplementary ventilation designed
to enter along the floor. The equation above can be solved
directly for T\(_{AirFloor}\) and is used in
the form of the equation below,
The upper air node temperature is obtained by solving the
overall air heat balance for the entire thermal zone for the
temperature of the air leaving the zone and going into the air
system return, T\(_{leaving}\).
where \({\dot Q_{sys}}\)
is the air system heat load with negative values indicating a
positive cooling load. Values for \({\dot Q_{sys}}\) are computed by
the load calculation routines and passed to the air model.
The vertical temperature gradient or slope, dT/dz, is
obtained from,
where H\(_{return}\) is the distance
between the air system return and the floor air node assumed
to be 0.1 m from the floor and z is the vertical
distance.
Height versus temperature
schematic for Mundt model [fig:height-versus-temperature-schematic-for-mundt]
The constant slope allows obtaining temperatures at any
vertical location using,
So for example the temperatures near the ceiling can easily
be determined. Accounting for the location of the thermostat
inside the zone (e.g. 1.1 m) is accomplished by returning the
temperature for the appropriate height to the appropriate air
node used for control. If the walls are subdivided in the
vertical direction as shown in the figure above, then the air
model can provide individual values for each surface based on
the height and slope. However, no additional heat balances
are necessarily made (in the air domain) at these points as
all the surface convection is passed to the model in the
totaled value for \({\dot
Q_{sys}}\).
The input object RoomAirSettings:ThreeNodeDisplacementVentilation
provides a simple model for heat transfer and vertical
temperature profile prediction in displacement ventilation.
The fully-mixed room air approximation that is currently used
in most whole building analysis tools is extended to a three
node approach, with the purpose of obtaining a first order
precision model for vertical temperature profiles in
displacement ventilation systems. The use of three nodes
allows for greatly improved prediction of thermal comfort and
overall building energy performance in low energy cooling
strategies that make use of unmixed stratified ventilation
flows.
The UCSD Displacement Ventilation Model is one of the
non-uniform zone models provided through the Room Air Manager
in EnergyPlus. The intent is to provide a selection of useful
non-uniform zone air models to enable the evaluation of
air-conditioning techniques that use stratified or partially
stratified room air. Such techniques include displacement
ventilation (DV) and underfloor air distribution (UFAD)
systems. The methodology can also include, in principle,
natural displacement ventilation and also wind-driven
cross-ventilation (CV).
A DV system is a complete contrast to a conventional forced
air system. In a conventional system conditioned air is
delivered at ceiling level and the intent is to create a fully
mixed space with uniform conditions. In a DV system
conditioned air is delivered at floor level and low velocity
in order to minimize mixing and to establish a vertical
temperature gradient. The incoming air “displaces” the air
above it which, in turn, is exhausted through ceiling level
vents. In DV a noticeable interface occurs between the
occupied zone of the room and a mixed hot layer near the
ceiling of the room(Dominique & Guitton, 1997).
Maintaining the lower boundary of this warm layer above the
occupied zone is one of the many unique challenges of
displacement ventilation design. Often DV systems use 100%
outside air. The vertical displacement air movement means that
convective heat gains introduced near the ceiling will be
removed without affecting the occupied region of the room.
Also a fraction of the heat gains that occur in the occupied
zones rise as plumes into the upper part of the space, thereby
reducing the cooling load. Similarly the fresh air will be
used more effectively than with a fully mixed system: the
fresh air won’t be “wasted” in the upper, unoccupied region of
the room. Finally, the vertical temperature gradient means
that the average room temperature can be higher for a DV
conditioned room than with a conventionally conditioned room:
the occupants feel the lower temperature in the lower region
of the room and are unaffected by the higher temperature near
the ceiling. However, whenever the outside air temperature is
above \(\approx\)19\(^{\circ}\)C this advantage is
mostly lost: the internal loads must be removed from the space
independently of the airflow pattern (during the warmer hours
buildings tend to be almost closed to the outside, operating
in closed loop). The inflow temperature advantage is then only
useful for the minimum outside air that must always be
provided (in most cases this remaining advantage is
negligible).
DV systems have limitations. In order to avoid chilling the
occupants the supply air temperature used for DV is
considerably higher than that used in conventional forced-air
systems. This can lead to problems in removing both sensible
and latent loads. Exterior spaces may have conditions that are
not conducive to establishing a vertical temperature gradient.
DV systems seem to be best suited to interior spaces with only
moderate loads.
Several types of models have been proposed as suitable for
inclusion in building energy simulation (BES) programs. These
models must be simple enough not to impose an undue
computational burden on a BES program, yet provide enough
predictive capability to produce useful comparisons between
conventional and stratified zone operation strategies. ASHRAE
RP-1222 (Chen & Griffith 2002) divides the candidate
models into two categories: nodal and zonal.
Nodal models describe the zone air as a network of nodes
connected by flow paths; each node couples convectively to one
or more surfaces. Zonal models are coarse–grained finite
volume models. ASHRAE RP-1222 provides a short history (and
examples) of each type of model. In terms of nodal models for
displacement ventilation we mention the Mundt model (Mundt
1996), since it is implemented in EnergyPlus, and the
Rees-Haves model (Rees & Haves 2001) since it is a well
developed nodal-type model and is implemented in the RP-1222
toolkit. The Rees-Haves model, while successful in predicting
the flow and temperature field for geometries similar to those
used in its development, can suffer from lack of flexibility
and clarity in the modeling approximations. When dealing with
diverse geometries it is not clear that the flow coefficients
used in the model are applicable or why they can be used since
plumes, the fundamental driving mechanisms of the displacement
flow, are not explicitly modeled. This is the main difference
between the DV models implemented in theRP-1222 toolkit and
the model that is described here.
The UCSD DV model is closer to a nodal model than to a
zonal model. However, it is best to classify it in a separate
category: plume equation based multi-layer models (Linden
et al. 1990, Morton et al. 1956). These
models assume that the dominant mechanism is plume-driven flow
from discrete internal sources and that other effects (such as
buoyancy driven flow at walls or windows) may be neglected.
Alternatively, these heat sources also produce plumes that can
be included in the model. The result is a zone divided
vertically into two or more well separated regions – each
region characterized by a single temperature or temperature
profile. This characterization allows the physics of the heat
gains and the ventilation flow to be represented in a
realistic manner, without the introduction of ad hoc
assumptions.
The simplest form of the plume equation based models is the
case of a single plume in an adiabatic box with constant
supply air flow. For this configuration two layers form in the
room: a lower layer with similar density and temperature as
the inflow air and a mixed upper layer with the same density /
temperature as the outflow air. The main assumption of this
model, successfully validated against scaled model experiments
(Linden et al. 1990), is that the interface between
the two layers occurs at the height (h) where the vertical
buoyancy driven plume flow rate is the same as the inflow
rate. For a point source of buoyancy in a non-stratified
environment (a plume) the airflow rate increases with vertical
distance from the source according to:
\(\alpha\) = plume
entrainment constant; a value of 0.127 is used, suitable for
top-hat profiles for density and velocity across the
plumes.
For an ideal gas:
\[\Delta \rho /\rho = \Delta
{\rm T}/{\rm T}\]
resulting in the following relation between heat input rate
and buoyancy flux:
\[B = \frac{{g\dot Q}}{{\rho
{C_p}T}}\]
where
\(\rho\) = density of air
[kg/m3]
\(T\) = air temperature
[K]
\(g\) = acceleration of
gravity [m/s2]
\(\dot Q\) = heat input
rate [W]
\({C_p}\) = specific heat
capacity of air [J/kgK]
Since the plume volume flow rate increases with height with
exponent 5/3, for any room inflow rate (F, (m\(^{3}\)/s)) there will always be a
height (h,(m)) where the plume driven flow rate matches the
inflow rate. This height is obtained by setting (1.1) equal to
F and solving for z = h:
\[h =
{F^{3/5}}/({C^{3/5}}{B^{1/5}})\]
Substituting in and introducing air properties at 20\(^{\circ}\)C gives:
Of course, it would be rare for a real world case to
consist of a single point-source plume originating on the
floor, unaffected by heat gains from walls and windows. For
multiple plumes of equal strength a straight-forward extension
of the single is possible. N plumes of unequal strength result
in the formation of n vertical layers. This case is much more
complex but if we are satisfied with a first order precision
model the equal strength model can be used by averaging the
plume strengths (Carrilho da Graça, 2003). Even in a case
where all plumes are of equal strength, nearby plumes may
coalesce. Plumes that are less than 0.5 meters apart at their
source will coalesce within 2 meters (Kaye &
Linden,2004).
As the complexity of the physical systems modeled increases
some limitations must be imposed. In particular, the biggest
challenge remains the interaction between wall driven boundary
layers (positively and negatively buoyant) and displacement
flows. For this reason, the model that is developed below is
not applicable when:
Downward moving buoyancy driven airflow rate is of the
same order of magnitude as plume driven flow (these airflow
currents are typically generated on lateral surfaces or in the
ceiling whenever these surfaces are much cooler than the room
air).
Upward moving wall or floor generated buoyancy flux in
the lower layer is of the same order of magnitude as plume
driven flow.
Although these limitations are significant it is important
to note that even in the presence of dominant convection from
the floor surface, a buoyancy, two layer flow can be
established whenever the plume buoyancy flux is more than 1/7
of the horizontal flux (Hunt et al. 2002). A two
layer structure can also originate when the only heat source
is a heated portion of the room floor, as long as the heated
area does not exceed 15% of the room floor (Holford et
al. 2002).
For the case of multiple non-coalescing plumes (n), with
equal strength, the total vertical airflow for a given height
is:
The model predicts three temperatures that characterize the
three main levels in the stratification of the room:
A floor level temperature T\(_{floor}\) to account for the
heat transfer from the floor into the supply air.
An occupied subzone temperature T\(_{oc}\) representing the
temperature of the occupied region.
An upper level temperature T\(_{mx}\) representing the
temperature of the upper, mixed region and the outflow
temperature.
We assume that the model for multiple, equal strength
plumes (equations and will be adequate for our calculations.
The supply air flow rate \(\dot
V\) is obtained by summing all the air flows entering
the zone: supply air, infiltration, ventilation, and
inter-zone flow. The heat gain \(\dot Q\) is estimated by summing
all the convective internal gains located in the occupied
subzone – task lights, people, equipment – and dividing this
power equally among the n plumes. With these assumptions we
can describe the implementation.
The UCSD DV model is controlled by the subroutine
ManageUCSDDVModel which is called from the
RoomAirModelManager. The RoomAirModelManager
selects which zone model will be used for each zone.
The calculation is done in subroutine CalcUCSDDV.
First we calculate the convective heat gain going into the
upper and lower regions.
Next we sum up the inlet air flows in the form of MCP (mass
flow rate times the air specific heat capacity) and MCPT (mass
flow rate times C\(_{p}\)
times air temperature).
The gains fraction \(F{r_{gains}}\) is a user input
via a schedule. It is the fraction of the convective gains in
the occupied subzone that remain in that subzone. Using this
we calculate the total power in the plumes and the power per
plume.
where 0.000833 = \(1/({\rho_{air}} \cdot
{c_{p,air}})\) converts \(MC{P_{tot}}\) to a volumetric
flow rate. Next we iterate over the following 3 steps.
Call subroutine HcUCSDDV to calculate a
convective heat transfer coefficient for each surface in the
zone, an effective air temperature for each surface, and
HA\(_{mx}\), HAT\(_{mx}\), HA\(_{oc}\), HAT\(_{oc}\), HA\(_{fl}\), and HAT\(_{fl}\). Here HA is \(\sum\limits_{surfaces} {{h_{c,i}}} \cdot
{A_i}\) for a region and HAT is \(\sum\limits_{surfaces} {{h_{c,i}}} \cdot
{A_i} \cdot {T_i}\) for a region. The sum is over all
the surfaces bounding the region; \({h_{c,i}}\) is the convective
heat transfer coefficient for surface i, \({A_i}\) is the area of surface
i, and \({T_i}\) is the
surface temperature of surface i.
Calculate the three subzone temperatures: T\(_{fl,}\) T\(_{oc}\) and T\(_{mx}\).
The h\(_{c}\)’s calculated
in step 1 depend on the subzone temperatures and the boundary
layer height. In turn the subzone temperatures depend on the
HA and HAT’s calculated in step 1. Hence the need for
iteration
Subroutine HcUCSDDV is quite straightforward. It
loops through all the surfaces in each zone and decides
whether the surface is located in the upper, mixed subzone or
the lower, occupied subzone, or if the surface is in both
subzones. If entirely in one subzone the subzone temperature
is stored in the surface effective temperature variable
SurfTempEffBulkAir(SurfNum) and h\(_{c}\) for the surface is
calculated by a call to subroutine
CalcDetailedHcInForDVModel. This routine uses the
“detailed” natural convection coefficient calculation that
depends on surface tilt and \(\Delta
{T^{1/3}}\). This calculation is appropriate for
situations with low air velocity.
For surfaces that bound 2 subzones, the subroutine
calculates h\(_{c}\) for each
subzone and then averages them, weighting by the amount of
surface in each subzone.
During the surface loop, once the h\(_{c}\) for a surface is
calculated, the appropriate subzone HA and HAT sums are
incremented. If a surface is in 2 subzones the HA and HAT for
each subzone are incremented based on the area of the surface
in each subzone.
The calculation of subzone temperatures follows the method
used in the
ZoneTempPredictorCorrector module
and described in the section Basis for the System and
Zone
Integration. Namely a third order finite difference
expansion of the temperature time derivative is used in
updating the subzone temperatures. Otherwise the subzone
temperatures are obtained straightforwardly by solving an
energy balance equation for each subzone.
Here \({C_{air,fl}}\),
\({C_{air,oc}}\), and \({C_{air,mx}}\) are the heat
capacities of the air volume in each subzone. \({C_{air,mx}}\) is calculated
by:
The above iterative procedure assumed that displacement
ventilation was taking place: i.e., conditions were favorable
temperature stratification in the zone. Now that this
calculation is complete and the subzone temperatures and
depths calculated, we check to see if this assumption was
justified. If not, zone conditions must be recalculated
assuming a well-mixed zone.
If \({T_{mx}} <
{T_{oc}}\) or \(MC{P_{tot}}
\le 0\) or \({H_{fr}} \cdot
{H_{ceil}} < {H_{fl,top}} + \Delta {z_{occ,\min }}\)
then the following mixed calculation will replace the
displacement ventilation calculation.
Note:\(\Delta
{z_{occ,\min }}\) is the minimum thickness of occupied
subzone. It is set to 0.2 meters. \({H_{fl,top}}\) is the height of
the top of the floor subzone. It is defined to be 0.2 meters;
that is, the floor subzone is always 0.2 meters thick and
\({T_{fl}}\) is the
temperature at 0.1 meter above the floor surface.
The mixed calculation iteratively calculates surface
convection coefficients and room temperature just like the
displacement ventilation calculation described above. In the
mixed case however, only one zone temperature T\(_{avg}\) is calculated. The
3 subzone temperatures are then set equal to T\(_{avg}\).
The displacement ventilation calculation finishes by
calculating some report variables. Using Equation [eq:RoomAirHeightBoundaryLayer],
setting the boundary height to 1.5 meters and solving for the
flow, we calculate a minimum flow fraction:
The input object RoomAirSettings:UnderFloorAirDistributionInterior
provides a simple model for heat transfer and nonuniform
vertical temperature profile for interior zones of a UFAD
system. These zones are expected to be dominated by internal
loads, a portion of which (such as occupants and workstations)
will act to generate plumes. The plumes act to potentially
create room air stratification, depending on the type &
number of diffusers, the amount and type of load, and the
system flowrate. In order to better model this situation the
fully-mixed room air approximation that is currently used in
most whole building analysis tools is extended to a two node
approach, with the purpose of obtaining a first order
precision model for vertical temperature profiles for the
interior zones of UFAD systems. The use of 2 nodes allows for
greatly improved prediction of thermal comfort and overall
building energy performance for the increasingly popular UFAD
systems.
The UCSD UFAD Interior Zone
Model is one of the non-uniform zone models provided through
the Room Air Manager in EnergyPlus. The intent is to provide a
selection of useful non-uniform zone air models to enable the
evaluation of air-conditioning techniques that use stratified
or partially stratified room air. Such techniques include
displacement ventilation (DV) and underfloor air distribution
(UFAD) systems. The methodology can also include natural
displacement ventilation and also wind-driven
cross-ventilation (CV).
UFAD systems represent, in terms of room air
stratification, an intermediate condition between a well-mixed
zone and displacement ventilation. Air is supplied through an
underfloor plenum at low pressure through diffusers in the
raised floor. The diffusers can be of various types: e.g.,
swirl, variable-area, displacement, and produce more or less
mixing in the zone. UFAD systems are promoted as saving energy
due to: higher supply air temperature; low static pressure;
cooler conditions in the occupied subzone than in the upper
subzone; and sweeping of some portion of the convective load
(from ceiling lights, for instance) into the return air
without interaction with the occupied region of the zone.
Modeling a UFAD system is quite complex and involves
considerably more than just a non-uniform zone model. The
zones’ coupling to the supply and return plenums must be
modeled accurately (particularly radiative transfer from a
warm ceiling to a cool floor and on into the supply plenum by
conduction). The supply plenum must be accurately modeled,
giving a good estimate of the supply air temperature and
conduction heat transfer between supply & return plenums
through the slab. The HVAC system must be modeled including
return air bypass and various types of fan powered terminal
units.
The UCSD UFAD interior zone model is similar to the UCSD DV
model. The most obvious difference is that the UFAD model has
no separate near-floor subzone. Like the UCSD DV model it is a
plume equation based multi-layer model (2 layers in this
case). The zone is modeled as being divided into 2 well
separated subzones which we denote as “occupied” and “upper”.
Each subzone is treated as having a single temperature. The
boundary between the 2 subzones moves up & down each time
step as a function of zone loads and supply air flow rate.
Thus at each HVAC time step, the height of the boundary above
the floor must be calculated, portions of surfaces assigned to
each subzone, and a separate convective heat balance performed
on each subzone.
The UFAD interior zone model is based upon non-dimensional
analysis of the system and using the non-dimensional
description to make a comparison between full-scale UCB test
chamber data & small-scale UCSD salt tank
measurements.
In order to do the non-dimensional comparisons, we need to
define two dimensionless parameters. One is \(\Gamma\), and the other is \(\phi\). Lin & Linden (Lin
& Linden, 2005) showed that in a UFAD system, the buoyancy
flux of the heat source \((B)\) and the momentum flux of
the cooling jets \((M)\) are
the controlling parameters on the stratification. Since \([B] = {L^4}{T^{ - 3}}\) and\([M] = {L^4}{T^{ - 2}}\) , we can
have a length scale as \({M^{3/4}}/{B^{1/2}}\) .
Definition of \(\Gamma\) for the single-plume
single-diffuser basic model
We observed, in our small-scale experiments, that the total
room height does not affect the interface position, or the
height of the occupied zone. In other words, H might
not be the critical length scale for the stratification.
Therefore, we started to use \(\sqrt
A\) as the reference length. Then \(\Gamma\) is defined as:
Definition for multi-diffuser and multi-source
cases
We only considered single-diffuser, single-source cases in
above analysis. Suppose there are n equal diffusers
and m equal heat sources in a UFAD room. We shall
divide the number of diffusers up into a number of separate
heat sources so that each subsection with n’ = n/m
diffusers per heat source will have the same stratification as
other subsections. Further, the air flow and the heat load
into the subsection Q’ and B’ will be \(Q' = Q/m\)\(B' = B/m\) respectively,
where Q’ and B’ are the total air flow and
the total heat load for the entire UFAD space. Then the
momentum flux each diffuser per heat source carries is \({M_d} =
{(\frac{1}{{n'}}Q')^2}/A\). Equation [eq:GammaSinglePlumeSingleDiffuser]
will be modified as:
Because B is the buoyancy flux of the heat sources
and M is the momentum flux of the cooling jets, in a
real full-scale room, we shall consider the total room net
heat load (plume heat input, minus the room losses) and the
total net flow rate coming from the diffusers (input room air
flow, minus the room leakage). Further, if the diffuser is
swirl type, the vertical momentum flux should be used.
where Q is the net flow rate coming out from all
diffusers (m\(^{3}\)/s);
W is the total net heat load (kW); A is the
effective area of each diffuser (m\(^{2}\)); n’ is the
number of diffusers per heat source; q\(\theta\) is the angle
between the diffuser slots and the vertical direction and
m is the number of heat sources.
Definition of \(\phi\)
In our theoretical model, two-layer stratification forms at
steady state, provided that each diffuser carries the same
momentum flux, and each heat source has the same heat load. We
could define a dimensionless parameter \(\phi\), which indicates the
strength of stratification.
Small-scale cases
In our salt-water tank experiments, fluid density
r is measured. Define that:
where \(\rho_{u}\) and \(\rho_{l}\) are the fluid
density of the upper layer and lower layer, separately; and
\(\rho_{o}\) is the
source density at the diffusers.
Therefore, \(\rho_{l}\) = \(\rho_{o}\) gives \(\phi\) = 1, which
means the largest stratification (displacement ventilation
case); \(\rho_{l}\)
= \(\rho_{u}\) leads
to \(\phi\) =
0, in which case there is no stratification (mixed
ventilation case).
Full-scale cases
Similarly, we can define \(\phi\) for full-scale cases
by using temperature.
where T\(_{r}\),
T\(_{oz}\), and
T\(_{s}\) are the
return air temperature, the occupied zone temperature and the
supply temperature, respectively (K). Again \(\phi\) = 1 occurs in
displacement ventilation; while \(\phi\) = 0 happens in mixed
ventilation.
Comparisons between full-scale UCB data and small-scale
UCSD data
Data comparisons in the
non-dimensional (a) regular \(\Gamma\)-\(\phi\) plot [fig:data-comparisons-in-the-non-dimensional-a](b) log-log \(\Gamma\)-\(\phi\) plot.} [fig:b-log-log-g-fplot.]
The figures (Figure 2.
Data comparisons in the non-dimensional (a) regular \(\Gamma\)-\(\phi\) plot and Figure 3. (b) log-log
\(\Gamma\)-\(\phi\) plot.} show the
comparisons between UCB’s data and the UCSD salt tank data in
the \(\Gamma\)-\(\phi\) plot. As seen in the
figures, the full-scale and small-scale data are on the same
trend curve. This provides the evidence that the salt-tank
experiments have included most characteristics of a UFAD
system. Note that big \(\Gamma\) (>20) of UCB’s
experiments all have large DDR (from 1.19 to
2.18). The largest DDR (2.18) even gives a
negative \(({T_r} - {T_{oz}})/({T_r}
- {T_s})\), which is NOT shown in the figures.)
We could work out the occupied zone temperature by using
the least-square fitting line suggested in Figure 3. Hence the
interface height is needed to determine a entire two-layer
stratification. Figure 4
shows the dimensionless interface height\((h/\sqrt {n'A} )\) of the
UCSD small-scale experiments plotted against \(\Gamma\). Note that for the
experiments with elevated heat source, the interface heights
have been modified by\(h' = h -
\frac{1}{2}{h_s}\) where h\(_{s}\) is the vertical
position of the elevated heat source. All data then are
located along a line in Figure 4.
Since the salt-tank experiments are concluded to represent
important characteristics of a full-scale UFAD room, this
figure provides some guidelines for estimate the interface
position in a real UFAD room.
Non-dimensional interface
height of small-scale experiments. [fig:non-dimensional-interface-height-of-small]
Formulas for EnergyPlus based on the dimensionless
parameter \(\Gamma\)
If we have input including the supply temperature
T\(_{s}\) (K); the
number of diffusers n; the number of heat sources
m; the vertical position of the heat sources
h\(_{s}\) (m); the
heat load W (kW); the effective area of a diffuser
A (m\(^{2}\)); and
the total supply air flow rate Q (m\(^{3}\)/s) then the output will
be;
where T\(_{r}\)
is the return temperature (K); T\(_{oz}\) is the occupied
subzone temperature (K); h is the interface height
(m); and \(\Gamma\)
is defined above.
The implementation closely follows the procedure described
in the displacement ventilation zone model. The model predicts
two temperatures that characterize the two main levels in the
stratification of the room:
An occupied subzone temperature T\(_{oc}\) representing the
temperature of the occupied region;
An upper level temperature T\(_{mx}\) representing the
temperature of the upper, mixed region and the outflow
temperature.
We will use Equation [eq:InterfaceHeightEquation]
to calculate the interface height and do a heat balance
calculation on each subzone. \(\Gamma\) is given by
Equation [eq:GammaFullScaleCases].
The supply air flow rate \(\dot
V\) is obtained by summing all the air flows entering
the zone: supply air, infiltration, ventilation, and
inter-zone flow. The heat gain \(\dot Q\) is estimated by summing
all the convective internal gains located in the occupied
subzone – task lights, people, equipment – and dividing this
power equally among the n plumes. With these assumptions we
can describe the implementation.
The UCSD UFI model is controlled by the subroutine
ManageUCSDUFModels which is called from the
RoomAirModelManager. The RoomAirModelManager
selects which zone model will be used for each zone.
The calculation is done in subroutine CalcUCSDUI.
First we calculate the convective heat gain going into the
upper and lower regions.
Next we sum up the inlet air flows in the form of MCP (mass
flow rate times the air specific heat capacity) and MCPT (mass
flow rate times C\(_{p}\)
times air temperature).
The area A\(_{diff}\) is also a user
input. For swirl diffusers and for displacement diffusers this
area is used as input. For the variable area diffusers,
though, we calculate the area. We assume 400 ft/min velocity
at the diffuser and a design flow rate per diffuser is 150 cfm
(.0708 m\(^{3}\)/s). The
design area of the diffuser is 150 ft\(^{3}\)/min / 400 ft/min = .575
ft\(^{2}\) = .035 m\(^{2}\). Then the variable area
each time step is
Call subroutine HcUCSDUF to calculate a
convective heat transfer coefficient for each surface in the
zone, an effective air temperature for each surface, and
HA\(_{mx}\), HAT\(_{mx}\), HA\(_{oc}\), HAT\(_{oc}\). Here HA is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot
{A_i}}\) for a region and HAT is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot
{A_i}} \cdot {T_i}\) for a region. The sum is over all
the surfaces bounding the region; \({{h_{c,i}}}\) is the convective
heat transfer coefficient for surface i, \({A_i}\) is the area of surface i,
and \({T_i}\) is the surface
temperature of surface i.
Calculate the two subzone temperatures: T\(_{oc}\) and T\(_{mx}\).
The h\(_{c}\)’s calculated
in step 1 depend on the subzone temperatures. In turn the
subzone temperatures depend on the HA and HAT’s calculated in
step 1. Hence the need for iteration.
Subroutine HcUCSDUF is quite straightforward. It
loops through all the surfaces in each zone and decides
whether the surface is located in the upper, mixed subzone or
the lower, occupied subzone, or if the surface is in both
subzones. If entirely in one subzone the subzone temperature
is stored in the surface effective temperature variable
SurfTempEffBulkAir(SurfNum) and h\(_{c}\) for the surface is
calculated by a call to subroutine
CalcDetailedHcInForDVModel. This routine uses the
“detailed” natural convection coefficient calculation that
depends on surface tilt and \(\Delta
{T^{1/3}}\) . This calculation is appropriate for
situations with low air velocity.
For surfaces that bound 2 subzones, the subroutine
calculates h\(_{c}\) for each
subzone and then averages them, weighting by the amount of
surface in each subzone.
During the surface loop, once the h\(_{c}\) for a surface is
calculated, the appropriate subzone HA and HAT sums are
incremented. If a surface is in 2 subzones the HA and HAT for
each subzone are incremented based on the area of the surface
in each subzone.
The calculation of subzone temperatures follows the method
used in the
ZoneTempPredictorCorrector module
and described in the section Basis for the System and
Zone
Integration. Namely a third order finite difference
expansion of the temperature time derivative is used in
updating the subzone temperatures. Otherwise the subzone
temperatures are obtained straightforwardly by solving an
energy balance equation for each subzone.
The gains fraction \(F{r_{gains}}\) is a user input
via a schedule. It is the fraction of the convective gains in
the occupied subzone that remain in that subzone.
The other subzone air heat capacities are calculated in the
same manner.
The above iterative procedure assumed that the UFAD
nonuniform zone model was appropriate: i.e., conditions were
favorable temperature stratification in the zone. Now that
this calculation is complete and the subzone temperatures and
depths calculated, we check to see if this assumption was
justified. If not, zone conditions must be recalculated
assuming a well-mixed zone.
If \({T_{mx}} <
{T_{oc}}\) or \(MC{P_{tot}}
\le 0\) or \({H_{fr}} \cdot
{H_{ceil}} < \Delta {z_{occ,\min }}\), then the
following mixed calculation will replace the UFAD interior
zone calculation.
Note:\(\Delta
{z_{occ,min}}\) is the minimum thickness of occupied
subzone. It is set to 0.2 meters.
The mixed calculation iteratively calculates surface
convection coefficients and room temperature just like the
displacement ventilation calculation described above. In the
mixed case however, only one zone temperature T\(_{avg}\) is calculated. The
3 subzone temperatures are then set equal to T\(_{avg}\).
The input object RoomAirSettings:UnderFloorAirDistributionExterior
provides a simple model for heat transfer and a nonuniform
vertical temperature profile for exterior zones of a UFAD
system. These zones are expected to be dominated by internal
loads, a portion of which (such as occupants and workstations)
will act to generate plumes, and by window solar and
conduction heat gains. The solar radiation penetrating the
room is not expected to generate plumes. However, a window
plume is likely to be generated in sunny conditions,
particularly if an interior blind is deployed. Thus the
exterior UFAD zone will have potentially have plumes from
people and equipment and plumes arising from the windows. The
plumes act to potentially create room air stratification,
depending on the type & number of diffusers, the amount
and type of load, and the system flowrate. In order to better
model this situation the fully-mixed room air approximation
that is currently used in most whole building analysis tools
is extended to a two node approach, with the purpose of
obtaining a first order precision model for vertical
temperature profiles for the exterior zones of UFAD systems.
The use of 2 nodes allows for greatly improved prediction of
thermal comfort and overall building energy performance for
the increasingly popular UFAD systems.
TheUCSD UFAD Exterior Zone
Model is one of the non-uniform zone models provided through
the Room Air Manager in EnergyPlus. The intent is to provide a
selection of useful non-uniform zone air models to enable the
evaluation of air-conditioning techniques that use stratified
or partially stratified room air. Such techniques include
displacement ventilation (DV) and underfloor air distribution
(UFAD) systems. The methodology can also include natural
displacement ventilation and also wind-driven
cross-ventilation (CV).
UFAD systems represent, in terms of room air
stratification, an intermediate condition between a well-mixed
zone and displacement ventilation. Air is supplied through an
underfloor plenum at low pressure through diffusers in the
raised floor. The diffusers can be of various types: e.g.,
swirl, variable-area, displacement, and produce more or less
mixing in the zone. UFAD systems are promoted as saving energy
due to: higher supply air temperature; low static pressure;
cooler conditions in the occupied subzone than in the upper
subzone; and sweeping of some portion of the convective load
(from ceiling lights, for instance) into the return air
without interaction with the occupied region of the zone.
Modeling a UFAD system is quite complex and involves
considerably more than just a non-uniform zone model. The
zones’ coupling to the supply and return plenums must be
modeled accurately (particularly radiative transfer from a
warm ceiling to a cool floor and on into the supply plenum by
conduction). The supply plenum must be accurately modeled,
giving a good estimate of the supply air temperature and
conduction heat transfer between supply & return plenums
through the slab. The HVAC system must be modeled including
return air bypass and various types of fan powered terminal
units.
The UCSD UFAD exterior zone model is similar to the UCSD
interior zone model. The most obvious difference is that the
exterior UFAD has 2 different types of plume sources: people
& equipment and windows. Like the UCSD UFAD interior model
it is a plume equation based multi-layer model (2 layers in
this case). The zone is modeled as being divided into 2 well
separated subzones which we denote as “occupied” and “upper”.
Each subzone is treated as having a single temperature. The
boundary between the 2 subzones moves up & down each time
step as a function of zone loads and supply air flow rate.
Thus at each HVAC time step, the height of the boundary above
the floor must be calculated, portions of surfaces assigned to
each subzone, and a separate convective heat balance performed
on each subzone.
where T\(_{oz}\)
is the occupied subzone temperature (K); T\(_{r}\) is the return
temperature (K); \(\Gamma\) is the
dimensionless height parameter defined above; T\(_{s}\) is the supply
temperature (K); h is the interface height (m);
n is the number of diffusers; m is the
number of heat sources; A is the effective area of a
diffuser (m\(^{2}\)); and
h\(_{s}\) is the
vertical position of the heat sources (m). The formula for
T\(_{r}\) is the
same as in the previous section.
The implementation closely follows the procedure described
in the UFAD interior zone model. The model predicts two
temperatures that characterize the two main levels in the
stratification of the room:
An occupied subzone temperature T\(_{oc}\) representing the
temperature of the occupied region;
An upper level temperature T\(_{mx}\) representing the
temperature of the upper, mixed region and the outflow
temperature.
We will use Equation [eq:InterfaceHeightUFAD]
to calculate the interface height and do a heat balance
calculation on each subzone. \(\Gamma\) is given by
Equation [eq:GammaFullScaleCases].
The supply air flow rate \(\dot
V\) is obtained by summing all the air flows entering
the zone: supply air, infiltration, ventilation, and
inter-zone flow. The heat gain \(\dot Q\) is estimated by summing
all the convective internal gains located in the occupied
subzone – task lights, people, equipment – and adding to this
the convective gain coming from the window surface. With these
assumptions we can describe the implementation.
The UCSD UFE model is controlled by the subroutine
ManageUCSDUFModels which is called from the
RoomAirModelManager. The RoomAirModelManager
selects which zone model will be used for each zone.
The calculation is done in subroutine CalcUCSDUEI.
First we calculate the convective heat gain going into the
upper and lower regions.
Next we sum up the inlet air flows in the form of MCP (mass
flow rate times the air specific heat capacity) and MCPT (mass
flow rate times C\(_{p}\)
times air temperature).
For exterior zone model, we assume one plume: \({N_{plumes}} = 1\). The number of
diffusers in the zone \({N_{diffusers}}\) is a user
input.
The area A\(_{diff}\) is also a user
input. For swirl diffusers, linear bar grilles, and
displacement diffusers this area is used as input. For the
variable area diffusers, though, we calculate the area. We
assume 400 ft/min velocity at the diffuser and a design flow
rate per diffuser is 150 cfm (.0708 m\(^{3}\)/s). The design area of the
diffuser is 150 ft\(^{3}\)/min / 400 ft/min = .575
ft\(^{2}\) = .035 m\(^{2}\). Then the variable area
each time step is:
Call subroutine HcUCSDUF to calculate a
convective heat transfer coefficient for each surface in the
zone, an effective air temperature for each surface, and
HA\(_{mx}\), HAT\(_{mx}\), HA\(_{oc}\), HAT\(_{oc}\), HA\(_{mx,win}\), HAT\(_{mx,win}\), HA\(_{oc,win}\),and HAT\(_{oc,win}\). Here HA is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot
{A_i}}\) for a region and HAT is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot
{A_i}} \cdot {T_i}\) for a region. The sum is over all
the surfaces bounding the region; \({h_{c,i}}\) is the convective
heat transfer coefficient for surface i, \({A_i}\) is the area of surface i,
and \({T_i}\) is the surface
temperature of surface i. Variables with the win
subscript are summed over window surfaces only. Then the
convective gain from the window is recalculated using
Equation [eq:QdotwinconvUSCDUFAD]
and the power in the plume is recalculated: \({\dot Q_{plumes}} = {\dot Q_{ocz}} +
{\dot Q_{win,conv}}\) .
Calculate the two subzone temperatures: T\(_{oc}\) and T\(_{mx}\).
The h\(_{c}\)’s calculated
in step 1 depend on the subzone temperatures. In turn the
subzone temperatures depend on the HA and HAT’s calculated in
step 1. Hence the need for iteration.
Subroutine HcUCSDUF is quite straightforward. It
loops through all the surfaces in each zone and decides
whether the surface is located in the upper, mixed subzone or
the lower, occupied subzone, or if the surface is in both
subzones. If entirely in one subzone the subzone temperature
is stored in the surface effective temperature variable
SurfTempEffBulkAir(SurfNum) and h\(_{c}\) for the surface is
calculated by a call to subroutine
CalcDetailedHcInForDVModel. This routine uses the
“detailed” natural convection coefficient calculation that
depends on surface tilt and \(\Delta
{T^{1/3}}\) . This calculation is appropriate for
situations with low air velocity.
For surfaces that bound 2 subzones, the subroutine
calculates h\(_{c}\) for each
subzone and then averages them, weighting by the amount of
surface in each subzone.
During the surface loop, once the h\(_{c}\) for a surface is
calculated, the appropriate subzone HA and HAT sums are
incremented. If a surface is in 2 subzones the HA and HAT for
each subzone are incremented based on the area of the surface
in each subzone. The subroutine calculates a separate HA and
HAT for the windows for use in calculating the window
convective gain.
The calculation of subzone temperatures follows the method
used in the
ZoneTempPredictorCorrector module
and described in the section Basis for the System and
Zone
Integration. Namely a third order finite difference
expansion of the temperature time derivative is used in
updating the subzone temperatures. Otherwise the subzone
temperatures are obtained straightforwardly by solving an
energy balance equation for each subzone.
The gains fraction \(F{r_{gains}}\) is a user input
via a schedule. It is the fraction of the convective gains in
the occupied subzone that remain in that subzone.
The other subzone air heat capacities are calculated in the
same manner.
The above iterative procedure assumed that the UFAD zone
model was applicable: i.e., conditions were favorable
temperature stratification in the zone. Now that this
calculation is complete and the subzone temperatures and
depths calculated, we check to see if this assumption was
justified. If not, zone conditions must be recalculated
assuming a well-mixed zone.
If \({T_{mx}} <
{T_{oc}}\) or \(MC{P_{tot}}
\le 0\) or \({H_{fr}} \cdot
{H_{ceil}} < \Delta {z_{occ,\min }}\), then the
following mixed calculation will replace the UFAD exterior
zone calculation.
Note:\(\Delta
{z_{occ,min}}\) is the minimum thickness of occupied
subzone. It is set to 0.2 meters.
The mixed calculation iteratively calculates surface
convection coefficients and room temperature just like the
displacement ventilation calculation described above. In the
mixed case however, only one zone temperature T\(_{avg}\) is calculated. The
2 subzone temperatures are then set equal to T\(_{avg}\).
The input object RoomAirSettings:CrossVentilation
provides a simple model, developed by UCSD, for heat transfer
and vertical temperature profile prediction in cross
ventilated rooms. This model was developed using scaling
analysis, experimental correlations, computational fluid
dynamics, and approximate solutions of the Navier-Stokes
equations. The model distinguishes two regions in the room,
the main jet region and the recirculations, and predicts
characteristic airflow velocities and average air
temperatures. The model consists of a set of analytical
expressions that clearly display the first order effects of
room geometry and inflow characteristics on Cross Ventilation
airflow and heat transfer.
The CV model is developed from analysis of the flow
structure inside a cross-ventilated room, which naturally
divides into distinct jet and recirculation flow regions. The
flows are calculated for different room and opening geometries
using CFD. Correlations with room and aperture parameters are
then deduced for a characteristic temperature and velocity in
the jet and recirculation regions.
The SS 2-opening model is developed from analysis of wind
tunnel data for pressure and ventilation rates, correlating
the flow rate with the pressure difference between the two
openings, with contributions from both the mean and unsteady
components. The local velocity estimate is also based on wind
tunnel data, and gives the velocity as a function of wind
angle.
With the widespread availability of air conditioning,
natural ventilation became a rare feature in commercial
buildings in the US. The recent resurgence of the use of
natural ventilation relies on systems whose performance can be
predicted and controlled, often working in conjunction with
mechanical ventilation and cooling systems in hybrid
ventilation configurations. In this contemporary approach
natural ventilation deals with thermal and air renewal
requirements in the cold and mild season and often plays a
limited role in the warm season (due to the need to control
indoor velocities and noise levels). Whereas traditional
natural ventilation systems often use large openings
(1-4m\(^{2}\)), both recent
and expected future examples tend to use smaller openings
(0.2-1m\(^{2}\)), often in
multiple opening configurations that allow for efficient heat
removal with controlled indoor velocities.
The modeling approach used is based on simplified solutions
of the Navier Stokes equations that are integrated into
scaling formulae that predict relevant flow quantities
(characteristic velocities and temperatures) in distinct
regions of the flow. The scaling expressions that will be
developed use the principle of flow similarity: for a given
room geometry and flow regime (either laminar or turbulent),
the flow pattern is constant. In the present case, where the
goal is to predict internal velocities using as input the
inflow velocity (that can be estimated by the airflow rate
predicted by the AirflowNetwork model), the principle can be
stated as:
(Velocity in a given position in the room) = (Inflow
velocity) × (Constant)
The constant is non-dimensional and depends on the position
or room zone where the prediction is being made. For a given
geometry the constant can be measured experimentally or
inferred from a detailed CFD simulation. It is expected that
the constant may depend on non-dimensional room geometry
parameters that can be inferred from a careful analysis of the
flow. The proposed CV model uses a set of CFD simulations, in
conjunction with simplified solutions of the Navier Stokes
equations, to obtain the constants and their dependence on
room geometry features.
The updated CV model proposed in this document was
developed using inlet areas A\(_{in}\) and room cross-section
areas A\(_{RM}\) in
the following range:
0.5% < A\(_{in}\)/A\(_{RM}\) < 5%
whereas the existing model used:
3.2% < A\(_{in}\)/A\(_{RM}\) < 21%
The room lengths for which these models are expected to be
applied range between the smallest room size of 4m and the
maximum achievable CV room flushing length of 15-18m.
Development of an axisymmetric
jet from airflow through a window. [fig:development-of-an-axisymmetric-jet-from]
Air flows into the room through an inflow window in the
form of an approximately axisymmetric jet, with characteristic
diameter L = A\(_{in}\)\(^{1/2}\), that is confined by the
room surfaces (see Figure 6).
The initial portion of the jet (up to six diameters) is called
the potential core and is characterized by shear layer
development along the perimeter of the jet. After this phase
the jet transitions into a self similar profile whose
centerline velocity decays as 1/x’, where x’
= x/D (Awbi, 2003).
For the relevant room length range the geometries used in
the two models result in:
New model: 6 < x’ < 60
Existing model: 2 < x’ < 18
The smaller openings used in the new model result in
significantly more space for jet development along the room
length. In all but the shortest rooms the jet will enter the
fully developed stage characterized by a 1/x’
centerline velocity decay rate. For this reason, the updated
model uses the average jet velocity profile as opposed to
shear layer momentum transfer.
Figure 5
shows the schematic development of a free axisymmetric jet in
an unconfined environment, and Figure 6
depicts the confinement effects. The air entrained by the jet
is detrained at the outlet, creating two recirculation zones
that do not exist in a free jet.
The recirculation flow is similar to the flow in a lid
driven cavity, where here the driving force is the confined
jet flow. Turbulent lid driven cavity flows have been shown to
have self-similar velocity profiles, as shown in Figure 7.
The model is based on the hypothesis that the flow is
composed of:
A – Lid driven cavity flows (in the recirculation
regions).
B – A confined axisymmetric jet flow.
Top view of a cross ventilated
room. The air entrained by the jet is detrained at the outlet,
typically creating two recirculation zones. [fig:top-view-of-a-cross-ventilated-room.-the-air]Top view of one half of a cross
ventilated room. The flow is approximately composed of a
confined jet and two lid driven cavity flows (one on each
side). [fig:top-view-of-one-half-of-a-cross-ventilated]
The goal of the model is to obtain the value and functional
form of the constant term shown in the scaling laws for the
characteristic average velocities in the two flow regions: jet
and recirculation regions.
To estimate the average jet velocity we integrate along the
depth of the room (including both the core region and the
1/x’ velocity decay region) to derive an average
maximum jet velocity, V\(_{J,m}\):∙
(To allow for wind angle effects the room depth D
has been replaced by D* - see a later section.) The
lid-driven cavity flow profile that characterizes the
recirculation region has a maximum velocity that is
approximately one half of the driving velocity, which in this
case is due to the jet. For this reason we define the
separation between the two regions along the virtual surface
where the jet reaches 50% of its maximum, centerline,
velocity.
The analysis of the flow revealed that the room containment
effect leads to higher velocities for cases that combine a
large inflow opening with a small room cross-section area. In
order to account for this effect we propose to use the
following non-dimensional scaling parameter:
where the constant C\(_{1}\) will be obtained from the
results of a set of CFD simulations shown in the next
section.
In addition to the jet velocity the characteristic
recirculation flow velocity \({U_R}\) is also an interesting
parameter that, according to the self-similar flow profile
hypothesis shown in Figure 7,
should scale with a similar expression to the one shown above
for the jet (although with a different correlation constant,
C\(_{2}\)). The
modeling hypothesis can also be tested by evaluating its
capability to predict the maximum recirculation flow rate,
which can be predicted by multiplying the average velocity by
the room cross-section area:
Because of the self-similar velocity profile shape that
occurs in the recirculation, the near wall velocity (used to
predict forced convection) can be estimated simply by
multiplying the characteristic recirculation velocity by
two.
The expressions obtained apply in the turbulent regime
(because the CFD simulations on which they are based are for
that regime). Both velocity correlations presented below have
the functional form:
\[V = aX + b\]
where a is C\(_{1}\), C\(_{2}\) or C\(_{3}\), X is a velocity scale and
b is a constant that results from the best fit. The
model has a lower limit on the value of X: if
X is too small the flow is no longer turbulent. This
limit implies that the point X = 0 is never
achieved, avoiding the unrealistic prediction of V(0)
= b.
In addition to the velocity and airflow rate predictions
described above it is also essential to predict the
temperature variations in the two zones of the flow. In most
cases these variations will be positive due to the internal
sensible heat gains. However’, as a result of room surface
heat transfer, in some cases the variations may be negative.
Analysis of the CFD simulations that are presented in the next
section revealed that approximately all the exhaust air comes
from the jet regions. For this reason, the heat gains affect
the two regions differently:
Heat gains in the jet region only affect this region
(no temperature increase in the recirculation
regions).
Heat gains in the recirculations affect both jet and
recirculation regions (because they must be exhausted by the
jets).
In order to predict the temperature increase, heat gains
were inserted into the recirculation regions for different
room geometries in the set of simulations presented in the
next section. The volume containing the heat gains extended
over the whole depth of the room, from floor to ceiling, and
in the lateral direction from halfway between the window edge
and the wall all the way to the wall. Applying energy
conservation principles to predict the temperature increase in
the jet and recirculation region as a function of the sensible
heat gains in each region (the sum of internal gains and
internal surface heat transfer) we can obtain the average
value of temperature increase in each region:
where q”\(_{RM}\) is the
total internal heat gain for the room, \(\rho\) and c\(_{p}\) are the density and
specific heat capacity of air, respectively, and
C\(_{T,J}\),
C\(_{T,R}\) are
constants.
Table 2
lists the cases used to develop the updated CV model, which
predicts the output variables defined in Table 3
(see also Figure 8).
Figures 9, 10,
and 11
show plots of the data derived from the CFD simulations of
these test cases, and the lines giving the best fit to the
data. Table 4
summarizes the correlation formulae and constants. Finally,
Table 5
provides estimates for the error of the correlation in each
test case.
Jet and recirculation regions
in typical vertical cross-section through room (y-z plane).
Jet boundary occurs where jet velocity has fallen to 50% of
centerline maximum. Remainder of cross-section is treated as
recirculation. Volumetric heat sources ar [fig:jet-and-recirculation-regions-in-typical]
Test cases used to develop correlations
Case
Opening area
Room width
Room depth
Room height
A (m\(^2\))
W (m)
D (m)
H (m)
1
0.25
6
9
2.3
2
0.25
11
9
2.3
3
0.5
9
4.5
2.3
4
0.5
9
6
2.3
5
0.5
6
9
2.3
6
0.5
9
9
2.3
7
0.5
11
9
2.3
8
0.5
13.5
9
2.3
9
0.5
9
13.5
2.3
10
0.5
9
18
2.3
11
0.5
13.5
18
2.3
12
1
9
9
2.3
13
1
9
13.5
2.3
14
1
9
18
2.3
RANGE (max:min ratio)
4
2.25
4
1
Definitions of output variables
Output variable
Symbol
Units
Definition
Jet velocity
U\(_J\)
m/s
Volume-averaged jet region
velocity. The averaging volume is bounded at each x along the
room depth by the line in the y-z plane where the jet velocity
drops to 50% of its maximum centerline value.
Recirculation zone velocity
U\(_R\)
m/s
Area-averaged velocity in the
y-z plane with maximum flow. The averaging area is the
recirculation part of the room cross-section. Typically the
plane of maximum flow occurs at x \(\sim\) 2D/3 (D/3 before the
outlet).
Recirculation zone flow
rate
Q\(_R\)
m\(^3\)/s
Total flow rate for the
recirculation regions in the plane of maximum flow (see
above).
Jet temperature rise
\(\Delta\)T\(_J\)
°C
Volume-averaged temperature
variations in the jet region, over the same volume used to
define the jet velocity average.
Recirculation zone temperature
rise
\(\Delta\)T\(_R\)
°C
Volume-average temperature
variations in the recirculation region. The average is
calculated over the cuboidal volume placed in each
recirculation containing the volumetric sensible heat gains
(see main text).
See also Figure 8.
Note the x-coordinate is along the room, the
y-coordinate is across the room and the
z-coordinate is vertical.
Accuracy of correlation formulae in predicting test
case results
Cases
Errors (%)
ID
A (m2)
W (m)
D (m)
H (m)
Jet Velocity
Recirc- ulation Velocity
Recirc- ulation Flow Rate
Jet Temperature Rise
Recirc- ulation Temperature Rise
1
0.25
6
9
2.3
16
45
60
1
10
2
0.25
11
9
2.3
5
43
32
5
7
3
0.5
9
4.5
2.3
2
18
27
15
2
4
0.5
9
6
2.3
6
32
28
7
1
5
0.5
6
9
2.3
7
6
21
7
0
6
0.5
9
9
2.3
10
4
6
15
17
7
0.5
11
9
2.3
6
7
0
14
20
8
0.5
13.5
9
2.3
14
9
5
21
2
9
0.5
9
13.5
2.3
10
11
17
7
8
10
0.5
9
18
2.3
10
25
29
8
3
11
0.5
13.5
18
2.3
8
2
23
9
6
12
1
9
9
2.3
2
5
3
3
4
13
1
9
13.5
2.3
8
10
10
3
6
14
1
9
18
2.3
0
18
6
13
2
Average
8
17
19
9
6
In each case the error is the percentage difference between
the given correlation prediction and the test case result
obtained from CFD, i.e. 100*(correlation prediction –
CFD)/CFD.
Jet velocity correlation. [fig:jet-veolicty-correlation]Recirculation region velocity
correlation. [fig:recirculation-region-velocity]Recirculation region flow rate.
[fig:recirculation-region-flow-rate]
CFD simulations for rooms with more than one inflow
aperture show that the model can be applied to these
configurations with minimal adjustments. As a result of
increased dissipation in the central recirculation region,
where two distinct recirculation flows collide (Figure 13),
the airflow velocities are reduced by 5-10%, as shown in the
charts in Figure 14.
Flow regions for a room with 2
inlets and 2 outlets. [fig:flow-regions-for-a-room-with-2-inlets-and-2]Jet velocity and recirculation
velocity for the two halves of 3 different 2-inlet rooms, A-C,
compared with the corresponding 1-inlet rooms. [fig:jet-velocity-and-recirculation-velocity-for]
Heat transfer in recirculating
flows
In most cross ventilation flows with recirculations the
mixing between recirculation and inflow jet is only partial.
Figure 15
shows a schematic representation of the heat transfer process
(showing a top view of one half of a symmetrical room). The
flow is divided into three distinct streams with connected
temperature variations:
The main jet (labeled J in the figure).
The part of the recirculation flow that exchanges heat
with the jet (label R).
The wall boundary layer part of the recirculation flow
(label W).
Top view of the flow structure
in case R. [fig:top-view-of-the-flow-structure-in-case-r.]
The light gray arrows show flow direction. The dark gray
arrows show heat transfer in the shear layer. The
recirculation region coordinate system is shown in the figure,
with coordinate r varying between 0 and
L.
It is possible to use a formulation with a heat transfer
area, a convection heat transfer coefficient and an average
shear layer temperature difference to model the heat exchange
between jet and recirculation flows:
Predicting
air temperature in the recirculation region of the flow[LINK]
An accurate representation of system behavior must consider
both the local heat transfer in the shear layer (varying along
the x direction, driven by the temperature profiles
T\(_{J}\)(x) and
T\(_{R}\)(x)) and
the heat transfer in the wall boundary layers (with
temperature T\(_{W}\)(r)). Both transfers
can be modeled using differential equations, resulting in the
following system of equations:
The first equation in Equation [eq:DiffEqForTWrTRxTJx]
models heat transfer in the wall layer. The second models the
temperature variation of the portion of the recirculation flow
that is in contact with the jet, it differs from the third
equation only in a sign (the temperature variation in
T\(_{R}\) is
symmetric to the variation in T\(_{J}\)) and the flow rate
(F\(\cdot\)R for the
recirculation and F for the main jet). The
temperature in the recirculation region is given by:
Heat
transfer in recirculating flows with internal gains[LINK]
Whereas surface heat transfer is an important component of
room heat transfer, internal gains often dominate the room
temperature field. Referring to the previously used
subdivision of the room into jet and recirculation regions we
can see that gains can occur exclusively in the jet region,
exclusively in the recirculation region or in both
regions.
The effect of gains in the jet region on room air
temperature distribution is simple to model and, within the
first order accuracy goal, can be characterized as
follows:
For gains occurring in the jet region (between the inlet
and the room outlet) a change in inflow temperature is an
adequate, conservative approximation. From energy conservation
we conclude that the altered inflow temperature is given
by:
Gains occurring in the jet region, close to the outlet can
be ignored in a first order accuracy model.
Clearly the two approaches described above are only exact
when: the gains occur at the inlet and perfectly mix with the
inflow jet (for the first case), or whenever the gains occur
very close to the outlet (in the second case). In all other
situations, the first approach provides a conservative
approximation. In the model implementation in EnergyPlus the
user will not define the exact location of the heat gains
within each region of the flow therefore, heat gains in the
jet region, are inserted in two equal parts in two separate
points of the jet flow (in the room entrance and before the
exit, see points r = 0 and r = L\(_{R}\) in Figure 15).
When heat gains occur in the recirculation region
significant heat accumulation occurs due to the limited heat
transfer ability of the shear layer. Due to this limited
ability, the recirculation is partially isolated from the main
jet flow and higher temperatures are generated in this region
whenever heat gains are present. The concept of a global heat
transfer coefficient is not applicable.
In the case of a flow in a room with adiabatic surfaces and
heat gains in the recirculation region the indoor air
temperatures can be predicted using the solution to the
following system of equations:
\[\begin{array}{rl}
\rho C_p F R \left(T_R\left(0\right) -
T_R\left(L\right)\right) & = G_R \\
\rho C_p F R \frac{\partial T_R \left(x\right)}{\partial
x} & = \frac{A_{SL} h_{SL}}{L} \left(T_J\left(x\right) -
T_R \left(x\right)\right) \\
\rho C_p F \frac{\partial T_J \left(x\right)}{\partial x}
& = \frac{A_{SL} h_{SL}}{L} \left(T_R\left(x\right) - T_J
\left(x\right)\right)
\end{array}
\label{eq:TRoTRLdTRxdTJxSystem}\]
This system of equations differs from the system shown in
Equation [eq:DiffEqForTWrTRxTJx]
in several ways. Because there is no wall heat transfer there
is no need to split the recirculation flow in two parts
(R and W). The temperature increase in the
recirculation flow is simply defined by the internal gains
(first equation in Equation [eq:TRoTRLdTRxdTJxSystem]).
The second and third equations in Equation [eq:TRoTRLdTRxdTJxSystem]
model shear layer heat transfer and are similar in the two
cases (Equation [eq:DiffEqForTWrTRxTJx]
and Equation [eq:TRoTRLdTRxdTJxSystem]).
The recirculation flow temperature at x = 0 is
given by:
The temperature in the recirculation is inversely
proportional to the recirculation flow ratio (R) and
the shear layer area (measured by the product: A\(_{SL}\) h\(_{SL}\)). As both of these
parameters increase this expression becomes similar to
Equation [eq:AlteredInflowTempEquation].
The recirculation flow temperature at x = L is given
by:
\[{T_R}(L) = {T_R}(0) + \Delta
{T_R}\]
The temperature variations in the recirculation flow is
given by:
With the boundary conditions: \({T_W}(L) = {T_R}(0)\,,\,{T_W}(0) =
{T_R}(L)\,,\,\,{T_J}(0) = {T_{IN}}\)
For simplicity the heat gains are considered to be evenly
distributed along the recirculation path. The temperature
variation in the recirculation region is given by:
The RoomAirflowNetwork integrates the RoomAir model with
the AirflowNetwork model. The model allows multiple Room Air
nodes to be associated with AirflowNetwork intra zone nodes.
The AirflowNetwork intra zone linkages provide links among the
intra zone nodes and users may specify airflow components
among these links. The AirflowNetwork model calculates
airflows of the links. The incoming airflows are part of the
zone heat and moisture balances of intra zone nodes. The model
also allows users to specify a node to be connected to
surfaces to have convective heat transfer and/or moisture
transfer between surfaces and the node, portion of internal
gains and supply air and return air fractions from zone
equipment and AirLoop terminals.
The input object RoomAirSettings:AirflowNetwork
lists multiple RoomAirflowNetwork nodes in a zone. The input
object of RoomAirflowNetwork:Node defines a list of
connections assigned to a particular RoomAirflowNetwork node
so that the connections will be used as components of node
heat and moisture balance equations. In addition, the input
objects AirflowNetwork:IntraZone:Node
and AirflowNetwork:IntraZone:Linkage
specify airflow nodes, linkages, and associated flow
components, so that the AirflowNetwork model is used to
calculate intrazone linkage airflows, which will be a part of
loads for the RoomAir node balances.
The model calculates each component of the governing
equations first. Then heat and moisture balance equations are
solved to get the node temperature and humidity ratio at the
current time step.
Governing equations for each RoomAirflowNetwork node are
described below:
T\(_{z,i}\) = Air
temperature at the ith zone RoomAir node
N\(_{i,g}\) = The number
of internal gains for the ith node, defined in the
RoomAir:Node:Airflownetwork:InternalGains object
F\(_{gain,j}\) = Fraction
of internal gain from jth internal gain object defined in the
RoomAir:Node:Airflownetwork:InternalGains object
\({\dot Q}_{i,j}\) = The
amount of internal sensible gain at the jth internal gain
object and ith node.
\({\sum\limits_{j = 1}^{N_{i,g}}
F_{gain,j}{\dot Q}_{i,j}}\) = The total amount of
internal sensible gains assigned to the ith node
N\(_{i,sur}\) = The number
of adjacent surface with convective heat transfer for the ith
node, defined in the
RoomAir:Node:Airflownetwork:AdjacentSurfaceList object
h\(_{i,j}\) = Heat
transfer coefficient between the jth adjacent surface and ith
node
A\(_{j}\) = Area of the
jth adjacent surface
T\(_{s,j}\) = Inside
temperature at the jth adjacent surface
\({\sum\limits_{j =
1}^{N_{i,sur}} h_{i,j} A_j \left( T_{s,j} -
T_{z,i}\right)}\) = The total amount of convective
gains from adjacent surfaces attached to the ith node
N\(_{i,AFN}\) = The number
of RoomAir:Node:Airflownetwork objects connected to the ith
node. The links are defined in AirflowNetwork:IntraZone:Linkage
objects.
\({\dot m}_{AFN,i,j}\) =
Mass flow rate from the jth node to the ith node. The flow
rates are calculated in the AirflowNetwork model. It should be
pointed out that the outgoing flows from the ith node will not
be used in the present energy and moisture balance
calculations, because they are a part of energy and moisture
balances.
T\(_{z,j}\) = Air
temperature at the jth zone RoomAir node
\({\sum\limits_{j = 1}^{N_{i,g}}
\dot m_{AFN i,j} {C_p}\left( T_{z,j} - T_{z,i}
\right)}\) = The total amount of heat transfer driven
by intra zonal flows
N\(_{i,HVAC}\) = The
number of RoomAir:Node:Airflownetwork:HVACEquipment objects
which provide supply air to the ith node.
F\(_{HVAC,j}\) = Fraction
of output or supply air to the ith node from jth HVAC
equipment defined in RoomAir:Node:Airflownetwork:HVACEquipment
objects
\({\dot m}_{sup,i,j}\) =
The supply mass flow rate from the jth HVAC equipment
T\(_{sup,j}\) = Supply air
temperature at the jth HVAC equipment
\({\sum\limits_{j =
1}^{N_{i,HVAC}} F_{HVAC,j} \dot m_{sup,i,j} \left( T_{sup,j} -
T_{z,i}\right)}\) = The total amount of heat transfer
from HVAC equipment to the ith node
W\(_{z,i}\) = Air humidity
ratio at the ith zone RoomAir node
N\(_{i,g}\) = The number
of internal gains for the ith node, defined in the
RoomAir:Node:Airflownetwork:InternalGains object
F\(_{gain,j}\) = Fraction
of internal gain from jth internal gain object defined in the
RoomAir:Node:Airflownetwork:InternalGains object
\({\dot Q}_{i,j}\) = The
amount of internal latent gain at the jth internal gain object
and ith node.
\({\sum\limits_{j = 1}^{N_{i,g}}
F_{gain,j}{\dot Q}_{i,j}}\) = The total amount of
internal latent gains attached to the ith node
N\(_{i,sur}\) = The number
of adjacent surface with convective heat transfer for the ith
node, defined in the
RoomAir:Node:Airflownetwork:AdjacentSurfaceList object
h\(_{M,j}\) = Mass
transfer coefficient between the jth adjacent surface and ith
node
A\(_{j}\) = Area of the
jth adjacent surface
W\(_{s,j}\) = Inside
humidity ratio at the jth adjacent surface
\({\sum\limits_{j =
1}^{N_{i,sur}} h_{M,j} A_j \left( W_{s,j} -
W_{z,i}\right)}\) = The total amount of convective
moisture gains from adjacent surfaces attached to the ith
node
N\(_{i,AFN}\) = The number
of RoomAir:Node:Airflownetwork objects connected to the ith
node. The links are defined in AirflowNetwork:IntraZone:Linkage
objects.
\({\dot m}_{AFN,i,j}\) =
Mass flow rate from the jth node to the ith node. The flow
rates are calculated in the AirflowNetwork model. It should be
pointed out that the outgoing flow from the ith node will not
be used in the present calculation, because they are a part of
energy balance.
W\(_{z,j}\) = Air humidity
ratio at the jth zone RoomAir node
\({\sum\limits_{j =
1}^{N_{i,AFN}} \dot m_{AFN,i,j} \left( W_{z,j} -
W_{z,i}\right)}\) = The total amount of moisture
transfer driven by intra zonal flows
N\(_{i,HVAC}\) = The
number of RoomAir:Node:Airflownetwork:HVACEquipment objects
which provide supply air to the ith node.
F\(_{HVAC,j}\) = Fraction
of output or supply air to the ith node from jth HVAC
equipment defined in RoomAir:Node:Airflownetwork:HVACEquipment
objects
\({\dot m}_{sup,i,j}\) =
The supply mass flow rate from the jth HVAC equipment
W\(_{sup,j}\) = Supply air
humidity ratio at the jth HVAC equipment
\({\sum\limits_{j =
1}^{N_{i,HVAC}} F_{HVAC,j} \dot m_{sup,i,j} \left( W_{sup,j} -
W_{z,i}\right)}\) = The total amount of moisture
transfer from HVAC equipment to the ith node
The AirflowNetwork model is called first to calculate
intrazone linkage air flows based on node temperatures and
humidity ratios at the previous time step. The linkage loads
are summarized based on each node with all incoming airflow,
while any outgoing airflows from the node of the interest are
excluded in the node energy and moisture balances. When the
RoomAirModelAirflowNetowrk module is called, it summarizes
internal gains at each node based on specification provided in
the RoomAirflowNetwork:Node:InternalGains object, convective
heat transfer from adjacent walls specified in the
RoomAirflowNetwork:Node:AdjacentSurfaceList object. These load
components are used in the predictor to predict system loads
based on thermostat setpoint and supply fraction in the
controlled zone node. In the corrector, the supply air
conditions defined in the
RoomAirflowNetwork:Node:HVACEquipment object are added. After
all load components are assembled together, the node energy
and moisture balance equations are solved. The node
temperature and humidity ratio are calculated at the current
time step.
Baturin, V. V. , Billington, N. S. 1972. Fundamentals of
Industrial Ventilation, Franklin Book Company 1972,
pp. 174-179.
Bejan, A. 1994. Convection Heat Transfer 2nd ed, Wiley,
USA.
Carrilho da Graça, G. 2003. Simplified models for heat
transfer in rooms. Ph. D. Thesis, University of California,
San Diego.
Chandra, S., Kerestecioglu, A.A. 1984. Heat transfer in
naturally ventilated rooms data from full-scale measurements,
ASHRAE Transactions, Volume 90, part 1b 211-224.
Chen, Q., and B. Griffith. 2002. Incorporating Nodal Room
Air Models into Building
Energy Calculation Procedures. ASHRAE RP-1222 Final
Report.
Cooper, P. and P.F. Linden. 1996. Natural ventilation of an
enclosure containing two buoyancy sources. Journal of Fluid
Mechanics, Vol. 311, pp. 153-176.
Dominique, M. and P. Guitton. 1997. Validation of
displacement ventilation simplified models. Proc. of Building
Simulation.
Griffith, B. and Q. Chen. 2004. Framework for coupling room
air models to heat balance load and energy calculations
(RP-1222). International Journal of Heating, Ventilating,
Air-conditioning and Refrigerating Research. ASHRAE, Atlanta
GA. Vol 10. No 2. April 2004.
Holford, J.M., G.R. Hunt and P.F. Linden. 2002. Competition
between heat sources in a ventilated space. Proceedings of
RoomVent 2002, pp. 577-580.
Hunt, G.R., J.M. Holford and P.F. Linden. 2002.
Characterization of the flow driven by a finite area heat
source in a ventilated enclosure. Proceedings of RoomVent
2002, pp. 581-584.
Hunt, G.R. and P.F. Linden. 2001. Steady-state flows in an
enclosure ventilated by buoyancy forces assisted by wind. .
Journal of Fluid Mechanics, Vol. 426, pp. 355-386.
Kalema, T., Haapala, T. ’995. Effect of interior heat
transfer coefficients on thermal dynamics and energy
consumption, Energy and Buildings 22 (2) pp. 101-113.
Kaye, K.N. and P.F. Linden. 2004. Coalescing axisymmetric
turbulent plumes. Journal of Fluid Mechanics, Vol.
502, pp. 41–63.
Lin, Y.J. and P.F. Linden. 2005. A model for an under floor
air distribution system. Energy&Building, Vol. 37,
pp. 399-409.
Linden, P.F., G.F. Lane-Serff and D.A. Smeed. 1990.
Emptying filling boxes: the fluid mechanics of natural
ventilation. Journal of Fluid Mechanics, Vol. 212,
pp. 309-335.
Linden, P.F. and P. Cooper. 1996. Multiple sources of
buoyancy in a naturally ventilated enclosure. Journal of Fluid
Mechanics, Vol. 311, pp. 177-192.
Liu, Q. 2006. The Fluid Dynamics of an Underfloor Air
Distribution System. A PhD dissertation in Engineering
Sciences (Systems Science) at UC San Diego.
Morton, B.R., G.I. Taylor andJ.S. Turner. 1956. Turbulent
gravitational convection from maintained and instantaneous
sources. Proceedings of the Royal Society of London, Vol A234,
pp. 1-23.
Mundt, E. 1996. The Performance of Displacement Ventilation
Systems – Experimental and Theoretical Studies, Ph. D. Thesis,
Bulletin N38, Building
Services Engineering KTH, Stockholm.
Neiswanger, L., Johnson, G.A., Carey, V.P. 1987. An
experimental study of high Raleigh number mixed convection in
a rectangular enclosure with restricted inlet and outlet
openings. Transactions of ASME, Vol. 109, 446-453.
Ohba, M., Irie, K., Kurabuchi, T. 2001. Study on airflow
characteristics inside and outside a CVmodel, and ventilation
flow rate using wind tunnel experiments. Journal of Wind
Engineering and Industrial Aerodynamics, in press.
Rees, S.J., and P. Haves. 2001. A nodal model for
displacement ventilation and chilled ceiling systems in office
spaces. Building
and Environment, Vol. 26, pp. 753-762.
Siebers, D. L., Schwind, R. G., Moffat, R. J. 1983.
Experimental Mixed Convection Heat Transfer From a Large
Vertical Surface in a Horizontal Flow. SAND 83-8225, Sandia
National Laboratories, Livermore CA.
Spitler, J. D., Pedersen, C.O., Fisher, D.E. 1991. Interior
Convective Heat Transfer in Buildings with Large Ventilative
Flow Rates. ASHRAE Transactions, Vol. 97, Pt.1,
pp. 505-515.
RoomAir Models[LINK]
The group of models described in this section is used to account for non-uniform room air temperatures that may occur within the interior air volume of a zone. These models are accessed using the RoomAirModelType input object. RoomAir modeling was added to EnergyPlus starting with Version 1.2. Although there are many types of analyses (comfort, indoor air quality, etc) that might benefit from localized modeling of how room air varies across space, most of the room air models in EnergyPlus only address the distribution of temperature within the zone. This allows surface heat transfer and air system heat balance calculations to be made taking into account natural thermal stratification of air and different types of intentional air distribution designs such as under-floor and side-wall displacement ventilation that purport to extract room air at higher-than-mean temperatures. The exception is the RoomAirflowNetwork model, which integrates the AirflowNetwork model and applies the nodal airflow model within zones. Note that EnergyPlus does not have completely general methods of modeling room air that are applicable to every conceivable type of airflow that might occur in a zone. Such models (e.g. RANS-CFD) are too computationally expensive to use with EnergyPlus for the foreseeable future. The models that are available in EnergyPlus offer only limited modeling capabilities for select room airflow configurations. Also note that because the complete mixing model for room air has long been the standard in building energy simulation, there is not currently a consensus on how to best model non-uniform air temperatures in buildings. Therefore, it is up to the user to have a good understanding of when, where, and how to apply the room air models available in EnergyPlus. The rest of this section provides some guidance in the way of examples and further discussion of the models available in EnergyPlus.
EnergyPlus offers the different types of air models listed in the table below along with the input objects associated with the use of that model.
The room air models are coupled to the heat balance routines using the framework described by Griffith and Chen (2004). Their framework was modified to include features needed for a comprehensive program for annual energy modeling rather than one for hourly load calculations. The formulation is largely shifted from being based on the setpoint temperature to one based on the current mean air temperature. This is necessary to allow for floating temperatures and dual setpoint control where there may be times that the mean zone temperatures are inside the dead band. The coupling framework was also extended to allow for exhaust air flows (e.g. bathroom exhaust fans) in addition to air system return flows.
The inside face temperature calculation is modified by rewriting the zone air temperature, T\(_{a}\), with an additional subscript, i, for the surface index (\({T_{{a_j}}} \to {T_{{a_{i,j}}}}\) or \({T_{{a_{}}}} \to {T_{{a_i}}}\) ). The inside face heat balance is solved for its surface temperature using,
\[{T_{{s_{i,j}}}} = \frac{{{T_{s{o_{i,j}}}}{Y_{i,o}} + \sum\limits_{k = 1}^{nz} {{T_{s{o_{i,j - k}}}}} {Y_{i,k}} - \sum\limits_{k = 1}^{nz} {{T_{{s_{i,j - k}}}}} {Z_{i,k}} + \sum\limits_{k = 1}^{nq} {{\Phi_{i,k}}{{q''}_{k{i_{i,j - k}}}} + {T_{{a_{i,j}}}}{h_{{c_{i,j}}}} + {{q''}_{LWS}} + {{q''}_{LWX}} + {{q''}_{SW}} + {{q''}_{sol}}} }}{{{Z_{i,o}} + {h_{{c_{i,j}}}}}}\]
where:
T\(_{s}\) is the inside face temperature
i is the subscript indicates individual surfaces
j is the subscript indicates current time step
k is the subscript indicates time history steps
T\(_{so}\) is the outside face temperature
Y\(_{i}\) are the cross CTF coefficients
Z\(_{i}\) are the inside CTF coefficients
\(\varphi_{i}\) are the flux CTF coefficients
\({q''_{ki}}\) is the conduction heat flux through the surface
\({h_{{c_i}}}\) is the surface convection heat transfer coefficient
T\(_{a}\) is the near-surface air temperature
\({q''_{LWS}}\) is the longwave radiation heat flux from equipment in zone
\({q''_{LWX}}\) is the net long wavelength radiation flux exchange between zone surfaces
\({q''_{SW}}\) is the net short wavelength radiation flux to surface from lights
\({q''_{sol}}\) is the absorbed direct and diffuse solar (short wavelength) radiation.
User Defined RoomAir Temperatures[LINK]
The input object RoomAir:TemperaturePattern:UserDefined provides a capabity for users to define the sort of air temperature pattern he or she expects in the zone. With these models, the pattern is generally set beforehand and does not respond to conditions that evolve during the simulation. (Exception: the pattern available through the RoomAir:TemperaturePattern:TwoGradient object will switch between two different pre-defined vertical gradients depending on the current value of certain temperatures or thermal loads. )
The user-defined patterns obtain the mean air temperature, \({T_{MAT}}\), from the heat balance domain and then produce modified values for:
\({T_{ai}}\) is the adjacent air temperature which is then used in the calculation of inside face surface temperature during the heat balance calculations.
\({T_{leaving}}\) is the temperature of air leaving the zone and entering the air system returns.
\({T_{exhaust}}\) is the temperature of air leaving the zone and entering the exhaust.
\({T_{stat}}\) is the temperature of air “sensed” at the thermostat (not currently used in air system control because air system flows use load-based control).
The user defined room air models used indirect coupling so that the patterns provide values for, or ways to calculate, how specific temperatures differ from \({T_{MAT}}\). The various \(\Delta T\) values determined from the model are applied to \({T_{MAT}}\) as follows:
\[{T_{ai}} = \Delta {T_{ai}} + {T_{MAT}}\]
(where “i’s” represents each surface in the zone that is affected by the model)
\[{T_{leaving}} = \Delta {T_{return}} + {T_{MAT}}\]
\[{T_{exhaust}} = \Delta {T_{exhaust}} + {T_{MAT}}\]
\[{T_{stat}} = \Delta {T_{stat}} + {T_{MAT}}\]
The patterns defined by the object ‘RoomAir:TemperaturePattern:SurfaceMapping’ are fairly straightforward. The user directly inputs values for \(\Delta {T_{ai}}\) for each surface. The pattern “maps” specific surfaces, identified by name, to \(\Delta {T_{ai}}\) values. This provides completely general control (but in practice may be cumbersome to use). The other patterns focus on temperature changes in the vertical direction. Surfaces do not need to be identified, but all the surfaces with the same height will be assigned the same \(\Delta {T_{ai}}\) values.
The patterns defined by the object ‘RoomAir:TemperaturePattern:NondimensonalHeight’ apply a temperature profile based on a non-dimensionalized height, \({\rm Z}\). The height of each surface is defined to be the z-coordinate of the surface’s centroid relative to the average z-coordinate of the floor surfaces. The zone ceiling height is used as the length scale to non-dimensionalize each surface’s height so that,
\[{{\rm Z}_i} = ({z_{i,centroid}} - {z_{floorAvg}})/Ceiling\;Height\]
(where “i’s” represents each surface in the zone that is affected by the model)
The values for \({Z_i}\) are constrained to be between 0.01 and 0.99 because the value is meant to describe the air layer near the surface (say approximate 0.1 m from the surface) rather than the surface itself.
The user-defined profile is treated as a look up table or piecewise linear model. The values for \(\Delta {T_{ai}}\) are determined by searching the \({\rm Z}\) values in the user-defined profile and performing linear interpolation on the associated \(\Delta {T_a}\) values.
The patterns defined by the object ‘RoomAir:TemperaturePattern:ConstantGradient’ apply a constant temperature gradient in the vertical direction. The model assumes that \({T_{MAT}}\) occurs at the mid-plane so that \({{\rm Z}_{{T_{MAT}}}} = 0.5\) (by definition). The surface \({Z_i}\) values are compared to \({Z_{{T_{MAT}}}}\) and then scaled with zone ceiling height to obtain values for the change in height (in units of meters), \(\Delta z\). The user defined gradient, \(grad\), (units of \(^{\circ}\)C/m) is then used to determine \(\Delta {T_{ai}}\) values using:
\[\Delta {T_{ai}} = \Delta z * grad\]
The patterns defined by the object ‘RoomAir:TemperaturePattern:TwoGradient’ are very similar to the constant gradient pattern above but the value of \(grad\) used at any given time is selected by interpolating between two user-defined values for \(grad\). Five options are available, three based on temperatures and two based on thermal loads – see the Input Output Reference. The user provides upper and lower bounding values. If the current value of the “sensing” variable lies between the upper and lower bounds, then \(grad\) is determined using linear interpolation. If the designated value is above the upper bound then the upper value for \(grad\) is used (no extrapolation). Similarly, if the designated value is below the lower bound, then the lower value for \(grad\) is used. Note that “upper” and “lower” indicate the temperature and heat rate bounds and that the values for \(grad\) do not have to follow in the same way; the \(grad\) value for the lower bound could be higher than the \(grad\) value for the upper bound (providing a something of a reverse control scheme). Rather than directly using \(\Delta T\) values from the user, the temperatures for the return air, exhaust and thermostat are determined based on user-entered heights (in units of meters from the floor) and applying the current value for \(grad\).
One-Node Displacement Ventilation RoomAir Model[LINK]
The input object RoomAirSettings:OneNodeDisplacementVentilation provides a simple model for displacement ventilation. Mundt (1996) points out that a floor air heat balance provides a simple and reasonably accurate method of modeling the temperature near the floor surface. The slope of a linear temperature gradient can then be obtained by adding a second upper air temperature value that comes from the usual overall air system cooling load heat balance. The figure below diagrams the temperature distribution versus height being calculated by the model. Mundt’s floor air heat balance is extended to include convection heat gain from equipment and by ventilation or infiltration that may be introduced near the floor in order to maintain all the terms in the air heat balance of the Heat Balance Model. This yields the following heat balance for a floor air node,
\[\rho \,{c_p}\dot V({T_{AirFloor}} - {T_{Supply}}) = {h_{cFloor}}{A_{Floor}}({T_{Floor}} - {T_{AirFloor}}) + {Q_{ConvSourceFloor}} + {Q_{InfilFloor}}\]
where:
\(\rho\) is the air density
c\(_{p}\) is the air specific heat at constant pressure
\(\mathop V\limits^\cdot\) is the air system flow rate
T\(_{supply}\) is the air system’s supply air drybulb temperature
h\(_{cFloor}\) is the convection heat transfer coefficient for the floor
A\(_{floor}\) is the surface area of the floor
T\(_{floor}\) is the surface temperature of the floor
Q\(_{convSourceFloor}\) is the convection from internal sources near the floor (< 0.2 m)
Q\(_{InfilFloor}\) is the heat gain (or loss) from infiltration or ventilation near the floor.
“Floor splits” are the fraction of total convective or infiltration loads that are dispersed so as to add heat to the air located near the floor. The user prescribes values for floor splits as input. No guidance is known to be available to use in recommending floor splits, but the user could for example account for equipment known to be near the floor, such as tower computer cases, or supplementary ventilation designed to enter along the floor. The equation above can be solved directly for T\(_{AirFloor}\) and is used in the form of the equation below,
\[{T_{AirFloor}} = \frac{{\rho \,{c_p}\dot V{T_{Supply}} + \sum {{h_{cFloor}}{A_{Floor}}{T_{Floor}}} + {{\dot Q}_{ConvSourceFloor}} + {{\dot Q}_{InfilFloor}}}}{{\rho \,{c_p}\dot V + \sum {{h_{cFloor}}{A_{Floor}}} }}\]
The upper air node temperature is obtained by solving the overall air heat balance for the entire thermal zone for the temperature of the air leaving the zone and going into the air system return, T\(_{leaving}\).
\[{T_{Leaving}} = \frac{{ - {{\dot Q}_{sys}}}}{{p{c_p}\dot V}} + {T_{Supply}}\]
where \({\dot Q_{sys}}\) is the air system heat load with negative values indicating a positive cooling load. Values for \({\dot Q_{sys}}\) are computed by the load calculation routines and passed to the air model. The vertical temperature gradient or slope, dT/dz, is obtained from,
\[\frac{{dT}}{{dz}} = \frac{{{T_{Leaving}} - {T_{AirFloor}}}}{{{H_{return}}}}\]
where H\(_{return}\) is the distance between the air system return and the floor air node assumed to be 0.1 m from the floor and z is the vertical distance.
The constant slope allows obtaining temperatures at any vertical location using,
\[{T_{{a_i}}} = {T_{leaving}} - \frac{{dT}}{{dz}}({z_{leaving}} - {z_i})\]
So for example the temperatures near the ceiling can easily be determined. Accounting for the location of the thermostat inside the zone (e.g. 1.1 m) is accomplished by returning the temperature for the appropriate height to the appropriate air node used for control. If the walls are subdivided in the vertical direction as shown in the figure above, then the air model can provide individual values for each surface based on the height and slope. However, no additional heat balances are necessarily made (in the air domain) at these points as all the surface convection is passed to the model in the totaled value for \({\dot Q_{sys}}\).
Three-Node Displacement Ventilation RoomAir Model[LINK]
Overview[LINK]
The input object RoomAirSettings:ThreeNodeDisplacementVentilation provides a simple model for heat transfer and vertical temperature profile prediction in displacement ventilation. The fully-mixed room air approximation that is currently used in most whole building analysis tools is extended to a three node approach, with the purpose of obtaining a first order precision model for vertical temperature profiles in displacement ventilation systems. The use of three nodes allows for greatly improved prediction of thermal comfort and overall building energy performance in low energy cooling strategies that make use of unmixed stratified ventilation flows.
The UCSD Displacement Ventilation Model is one of the non-uniform zone models provided through the Room Air Manager in EnergyPlus. The intent is to provide a selection of useful non-uniform zone air models to enable the evaluation of air-conditioning techniques that use stratified or partially stratified room air. Such techniques include displacement ventilation (DV) and underfloor air distribution (UFAD) systems. The methodology can also include, in principle, natural displacement ventilation and also wind-driven cross-ventilation (CV).
Displacement Ventilation[LINK]
A DV system is a complete contrast to a conventional forced air system. In a conventional system conditioned air is delivered at ceiling level and the intent is to create a fully mixed space with uniform conditions. In a DV system conditioned air is delivered at floor level and low velocity in order to minimize mixing and to establish a vertical temperature gradient. The incoming air “displaces” the air above it which, in turn, is exhausted through ceiling level vents. In DV a noticeable interface occurs between the occupied zone of the room and a mixed hot layer near the ceiling of the room(Dominique & Guitton, 1997). Maintaining the lower boundary of this warm layer above the occupied zone is one of the many unique challenges of displacement ventilation design. Often DV systems use 100% outside air. The vertical displacement air movement means that convective heat gains introduced near the ceiling will be removed without affecting the occupied region of the room. Also a fraction of the heat gains that occur in the occupied zones rise as plumes into the upper part of the space, thereby reducing the cooling load. Similarly the fresh air will be used more effectively than with a fully mixed system: the fresh air won’t be “wasted” in the upper, unoccupied region of the room. Finally, the vertical temperature gradient means that the average room temperature can be higher for a DV conditioned room than with a conventionally conditioned room: the occupants feel the lower temperature in the lower region of the room and are unaffected by the higher temperature near the ceiling. However, whenever the outside air temperature is above \(\approx\)19\(^{\circ}\)C this advantage is mostly lost: the internal loads must be removed from the space independently of the airflow pattern (during the warmer hours buildings tend to be almost closed to the outside, operating in closed loop). The inflow temperature advantage is then only useful for the minimum outside air that must always be provided (in most cases this remaining advantage is negligible).
DV systems have limitations. In order to avoid chilling the occupants the supply air temperature used for DV is considerably higher than that used in conventional forced-air systems. This can lead to problems in removing both sensible and latent loads. Exterior spaces may have conditions that are not conducive to establishing a vertical temperature gradient. DV systems seem to be best suited to interior spaces with only moderate loads.
Non-uniform zone models[LINK]
Several types of models have been proposed as suitable for inclusion in building energy simulation (BES) programs. These models must be simple enough not to impose an undue computational burden on a BES program, yet provide enough predictive capability to produce useful comparisons between conventional and stratified zone operation strategies. ASHRAE RP-1222 (Chen & Griffith 2002) divides the candidate models into two categories: nodal and zonal. Nodal models describe the zone air as a network of nodes connected by flow paths; each node couples convectively to one or more surfaces. Zonal models are coarse–grained finite volume models. ASHRAE RP-1222 provides a short history (and examples) of each type of model. In terms of nodal models for displacement ventilation we mention the Mundt model (Mundt 1996), since it is implemented in EnergyPlus, and the Rees-Haves model (Rees & Haves 2001) since it is a well developed nodal-type model and is implemented in the RP-1222 toolkit. The Rees-Haves model, while successful in predicting the flow and temperature field for geometries similar to those used in its development, can suffer from lack of flexibility and clarity in the modeling approximations. When dealing with diverse geometries it is not clear that the flow coefficients used in the model are applicable or why they can be used since plumes, the fundamental driving mechanisms of the displacement flow, are not explicitly modeled. This is the main difference between the DV models implemented in theRP-1222 toolkit and the model that is described here.
The UCSD DV model is closer to a nodal model than to a zonal model. However, it is best to classify it in a separate category: plume equation based multi-layer models (Linden et al. 1990, Morton et al. 1956). These models assume that the dominant mechanism is plume-driven flow from discrete internal sources and that other effects (such as buoyancy driven flow at walls or windows) may be neglected. Alternatively, these heat sources also produce plumes that can be included in the model. The result is a zone divided vertically into two or more well separated regions – each region characterized by a single temperature or temperature profile. This characterization allows the physics of the heat gains and the ventilation flow to be represented in a realistic manner, without the introduction of ad hoc assumptions.
Model Description[LINK]
Single Plume Two Layer Model[LINK]
The simplest form of the plume equation based models is the case of a single plume in an adiabatic box with constant supply air flow. For this configuration two layers form in the room: a lower layer with similar density and temperature as the inflow air and a mixed upper layer with the same density / temperature as the outflow air. The main assumption of this model, successfully validated against scaled model experiments (Linden et al. 1990), is that the interface between the two layers occurs at the height (h) where the vertical buoyancy driven plume flow rate is the same as the inflow rate. For a point source of buoyancy in a non-stratified environment (a plume) the airflow rate increases with vertical distance from the source according to:
\[\dot V = C\cdot {B^{1/3}}{z^{5/3}}\]
where:
\(\dot V\) = plume volume flux [m3/s]
\(B\) = buoyancy flux [m4/s3]
\(z\) = vertical distance above source [m]
\[C = \frac{6}{5}\alpha {(\frac{9}{{10}}\alpha )^{1/3}}{\pi ^{2/3}}\]
\(\alpha\) = plume entrainment constant; a value of 0.127 is used, suitable for top-hat profiles for density and velocity across the plumes.
For an ideal gas:
\[\Delta \rho /\rho = \Delta {\rm T}/{\rm T}\]
resulting in the following relation between heat input rate and buoyancy flux:
\[B = \frac{{g\dot Q}}{{\rho {C_p}T}}\]
where
\(\rho\) = density of air [kg/m3]
\(T\) = air temperature [K]
\(g\) = acceleration of gravity [m/s2]
\(\dot Q\) = heat input rate [W]
\({C_p}\) = specific heat capacity of air [J/kgK]
Since the plume volume flow rate increases with height with exponent 5/3, for any room inflow rate (F, (m\(^{3}\)/s)) there will always be a height (h,(m)) where the plume driven flow rate matches the inflow rate. This height is obtained by setting (1.1) equal to F and solving for z = h:
\[h = {F^{3/5}}/({C^{3/5}}{B^{1/5}})\]
Substituting in and introducing air properties at 20\(^{\circ}\)C gives:
\[h = 24.55\cdot {F^{3/5}}/{\dot Q^{1/5}}\]
Multiple Plumes and Wall Heat Transfer[LINK]
Of course, it would be rare for a real world case to consist of a single point-source plume originating on the floor, unaffected by heat gains from walls and windows. For multiple plumes of equal strength a straight-forward extension of the single is possible. N plumes of unequal strength result in the formation of n vertical layers. This case is much more complex but if we are satisfied with a first order precision model the equal strength model can be used by averaging the plume strengths (Carrilho da Graça, 2003). Even in a case where all plumes are of equal strength, nearby plumes may coalesce. Plumes that are less than 0.5 meters apart at their source will coalesce within 2 meters (Kaye & Linden,2004).
As the complexity of the physical systems modeled increases some limitations must be imposed. In particular, the biggest challenge remains the interaction between wall driven boundary layers (positively and negatively buoyant) and displacement flows. For this reason, the model that is developed below is not applicable when:
Downward moving buoyancy driven airflow rate is of the same order of magnitude as plume driven flow (these airflow currents are typically generated on lateral surfaces or in the ceiling whenever these surfaces are much cooler than the room air).
Upward moving wall or floor generated buoyancy flux in the lower layer is of the same order of magnitude as plume driven flow.
Although these limitations are significant it is important to note that even in the presence of dominant convection from the floor surface, a buoyancy, two layer flow can be established whenever the plume buoyancy flux is more than 1/7 of the horizontal flux (Hunt et al. 2002). A two layer structure can also originate when the only heat source is a heated portion of the room floor, as long as the heated area does not exceed 15% of the room floor (Holford et al. 2002).
For the case of multiple non-coalescing plumes (n), with equal strength, the total vertical airflow for a given height is:
\[\dot{V} = CnB^{1/3}z^{5/3}\]
resulting in a mixed layer height of:
\[h = \frac{24.55F^{3/5}}{{n^3 \dot{Q}}^{1/5}}\]
Implementation[LINK]
The model predicts three temperatures that characterize the three main levels in the stratification of the room:
A floor level temperature T\(_{floor}\) to account for the heat transfer from the floor into the supply air.
An occupied subzone temperature T\(_{oc}\) representing the temperature of the occupied region.
An upper level temperature T\(_{mx}\) representing the temperature of the upper, mixed region and the outflow temperature.
We assume that the model for multiple, equal strength plumes (equations and will be adequate for our calculations. The supply air flow rate \(\dot V\) is obtained by summing all the air flows entering the zone: supply air, infiltration, ventilation, and inter-zone flow. The heat gain \(\dot Q\) is estimated by summing all the convective internal gains located in the occupied subzone – task lights, people, equipment – and dividing this power equally among the n plumes. With these assumptions we can describe the implementation.
The UCSD DV model is controlled by the subroutine ManageUCSDDVModel which is called from the RoomAirModelManager. The RoomAirModelManager selects which zone model will be used for each zone.
The calculation is done in subroutine CalcUCSDDV. First we calculate the convective heat gain going into the upper and lower regions.
\[\dot Q_{ocz} = \dot Q_{oc,conv} + \dot Q_{tl,conv} + \dot Q_{eleq,conv} + \dot Q_{gaseq,conv} + \dot Q_{otheq,conv} + \dot Q_{hw,conv} + \dot Q_{stmeq,conv} + \dot Q_{bb,conv}\]
\[{\dot Q_{mxz}} = {\dot Q_{gl,conv}} + {\dot Q_{ltp}} + {\dot Q_{htrad,conv}}\]
\[{\dot Q_{tot,conv}} = {\dot Q_{ocz}} + {\dot Q_{mxz}}\]
Next we sum up the inlet air flows in the form of MCP (mass flow rate times the air specific heat capacity) and MCPT (mass flow rate times C\(_{p}\) times air temperature).
\[MC{P_{zone}} = MC{P_i} + MC{P_{vent}} + MC{P_{mix}}\]
\[MC{P_{sys}} = \sum\limits_{inlets} {{{\dot m}_i}} {C_{p,i}}\]
\[MC{P_{tot}} = MC{P_{zone}} + MC{P_{sys}}\]
\[MCP{T_{zone}} = MCP{T_i} + MCP{T_{vent}} + MCP{T_{mix}}\]
\[MCP{T_{sys}} = \sum\limits_{inlets} {{{\dot m}_i}} {C_{p,i}}{T_i}\]
\[MCP{T_{tot}} = MCP{T_{zone}} + MCP{T_{sys}}\]
The number of plumes per occupant \({N_{plumesperpers}}\) is a user input. The total number of plumes in the zone is:
\[{N_{plumes}} = {N_{occ}} \cdot {N_{plumesperperson}}\]
The gains fraction \(F{r_{gains}}\) is a user input via a schedule. It is the fraction of the convective gains in the occupied subzone that remain in that subzone. Using this we calculate the total power in the plumes and the power per plume.
\[{\dot Q_{plumes}} = (1 - F{r_{gains}}) \cdot {\dot Q_{tot,conv}}\]
\[{\dot Q_{perplume}} = {\dot Q_{plumes}}/{N_{plumes}}\]
We now make an initial estimate of the height fraction Fr\(_{hb}\) (height of the boundary layer divided by the total zone height).
\[F{r_{hb}} = \left( {{{24.55} \mathord{\left/ {\vphantom {{24.55} {{H_{ceil}}}}} \right. } {{H_{ceil}}}}} \right){\left( {\frac{{0.000833 \cdot MC{P_{tot}}}}{{{N_{plumes}} \cdot \dot Q_{perplume}^{1/3}}}} \right)^{3/5}} \label{eq:RoomAirHeightBoundaryLayer}\]
where 0.000833 = \(1/({\rho_{air}} \cdot {c_{p,air}})\) converts \(MC{P_{tot}}\) to a volumetric flow rate. Next we iterate over the following 3 steps.
Iterative procedure[LINK]
Call subroutine HcUCSDDV to calculate a convective heat transfer coefficient for each surface in the zone, an effective air temperature for each surface, and HA\(_{mx}\), HAT\(_{mx}\), HA\(_{oc}\), HAT\(_{oc}\), HA\(_{fl}\), and HAT\(_{fl}\). Here HA is \(\sum\limits_{surfaces} {{h_{c,i}}} \cdot {A_i}\) for a region and HAT is \(\sum\limits_{surfaces} {{h_{c,i}}} \cdot {A_i} \cdot {T_i}\) for a region. The sum is over all the surfaces bounding the region; \({h_{c,i}}\) is the convective heat transfer coefficient for surface i, \({A_i}\) is the area of surface i, and \({T_i}\) is the surface temperature of surface i.
Recalculate \(F{r_{hb}}\) using the Equation [eq:RoomAirHeightBoundaryLayer].
Calculate the three subzone temperatures: T\(_{fl,}\) T\(_{oc}\) and T\(_{mx}\).
The h\(_{c}\)’s calculated in step 1 depend on the subzone temperatures and the boundary layer height. In turn the subzone temperatures depend on the HA and HAT’s calculated in step 1. Hence the need for iteration
Next we describe Steps 1 and 3 in more detail.
Step 1[LINK]
Subroutine HcUCSDDV is quite straightforward. It loops through all the surfaces in each zone and decides whether the surface is located in the upper, mixed subzone or the lower, occupied subzone, or if the surface is in both subzones. If entirely in one subzone the subzone temperature is stored in the surface effective temperature variable SurfTempEffBulkAir(SurfNum) and h\(_{c}\) for the surface is calculated by a call to subroutine CalcDetailedHcInForDVModel. This routine uses the “detailed” natural convection coefficient calculation that depends on surface tilt and \(\Delta {T^{1/3}}\). This calculation is appropriate for situations with low air velocity.
For surfaces that bound 2 subzones, the subroutine calculates h\(_{c}\) for each subzone and then averages them, weighting by the amount of surface in each subzone.
During the surface loop, once the h\(_{c}\) for a surface is calculated, the appropriate subzone HA and HAT sums are incremented. If a surface is in 2 subzones the HA and HAT for each subzone are incremented based on the area of the surface in each subzone.
Step 3[LINK]
The calculation of subzone temperatures follows the method used in the ZoneTempPredictorCorrector module and described in the section Basis for the System and Zone Integration. Namely a third order finite difference expansion of the temperature time derivative is used in updating the subzone temperatures. Otherwise the subzone temperatures are obtained straightforwardly by solving an energy balance equation for each subzone.
\[T_{fl} = (C_{air,fl} \cdot (3 \cdot T_{ - 1,fl} - (3/2) \cdot T_{ - 2,fl} + (1/3) \cdot T_{ - 3,fl}) + HAT_{fl} + MCPT_{tot} + 0.6 \cdot T_{oc} \cdot MCP_{tot}) / ((11/6) \cdot C_{air,fl} + HA_{fl} + 1.6 \cdot MCP_{tot})\]
\[\begin{array}{l} T_{oc} = (C_{air,oc} \cdot (3 \cdot T_{ - 1,oc} - (3/2) \cdot T_{ - 2,oc} + (1/3) \cdot T_{ - 3,oc}) + \dot Q_{ocz} \cdot Fr_{gains} + HAT_{oc} + 1.6 \cdot T_{fl} \cdot MCP_{tot}) \\ / ((11/6) \cdot C_{air,oc} + HA_{oc} + 1.6 \cdot MCP_{tot}) \end{array}\]
\[\begin{array}{l} T_{mx} = (C_{air,mx} \cdot (3 \cdot T_{ - 1,mx} - (3/2) \cdot T_{ - 2,mx} + (1/3) \cdot T_{ - 3,mx}) + \dot Q_{ocz} \cdot (1 - Fr_{gains}) + \dot Q_{mxz} + HAT_{mx} + T_{oc} \cdot MCP_{tot}) \\ / ((11/6) \cdot C_{air,mx} + HA_{mx} + MCP_{tot}) \end{array}\]
Here \({C_{air,fl}}\), \({C_{air,oc}}\), and \({C_{air,mx}}\) are the heat capacities of the air volume in each subzone. \({C_{air,mx}}\) is calculated by:
\[{R_{air,mx}} = {V_{mx}} \cdot (\Delta {z_{mx}}/{z_{ceil}}) \cdot {\rho_{air,mx}} \cdot {c_{p,air,mx}} \cdot Mu{l_{cap}}/(\Delta {t_z} \cdot 3600)\]
\[{C_{air,mx}} = {R_{air,mx}} \cdot \Delta {t_z}/\Delta {t_{sys}}\]
The other subzone air heat capacities are calculated in the same manner.
Mixed calculation[LINK]
The above iterative procedure assumed that displacement ventilation was taking place: i.e., conditions were favorable temperature stratification in the zone. Now that this calculation is complete and the subzone temperatures and depths calculated, we check to see if this assumption was justified. If not, zone conditions must be recalculated assuming a well-mixed zone.
If \({T_{mx}} < {T_{oc}}\) or \(MC{P_{tot}} \le 0\) or \({H_{fr}} \cdot {H_{ceil}} < {H_{fl,top}} + \Delta {z_{occ,\min }}\) then the following mixed calculation will replace the displacement ventilation calculation.
Note: \(\Delta {z_{occ,\min }}\) is the minimum thickness of occupied subzone. It is set to 0.2 meters. \({H_{fl,top}}\) is the height of the top of the floor subzone. It is defined to be 0.2 meters; that is, the floor subzone is always 0.2 meters thick and \({T_{fl}}\) is the temperature at 0.1 meter above the floor surface.
The mixed calculation iteratively calculates surface convection coefficients and room temperature just like the displacement ventilation calculation described above. In the mixed case however, only one zone temperature T\(_{avg}\) is calculated. The 3 subzone temperatures are then set equal to T\(_{avg}\).
First, Fr\(_{hb}\) is set equal to zero.
Then the code iterates over these steps.
1. Calculate T\(_{avg}\) using:
\[\begin{array}{l} T_{avg} = (C_{air,z} \cdot (3 \cdot T_{ - 1,z} - (3/2) \cdot T_{ - 2,z} + (1/3) \cdot T_{ - 3,z}) + \dot Q_{tot,conv} + HAT_{oc} + HAT_{mx} + HAT_{fl} + MCPT_{tot}) \\ \quad \quad \quad \quad / ((11/6) \cdot C_{air,z} + HA_{oc} + HA_{mx} + HA_{fl} + MCP_{tot}) \end{array}\]
\[{T_{mx}} = {T_{avg}}\]
\[{T_{oc}} = {T_{avg}}\]
\[{T_{fl}} = {T_{avg}}\]
2. Call HcUCSDDV to calculate the h\(_{c}\)’s.
3. Repeat step 1
Final calculations[LINK]
The displacement ventilation calculation finishes by calculating some report variables. Using Equation [eq:RoomAirHeightBoundaryLayer], setting the boundary height to 1.5 meters and solving for the flow, we calculate a minimum flow fraction:
\[{\dot V_{\min }} = {(1.5/{\rm{24}}{\rm{.55)}}^{{\rm{5/3}}}} \cdot {N_{plumes}} \cdot \dot Q_{perplume}^{1/3}\]
\[F{r_{\min flow}} = .000833 \cdot MC{P_{tot}}/{\dot V_{\min }}\]
We define heights:
\[{H_{trans}} = F{r_{hb}} \cdot {H_{ceil}}\]
\[{H_{mxavg}} = ({H_{ceil}} + {H_{trans}})/2\]
\[{H_{ocavg}} = ({H_{fltop}} + {H_{trans}})/2\]
\[{H_{flavg}} = {H_{fltop}}/2\]
Using the user defined comfort height we calculate the comfort temperature.
If mixing:
\[{T_{comf}} = {T_{avg}}\]
If displacement ventilation:
If H\(_{comf}\) < H\(_{flavg}\):
\[{T_{comf}} = {T_{fl}}\]
Else if \({H_{comf}} \ge {H_{flavg}}\) and \({H_{comf}} < {H_{ocavg}}\):
\[{T_{comf}} = ({T_{fl}}({H_{ocavg}} - {H_{comf}}) + {T_{mx}}({H_{comf}} - {H_{flavg}})/({H_{ocavg}} - {H_{flavg}})\]
Else if \({H_{comf}} \ge {H_{ocavg}}\) and \({H_{comf}} < {H_{mxavg}}\):
\[{T_{comf}} = ({T_{oc}}({H_{mxavg}} - {H_{comf}}) + {T_{mx}}({H_{comf}} - {H_{ocavg}})/({H_{mxavg}} - {H_{ocavg}})\]
Else if \({H_{comf}} \ge {H_{mxavg}}\) and \({H_{comf}} < {H_{ceil}}\):
\[{T_{comf}} = {T_{mx}}\]
Using the user defined thermostat height we calculate the temperature at the thermostat.
If mixing:
\[{T_{stat}} = {T_{avg}}\]
If displacement ventilation:
If H\(_{stat}\) < H\(_{flavg}\):
\[{T_{stat}} = {T_{fl}}\]
Else if \({H_{stat}} \ge {H_{flavg}}\) and \({H_{stat}} < {H_{ocavg}}\):
\[{T_{stat}} = ({T_{fl}}({H_{ocavg}} - {H_{stat}}) + {T_{mx}}({H_{stat}} - {H_{flavg}})/({H_{ocavg}} - {H_{flavg}})\]
Else if \({H_{stat}} \ge {H_{ocavg}}\) and \({H_{stat}} < {H_{mxavg}}\):
\[{T_{stat}} = ({T_{oc}}({H_{mxavg}} - {H_{stat}}) + {T_{mx}}({H_{stat}} - {H_{ocavg}})/({H_{mxavg}} - {H_{ocavg}})\]
Else if \({H_{stat}} \ge {H_{mxavg}}\) and \({H_{stat}} < {H_{ceil}}\):
\[{T_{stat}} = {T_{mx}}\]
The average temperature gradient is:
If \({H_{mxavg}} - {H_{flavg}} > 0.1\):
\[Grad{T_{avg}} = ({T_{mx}} - {T_{fl}})/({H_{mxavg}} - {H_{flavg}})\]
else \(Grad{T_{avg}} = - 9.999\)
The maximum temperature gradient is:
If \({H_{ocavg}} - {H_{flavg}} > 0.1\):
\[Grad{T_{\max ,1}} = ({T_{oc}} - {T_{fl}})/({H_{ocavg}} - {H_{flavg}})\]
else \(Grad{T_{\max ,1}} = - 9.999\)
If \({H_{mxavg}} - {H_{ocavg}} > 0.1\):
\[Grad{T_{\max ,2}} = ({T_{mx}} - {T_{oc}})/({H_{mxavg}} - {H_{ocavg}})\]
else \(Grad{T_{\max ,2}} = - 9.999\) and
\[Grad{T_{\max }} = \max (Grad{T_{\max ,1}},Grad{T_{\max ,2}})\]
For reporting purposes, if the zone is deemed to be mixed, the displacement ventilation report variables are set to flag values.
If \({T_{mx}} < {T_{oc}}\) or \(MC{P_{tot}} \le 0\) or \({H_{fr}} \cdot {H_{ceil}} < {H_{fl,top}} + \Delta {z_{occ,\min }}\) or \({T_{mx}} - {T_{oc}} < \Delta {T_{Crit{\mathop{\rm Re}\nolimits} p}}\)
\[Grad{T_{avg}} = - 9.999\]
\[Grad{T_{\max }} = - 9.999\]
\[F{R_{\min ,flow}} = - 1.0\]
\[{H_{trans}} = - 9.999\]
Finally, the zone node temperature is set to T\(_{mx}\).
Under-Floor Air Distribution Interior Zone Model[LINK]
Overview[LINK]
The input object RoomAirSettings:UnderFloorAirDistributionInterior provides a simple model for heat transfer and nonuniform vertical temperature profile for interior zones of a UFAD system. These zones are expected to be dominated by internal loads, a portion of which (such as occupants and workstations) will act to generate plumes. The plumes act to potentially create room air stratification, depending on the type & number of diffusers, the amount and type of load, and the system flowrate. In order to better model this situation the fully-mixed room air approximation that is currently used in most whole building analysis tools is extended to a two node approach, with the purpose of obtaining a first order precision model for vertical temperature profiles for the interior zones of UFAD systems. The use of 2 nodes allows for greatly improved prediction of thermal comfort and overall building energy performance for the increasingly popular UFAD systems.
The UCSD UFAD Interior Zone Model is one of the non-uniform zone models provided through the Room Air Manager in EnergyPlus. The intent is to provide a selection of useful non-uniform zone air models to enable the evaluation of air-conditioning techniques that use stratified or partially stratified room air. Such techniques include displacement ventilation (DV) and underfloor air distribution (UFAD) systems. The methodology can also include natural displacement ventilation and also wind-driven cross-ventilation (CV).
Underfloor air distribution systems[LINK]
UFAD systems represent, in terms of room air stratification, an intermediate condition between a well-mixed zone and displacement ventilation. Air is supplied through an underfloor plenum at low pressure through diffusers in the raised floor. The diffusers can be of various types: e.g., swirl, variable-area, displacement, and produce more or less mixing in the zone. UFAD systems are promoted as saving energy due to: higher supply air temperature; low static pressure; cooler conditions in the occupied subzone than in the upper subzone; and sweeping of some portion of the convective load (from ceiling lights, for instance) into the return air without interaction with the occupied region of the zone.
Modeling a UFAD system is quite complex and involves considerably more than just a non-uniform zone model. The zones’ coupling to the supply and return plenums must be modeled accurately (particularly radiative transfer from a warm ceiling to a cool floor and on into the supply plenum by conduction). The supply plenum must be accurately modeled, giving a good estimate of the supply air temperature and conduction heat transfer between supply & return plenums through the slab. The HVAC system must be modeled including return air bypass and various types of fan powered terminal units.
The UCSD UFAD interior zone model is similar to the UCSD DV model. The most obvious difference is that the UFAD model has no separate near-floor subzone. Like the UCSD DV model it is a plume equation based multi-layer model (2 layers in this case). The zone is modeled as being divided into 2 well separated subzones which we denote as “occupied” and “upper”. Each subzone is treated as having a single temperature. The boundary between the 2 subzones moves up & down each time step as a function of zone loads and supply air flow rate. Thus at each HVAC time step, the height of the boundary above the floor must be calculated, portions of surfaces assigned to each subzone, and a separate convective heat balance performed on each subzone.
Model Description[LINK]
The UFAD interior zone model is based upon non-dimensional analysis of the system and using the non-dimensional description to make a comparison between full-scale UCB test chamber data & small-scale UCSD salt tank measurements.
In order to do the non-dimensional comparisons, we need to define two dimensionless parameters. One is \(\Gamma\), and the other is \(\phi\). Lin & Linden (Lin & Linden, 2005) showed that in a UFAD system, the buoyancy flux of the heat source \((B)\) and the momentum flux of the cooling jets \((M)\) are the controlling parameters on the stratification. Since \([B] = {L^4}{T^{ - 3}}\) and\([M] = {L^4}{T^{ - 2}}\) , we can have a length scale as \({M^{3/4}}/{B^{1/2}}\) .
Definition of \(\Gamma\) for the single-plume single-diffuser basic model
We observed, in our small-scale experiments, that the total room height does not affect the interface position, or the height of the occupied zone. In other words, H might not be the critical length scale for the stratification. Therefore, we started to use \(\sqrt A\) as the reference length. Then \(\Gamma\) is defined as:
\[\Gamma = \frac{{{M^{3/4}}}}{{{B^{1/2}}{A^{1/2}}}} = \frac{{{{({Q^2}/A)}^{3/4}}}}{{{B^{1/2}}{A^{1/2}}}} = \frac{{{Q^{3/2}}}}{{{A^{5/4}}{B^{1/2}}}} \label{eq:GammaSinglePlumeSingleDiffuser}\]
Definition for multi-diffuser and multi-source cases
We only considered single-diffuser, single-source cases in above analysis. Suppose there are n equal diffusers and m equal heat sources in a UFAD room. We shall divide the number of diffusers up into a number of separate heat sources so that each subsection with n’ = n/m diffusers per heat source will have the same stratification as other subsections. Further, the air flow and the heat load into the subsection Q’ and B’ will be \(Q' = Q/m\) \(B' = B/m\) respectively, where Q’ and B’ are the total air flow and the total heat load for the entire UFAD space. Then the momentum flux each diffuser per heat source carries is \({M_d} = {(\frac{1}{{n'}}Q')^2}/A\). Equation [eq:GammaSinglePlumeSingleDiffuser] will be modified as:
\[\Gamma = \frac{{{{(n'{M_d})}^{3/4}}}}{{B{'^{1/2}}\sqrt {n'A} }} = \frac{{Q{'^{3/2}}}}{{{{(n'A)}^{5/4}}B{'^{1/2}}}} = \frac{{{Q^{3/2}}}}{{m{{(n'A)}^{5/4}}{B^{1/2}}}} \label{eq:GammaMultiDiffuserMultiSource}\]
Full-scale cases
Because B is the buoyancy flux of the heat sources and M is the momentum flux of the cooling jets, in a real full-scale room, we shall consider the total room net heat load (plume heat input, minus the room losses) and the total net flow rate coming from the diffusers (input room air flow, minus the room leakage). Further, if the diffuser is swirl type, the vertical momentum flux should be used.
\[\Gamma = \frac{{{{(Q\cos \theta )}^{3/2}}}}{{m{{(n'A)}^{5/4}}{{(0.0281W)}^{1/2}}}} \label{eq:GammaFullScaleCases}\]
where Q is the net flow rate coming out from all diffusers (m\(^{3}\)/s); W is the total net heat load (kW); A is the effective area of each diffuser (m\(^{2}\)); n’ is the number of diffusers per heat source; q\(\theta\) is the angle between the diffuser slots and the vertical direction and m is the number of heat sources.
Definition of \(\phi\)
In our theoretical model, two-layer stratification forms at steady state, provided that each diffuser carries the same momentum flux, and each heat source has the same heat load. We could define a dimensionless parameter \(\phi\), which indicates the strength of stratification.
Small-scale cases
In our salt-water tank experiments, fluid density r is measured. Define that:
\[\phi = \frac{{{\rho_u} - {\rho_l}}}{{{\rho_u} - {\rho_o}}} \label{eq:PhiSmallScaleCases}\]
where \(\rho_{u}\) and \(\rho_{l}\) are the fluid density of the upper layer and lower layer, separately; and \(\rho_{o}\) is the source density at the diffusers.
Therefore, \(\rho_{l}\) = \(\rho_{o}\) gives \(\phi\) = 1, which means the largest stratification (displacement ventilation case); \(\rho_{l}\) = \(\rho_{u}\) leads to \(\phi\) = 0, in which case there is no stratification (mixed ventilation case).
Full-scale cases
Similarly, we can define \(\phi\) for full-scale cases by using temperature.
\[\phi = \frac{{{T_r} - {T_{oz}}}}{{{T_r} - {T_s}}} \label{eq:PhiFullScaleCases}\]
where T\(_{r}\), T\(_{oz}\), and T\(_{s}\) are the return air temperature, the occupied zone temperature and the supply temperature, respectively (K). Again \(\phi\) = 1 occurs in displacement ventilation; while \(\phi\) = 0 happens in mixed ventilation.
Comparisons between full-scale UCB data and small-scale UCSD data
The figures (Figure 2. Data comparisons in the non-dimensional (a) regular \(\Gamma\)-\(\phi\) plot and Figure 3. (b) log-log \(\Gamma\)-\(\phi\) plot.} show the comparisons between UCB’s data and the UCSD salt tank data in the \(\Gamma\)-\(\phi\) plot. As seen in the figures, the full-scale and small-scale data are on the same trend curve. This provides the evidence that the salt-tank experiments have included most characteristics of a UFAD system. Note that big \(\Gamma\) (>20) of UCB’s experiments all have large DDR (from 1.19 to 2.18). The largest DDR (2.18) even gives a negative \(({T_r} - {T_{oz}})/({T_r} - {T_s})\), which is NOT shown in the figures.)
We could work out the occupied zone temperature by using the least-square fitting line suggested in Figure 3. Hence the interface height is needed to determine a entire two-layer stratification. Figure 4 shows the dimensionless interface height\((h/\sqrt {n'A} )\) of the UCSD small-scale experiments plotted against \(\Gamma\). Note that for the experiments with elevated heat source, the interface heights have been modified by\(h' = h - \frac{1}{2}{h_s}\) where h\(_{s}\) is the vertical position of the elevated heat source. All data then are located along a line in Figure 4. Since the salt-tank experiments are concluded to represent important characteristics of a full-scale UFAD room, this figure provides some guidelines for estimate the interface position in a real UFAD room.
Formulas for EnergyPlus based on the dimensionless parameter \(\Gamma\)
If we have input including the supply temperature T\(_{s}\) (K); the number of diffusers n; the number of heat sources m; the vertical position of the heat sources h\(_{s}\) (m); the heat load W (kW); the effective area of a diffuser A (m\(^{2}\)); and the total supply air flow rate Q (m\(^{3}\)/s) then the output will be;
\[{T_r} = \frac{{0.0281W}}{{Qg}}{T_s} + {T_s}\]
\[{T_{oz}} = {T_r} - 1.6{\Gamma ^{ - 0.76}}({T_r} - {T_s})\]
\[h = \sqrt {\frac{n}{m}A} (7.43ln(\Gamma ) - 1.35) + \frac{1}{2}{h_s} \label{eq:InterfaceHeightEquation}\]
where T\(_{r}\) is the return temperature (K); T\(_{oz}\) is the occupied subzone temperature (K); h is the interface height (m); and \(\Gamma\) is defined above.
Implementation[LINK]
The implementation closely follows the procedure described in the displacement ventilation zone model. The model predicts two temperatures that characterize the two main levels in the stratification of the room:
An occupied subzone temperature T\(_{oc}\) representing the temperature of the occupied region;
An upper level temperature T\(_{mx}\) representing the temperature of the upper, mixed region and the outflow temperature.
We will use Equation [eq:InterfaceHeightEquation] to calculate the interface height and do a heat balance calculation on each subzone. \(\Gamma\) is given by Equation [eq:GammaFullScaleCases]. The supply air flow rate \(\dot V\) is obtained by summing all the air flows entering the zone: supply air, infiltration, ventilation, and inter-zone flow. The heat gain \(\dot Q\) is estimated by summing all the convective internal gains located in the occupied subzone – task lights, people, equipment – and dividing this power equally among the n plumes. With these assumptions we can describe the implementation.
The UCSD UFI model is controlled by the subroutine ManageUCSDUFModels which is called from the RoomAirModelManager. The RoomAirModelManager selects which zone model will be used for each zone.
The calculation is done in subroutine CalcUCSDUI. First we calculate the convective heat gain going into the upper and lower regions.
\[\begin{split} \dot Q_{ocz} =& \dot Q_{oc,conv} + \dot Q_{tl,conv} + \dot Q_{eleq,conv} + \dot Q_{gaseq,conv} + \dot Q_{otheq,conv} + \dot Q_{hw,conv} + \dot Q_{stmeq,conv} \\ &+ \dot Q_{bb,conv} + \dot Q_{waterheater,conv} + \dot Q_{fuelcell,conv} \end{split}\]
\[{\dot Q_{mxz}} = {\dot Q_{gl,conv}} + {\dot Q_{ltp}} + {\dot Q_{htrad,conv}}\]
\[{\dot Q_{tot,conv}} = {\dot Q_{ocz}} + {\dot Q_{mxz}}\]
Next we sum up the inlet air flows in the form of MCP (mass flow rate times the air specific heat capacity) and MCPT (mass flow rate times C\(_{p}\) times air temperature).
\[MC{P_{zone}} = MC{P_i} + MC{P_{vent}} + MC{P_{mix}}\]
\[MC{P_{sys}} = \sum\limits_{inlets} {{{\dot m}_i}} {C_{p,i}}\]
\[{\dot m_{systot}} = \sum\limits_{inlets} {{{\dot m}_i}}\]
\[MC{P_{tot}} = MC{P_{zone}} + MC{P_{sys}}\]
\[MCP{T_{zone}} = MCP{T_i} + MCP{T_{vent}} + MCP{T_{mix}}\]
\[MCP{T_{sys}} = \sum\limits_{inlets} {{{\dot m}_i}} {C_{p,i}}{T_i}\]
\[MCP{T_{tot}} = MCP{T_{zone}} + MCP{T_{sys}}\]
The number of plumes per occupant \({N_{plumesperpers}}\) is a user input. The total number of plumes in the zone is:
\[{N_{plumes}} = {N_{occ}} \cdot {N_{plumesperperson}}\]
Using this we calculate the total power in the plumes and the power per plume.
\[{\dot Q_{plumes}} = {\dot Q_{ocz}}\]
\[{\dot Q_{perplume}} = {\dot Q_{plumes}}/{N_{plumes}}\]
The number of diffusers per plumes is also a user input. To obtain the number of diffusers in the zone:
\[{N_{diffusers}} = {N_{plumes}}\cdot {N_{diffusersperplume}}\]
The area A\(_{diff}\) is also a user input. For swirl diffusers and for displacement diffusers this area is used as input. For the variable area diffusers, though, we calculate the area. We assume 400 ft/min velocity at the diffuser and a design flow rate per diffuser is 150 cfm (.0708 m\(^{3}\)/s). The design area of the diffuser is 150 ft\(^{3}\)/min / 400 ft/min = .575 ft\(^{2}\) = .035 m\(^{2}\). Then the variable area each time step is
\[{A_{diff}} = {\dot m_{systot}}/(.0708\cdot {N_{diffusers}})\]
We now calculate the height fraction Fr\(_{hb}\) (height of boundary layer divided by the total zone height).
\[\Gamma = \frac{{{{({{\dot m}_{systot}}\cdot \cos ({\theta_{throw}}))}^{3/2}}}}{{{N_{plumes}}{{({N_{diffusersperplume}}\cdot {A_{diff}})}^{5/4}}{{(0.0281\cdot 0.001\cdot {{\dot Q}_{plumes}})}^{1/2}}}}\]
\[F{r_{hb}} = \frac{{{{({N_{diffusersperplume}}\cdot {A_{diff}})}^{1/2}}(7.43\ln (\Gamma ) - 1.35 + 0.5{H_s})}}{{{H_{ceil}}}}\]
where \(\theta_{throw}\) is a user input: the angle between the diffuser slots and vertical; and H\(_{s}\) is the source height above the floor (m).
Next we iterate over the following 2 steps.
Iterative procedure[LINK]
Call subroutine HcUCSDUF to calculate a convective heat transfer coefficient for each surface in the zone, an effective air temperature for each surface, and HA\(_{mx}\), HAT\(_{mx}\), HA\(_{oc}\), HAT\(_{oc}\). Here HA is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot {A_i}}\) for a region and HAT is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot {A_i}} \cdot {T_i}\) for a region. The sum is over all the surfaces bounding the region; \({{h_{c,i}}}\) is the convective heat transfer coefficient for surface i, \({A_i}\) is the area of surface i, and \({T_i}\) is the surface temperature of surface i.
Calculate the two subzone temperatures: T\(_{oc}\) and T\(_{mx}\).
The h\(_{c}\)’s calculated in step 1 depend on the subzone temperatures. In turn the subzone temperatures depend on the HA and HAT’s calculated in step 1. Hence the need for iteration.
Next we describe Steps 1 and 2 in more detail.
Step 1[LINK]
Subroutine HcUCSDUF is quite straightforward. It loops through all the surfaces in each zone and decides whether the surface is located in the upper, mixed subzone or the lower, occupied subzone, or if the surface is in both subzones. If entirely in one subzone the subzone temperature is stored in the surface effective temperature variable SurfTempEffBulkAir(SurfNum) and h\(_{c}\) for the surface is calculated by a call to subroutine CalcDetailedHcInForDVModel. This routine uses the “detailed” natural convection coefficient calculation that depends on surface tilt and \(\Delta {T^{1/3}}\) . This calculation is appropriate for situations with low air velocity.
For surfaces that bound 2 subzones, the subroutine calculates h\(_{c}\) for each subzone and then averages them, weighting by the amount of surface in each subzone.
During the surface loop, once the h\(_{c}\) for a surface is calculated, the appropriate subzone HA and HAT sums are incremented. If a surface is in 2 subzones the HA and HAT for each subzone are incremented based on the area of the surface in each subzone.
Step 2[LINK]
The calculation of subzone temperatures follows the method used in the ZoneTempPredictorCorrector module and described in the section Basis for the System and Zone Integration. Namely a third order finite difference expansion of the temperature time derivative is used in updating the subzone temperatures. Otherwise the subzone temperatures are obtained straightforwardly by solving an energy balance equation for each subzone.
\[\begin{array}{cl} T_{oc} & = \frac{(C_{air,oc} \cdot (3 \cdot T_{ - 1,oc} - (3/2) \cdot T_{ - 2,oc} + (1/3) \cdot T_{ - 3,oc}) + \dot Q_{ocz} \cdot Fr_{gains} + HAT_{oc} + MCPT_{tot})} {((11/6) \cdot C_{air,oc} + HA_{oc} + MCP_{tot})} \\ T_{mx} & = \frac{(C_{air,mx} \cdot (3 \cdot T_{ - 1,mx} - (3/2) \cdot T_{ - 2,mx} + (1/3) \cdot T_{ - 3,mx}) + \dot Q_{ocz} \cdot (1 - Fr_{gains}) + \dot Q_{mxz} + HAT_{mx} + T_{oc} \cdot MCP_{tot})} {((11/6) \cdot C_{air,mx} + HA_{mx} + MCP_{tot})} \end{array}\]
Here \({C_{air,oc}}\) and \({C_{air,mx}}\) are the heat capacities of the air volume in each subzone. \({C_{air,mx}}\) is calculated by:
\[{R_{air,mx}} = {V_{mx}} \cdot \left( {\Delta {z_{mx}}/{z_{ceil}}} \right) \cdot {\rho_{air,mx}} \cdot {c_{p,air,mx}} \cdot Mu{l_{cap}}/\left( {\Delta {t_z} \cdot 3600} \right)\]
\[{c_{air,mx}} = {R_{air,mx}} \cdot \Delta {t_z}/\Delta {t_{sys}}\]
The gains fraction \(F{r_{gains}}\) is a user input via a schedule. It is the fraction of the convective gains in the occupied subzone that remain in that subzone.
The other subzone air heat capacities are calculated in the same manner.
Mixed calculation[LINK]
The above iterative procedure assumed that the UFAD nonuniform zone model was appropriate: i.e., conditions were favorable temperature stratification in the zone. Now that this calculation is complete and the subzone temperatures and depths calculated, we check to see if this assumption was justified. If not, zone conditions must be recalculated assuming a well-mixed zone.
If \({T_{mx}} < {T_{oc}}\) or \(MC{P_{tot}} \le 0\) or \({H_{fr}} \cdot {H_{ceil}} < \Delta {z_{occ,\min }}\), then the following mixed calculation will replace the UFAD interior zone calculation.
Note: \(\Delta {z_{occ,min}}\) is the minimum thickness of occupied subzone. It is set to 0.2 meters.
The mixed calculation iteratively calculates surface convection coefficients and room temperature just like the displacement ventilation calculation described above. In the mixed case however, only one zone temperature T\(_{avg}\) is calculated. The 3 subzone temperatures are then set equal to T\(_{avg}\).
First, Fr\(_{hb}\) is set equal to zero.
Then the code iterates over these steps.
Calculate T\(_{avg}\) using:
\[\begin{array}{rl} T_{avg} &= ({C_{air,z}} \cdot (3 \cdot {T_{ - 1,z}} - (3/2) \cdot {T_{ - 2,z}} + (1/3) \cdot {T_{ - 3,z}}) \\ &+ {{\dot Q}_{tot,conv}} + HA{T_{oc}} + HA{T_{mx}} + MCP{T_{tot}})/((11/6) \cdot {C_{air,z}} + H{A_{oc}} + H{A_{mx}} + MC{P_{tot}}) \end{array}\]
\[{T_{mx}} = {T_{avg}}\]
\[{T_{oc}} = {T_{avg}}\]
Call HcUCSDUF to calculate the h\(_{c}\)’s.
Repeat step 1
Final calculations[LINK]
The UFAD interior zone calculation finishes by calculating some report variables.
We define heights:
\[{H_{trans}} = F{r_{hb}} \cdot {H_{ceil}}\]
\[{H_{mxavg}} = \left( {{H_{ceil}} + {H_{trans}}} \right)/2\]
\[{H_{ocavg}} = {H_{trans}}/2\]
Using the user defined comfort height we calculate the comfort temperature.
If mixing:
\[{T_{comf}} = {T_{avg}}\]
If UFAD:
If \({H_{comf}} < {H_{ocavg}}\):
\[{T_{comf}} = {T_{occ}}\]
Else if \({H_{comf}} \ge {H_{ocavg}}\) and \({H_{comf}} < {H_{mxavg}}\):
\[{T_{comf}} = \left( {{T_{oc}}\left( {{H_{mxavg}} - {H_{comf}}} \right) + {T_{mx}}\left( {{H_{comf}} - {H_{ocavg}}} \right)} \right)/\left( {{H_{mxavg}} - {H_{ocavg}}} \right)\]
Else if \({H_{comf}} \ge {H_{mxavg}}\) and \({H_{comf}} < {H_{ceil}}\):
\[{T_{comf}} < {T_{mx}}\]
Using the user defined thermostat height we calculate the temperature at the thermostat.
If mixing:
\[{T_{stat}} < {T_{avg}}\]
If UFAD:
If\({H_{stat}} < {H_{ocavg}}\):
\[{T_{stat}} = {T_{occ}}\]
Else if \({H_{stat}} \ge {H_{ocavg}}\) and \({H_{stat}} < {H_{mxavg}}\):
\[{T_{stat}} = \left( {{T_{oc}}\left( {{H_{mxavg}} - {H_{stat}}} \right) + {T_{mx}}\left( {{H_{stat}} - {H_{ocavg}}} \right)} \right)/\left( {{H_{mxavg}} - {H_{ocavg}}} \right)\]
Else if \({H_{stat}} \ge {H_{mxavg}}\) and \({H_{stat}} < {H_{ceil}}\):
\[{T_{stat}} = {T_{mx}}\]
The average temperature gradient is:
If \({H_{mxavg}} - {H_{occavg}} > 0.1\):
\[Grad{T_{avg}} = ({T_{mx}} - {T_{occ}})/({H_{mxavg}} - {H_{occavg}})\]
else \(Grad{T_{avg}} = 0.0\).
Finally, the zone node temperature is set to T\(_{mx}\).
Other variables that are reported out are \(\Gamma\) and \(\phi\).
\[\phi = ({T_{mx}} - {T_{occ}})/({T_{mx}} - {T_{sup}})\]
where \({T_{sup}}\) is the zone supply air temperature.
Under-Floor Air Distribution Exterior Zone Model[LINK]
Overview[LINK]
The input object RoomAirSettings:UnderFloorAirDistributionExterior provides a simple model for heat transfer and a nonuniform vertical temperature profile for exterior zones of a UFAD system. These zones are expected to be dominated by internal loads, a portion of which (such as occupants and workstations) will act to generate plumes, and by window solar and conduction heat gains. The solar radiation penetrating the room is not expected to generate plumes. However, a window plume is likely to be generated in sunny conditions, particularly if an interior blind is deployed. Thus the exterior UFAD zone will have potentially have plumes from people and equipment and plumes arising from the windows. The plumes act to potentially create room air stratification, depending on the type & number of diffusers, the amount and type of load, and the system flowrate. In order to better model this situation the fully-mixed room air approximation that is currently used in most whole building analysis tools is extended to a two node approach, with the purpose of obtaining a first order precision model for vertical temperature profiles for the exterior zones of UFAD systems. The use of 2 nodes allows for greatly improved prediction of thermal comfort and overall building energy performance for the increasingly popular UFAD systems.
TheUCSD UFAD Exterior Zone Model is one of the non-uniform zone models provided through the Room Air Manager in EnergyPlus. The intent is to provide a selection of useful non-uniform zone air models to enable the evaluation of air-conditioning techniques that use stratified or partially stratified room air. Such techniques include displacement ventilation (DV) and underfloor air distribution (UFAD) systems. The methodology can also include natural displacement ventilation and also wind-driven cross-ventilation (CV).
Underfloor air distribution systems[LINK]
UFAD systems represent, in terms of room air stratification, an intermediate condition between a well-mixed zone and displacement ventilation. Air is supplied through an underfloor plenum at low pressure through diffusers in the raised floor. The diffusers can be of various types: e.g., swirl, variable-area, displacement, and produce more or less mixing in the zone. UFAD systems are promoted as saving energy due to: higher supply air temperature; low static pressure; cooler conditions in the occupied subzone than in the upper subzone; and sweeping of some portion of the convective load (from ceiling lights, for instance) into the return air without interaction with the occupied region of the zone.
Modeling a UFAD system is quite complex and involves considerably more than just a non-uniform zone model. The zones’ coupling to the supply and return plenums must be modeled accurately (particularly radiative transfer from a warm ceiling to a cool floor and on into the supply plenum by conduction). The supply plenum must be accurately modeled, giving a good estimate of the supply air temperature and conduction heat transfer between supply & return plenums through the slab. The HVAC system must be modeled including return air bypass and various types of fan powered terminal units.
The UCSD UFAD exterior zone model is similar to the UCSD interior zone model. The most obvious difference is that the exterior UFAD has 2 different types of plume sources: people & equipment and windows. Like the UCSD UFAD interior model it is a plume equation based multi-layer model (2 layers in this case). The zone is modeled as being divided into 2 well separated subzones which we denote as “occupied” and “upper”. Each subzone is treated as having a single temperature. The boundary between the 2 subzones moves up & down each time step as a function of zone loads and supply air flow rate. Thus at each HVAC time step, the height of the boundary above the floor must be calculated, portions of surfaces assigned to each subzone, and a separate convective heat balance performed on each subzone.
Model Description[LINK]
As in the interior zone case, we define 2 dimensionless parameters: \(\Gamma\) and \(\phi\). The definitions of the 2 parameters are the same as in the previous section (Equations [eq:GammaSinglePlumeSingleDiffuser], [eq:GammaMultiDiffuserMultiSource], [eq:GammaFullScaleCases], [eq:PhiSmallScaleCases], and [eq:PhiFullScaleCases]). As in the previous case, the experimental data can be plotted versus \(\Gamma\) and lines fitted to the data give the following formulas for occupied subzone temperature and interface height.
\[{T_{oz}} = {T_r} - 1.4{\Gamma ^{ - 0.6}}({T_r} - {T_s})\]
\[h = \sqrt {\frac{n}{m}A} (11.03ln(\Gamma ) - 10.73) + \frac{1}{2}{h_s} \label{eq:InterfaceHeightUFAD}\]
where T\(_{oz}\) is the occupied subzone temperature (K); T\(_{r}\) is the return temperature (K); \(\Gamma\) is the dimensionless height parameter defined above; T\(_{s}\) is the supply temperature (K); h is the interface height (m); n is the number of diffusers; m is the number of heat sources; A is the effective area of a diffuser (m\(^{2}\)); and h\(_{s}\) is the vertical position of the heat sources (m). The formula for T\(_{r}\) is the same as in the previous section.
Implementation[LINK]
The implementation closely follows the procedure described in the UFAD interior zone model. The model predicts two temperatures that characterize the two main levels in the stratification of the room:
An occupied subzone temperature T\(_{oc}\) representing the temperature of the occupied region;
An upper level temperature T\(_{mx}\) representing the temperature of the upper, mixed region and the outflow temperature.
We will use Equation [eq:InterfaceHeightUFAD] to calculate the interface height and do a heat balance calculation on each subzone. \(\Gamma\) is given by Equation [eq:GammaFullScaleCases]. The supply air flow rate \(\dot V\) is obtained by summing all the air flows entering the zone: supply air, infiltration, ventilation, and inter-zone flow. The heat gain \(\dot Q\) is estimated by summing all the convective internal gains located in the occupied subzone – task lights, people, equipment – and adding to this the convective gain coming from the window surface. With these assumptions we can describe the implementation.
The UCSD UFE model is controlled by the subroutine ManageUCSDUFModels which is called from the RoomAirModelManager. The RoomAirModelManager selects which zone model will be used for each zone.
The calculation is done in subroutine CalcUCSDUEI. First we calculate the convective heat gain going into the upper and lower regions.
\[\begin{split} \dot Q_{ocz} =& \dot Q_{oc,conv} + \dot Q_{tl,conv} + \dot Q_{eleq,conv} + \dot Q_{gaseq,conv} + \dot Q_{otheq,conv} + \dot Q_{hw,conv} + \dot Q_{stmeq,conv} \\ &+ \dot Q_{bb,conv} + \dot Q_{waterheater,conv} + \dot Q_{fuelcell,conv} \end{split}\]
\[{\dot Q_{mxz}} = {\dot Q_{gl,conv}} + {\dot Q_{ltp}} + {\dot Q_{htrad,conv}}\]
\[{\dot Q_{tot,conv}} = {\dot Q_{ocz}} + {\dot Q_{mxz}}\]
Next we sum up the inlet air flows in the form of MCP (mass flow rate times the air specific heat capacity) and MCPT (mass flow rate times C\(_{p}\) times air temperature).
\[MC{P_{zone}} = MC{P_i} + MC{P_{vent}} + MC{P_{mix}}\]
\[MCP_{sys} = \sum\limits_{inlets} {\dot m_i C_{p,i}}\]
\[{\dot m_{systot}} = \sum\limits_{inlets} {{{\dot m}_i}}\]
\[MC{P_{tot}} = MC{P_{zone}} + MC{P_{sys}}\]
\[MCP{T_{zone}} = MCP{T_i} + MCP{T_{vent}} + MCP{T_{mix}}\]
\[MCP{T_{sys}} = \sum\limits_{inlets} {{{\dot m}_i}} {C_{p,i}}{T_i}\]
\[MCP{T_{tot}} = MCP{T_{zone}} + MCP{T_{sys}}\]
For exterior zone model, we assume one plume: \({N_{plumes}} = 1\). The number of diffusers in the zone \({N_{diffusers}}\) is a user input.
The area A\(_{diff}\) is also a user input. For swirl diffusers, linear bar grilles, and displacement diffusers this area is used as input. For the variable area diffusers, though, we calculate the area. We assume 400 ft/min velocity at the diffuser and a design flow rate per diffuser is 150 cfm (.0708 m\(^{3}\)/s). The design area of the diffuser is 150 ft\(^{3}\)/min / 400 ft/min = .575 ft\(^{2}\) = .035 m\(^{2}\). Then the variable area each time step is:
\[{A_{diff}} = {\dot m_{systot}}/(.0708\cdot {N_{diffusers}})\]
We now make an initial estimate of the convective gain from the windows.
\[{\dot Q_{win,conv}} = \sum\limits_{i = 1,{n_{win}}} {{h_{c,i}}} {A_{win,i}}{T_{winsurf,i}} - {h_{c,i}}{A_{win,i}}{T_{mat}}\]
Then:
\[{\dot Q_{plumes}} = {\dot Q_{ocz}} + {\dot Q_{win,conv}}\]
We now calculate the height fraction Fr\(_{hb}\) (height of boundary layer divided by the total zone height).
\[\Gamma = \frac{{{{({{\dot m}_{systot}}\cdot \cos ({\theta_{throw}}))}^{3/2}}}}{{{{({N_{diffusers}}\cdot {A_{diff}})}^{5/4}}{{(0.0281\cdot 0.001\cdot {{\dot Q}_{plumes}})}^{1/2}}}}\]
\[F{r_{hb}} = \frac{{{{({N_{diffusers}}\cdot {A_{diff}})}^{1/2}}(11.03ln(\Gamma ) - 10.73 + 0.5{H_s})}}{{{H_{ceil}}}}\]
where \(\theta_{throw}\) is a user input: the angle between the diffuser slots and vertical; and H\(_{s}\) is the source height above the floor (m).
Next we iterate over the following 2 steps.
Iterative procedure[LINK]
Call subroutine HcUCSDUF to calculate a convective heat transfer coefficient for each surface in the zone, an effective air temperature for each surface, and HA\(_{mx}\), HAT\(_{mx}\), HA\(_{oc}\), HAT\(_{oc}\), HA\(_{mx,win}\), HAT\(_{mx,win}\), HA\(_{oc,win}\),and HAT\(_{oc,win}\). Here HA is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot {A_i}}\) for a region and HAT is \(\sum\limits_{surfaces} {{h_{c,i}} \cdot {A_i}} \cdot {T_i}\) for a region. The sum is over all the surfaces bounding the region; \({h_{c,i}}\) is the convective heat transfer coefficient for surface i, \({A_i}\) is the area of surface i, and \({T_i}\) is the surface temperature of surface i. Variables with the win subscript are summed over window surfaces only. Then the convective gain from the window is recalculated using Equation [eq:QdotwinconvUSCDUFAD] and the power in the plume is recalculated: \({\dot Q_{plumes}} = {\dot Q_{ocz}} + {\dot Q_{win,conv}}\) .
Calculate the two subzone temperatures: T\(_{oc}\) and T\(_{mx}\).
\[{\dot Q_{win,conv}} = HA{T_{mx,win}} + HA{T_{oc,win}} - H{A_{mx,win}}{T_{mx}} - H{A_{oc,win}}{T_{oc}} \label{eq:QdotwinconvUSCDUFAD}\]
The h\(_{c}\)’s calculated in step 1 depend on the subzone temperatures. In turn the subzone temperatures depend on the HA and HAT’s calculated in step 1. Hence the need for iteration.
Next we describe Steps 1 and 3 in more detail.
Step 1[LINK]
Subroutine HcUCSDUF is quite straightforward. It loops through all the surfaces in each zone and decides whether the surface is located in the upper, mixed subzone or the lower, occupied subzone, or if the surface is in both subzones. If entirely in one subzone the subzone temperature is stored in the surface effective temperature variable SurfTempEffBulkAir(SurfNum) and h\(_{c}\) for the surface is calculated by a call to subroutine CalcDetailedHcInForDVModel. This routine uses the “detailed” natural convection coefficient calculation that depends on surface tilt and \(\Delta {T^{1/3}}\) . This calculation is appropriate for situations with low air velocity.
For surfaces that bound 2 subzones, the subroutine calculates h\(_{c}\) for each subzone and then averages them, weighting by the amount of surface in each subzone.
During the surface loop, once the h\(_{c}\) for a surface is calculated, the appropriate subzone HA and HAT sums are incremented. If a surface is in 2 subzones the HA and HAT for each subzone are incremented based on the area of the surface in each subzone. The subroutine calculates a separate HA and HAT for the windows for use in calculating the window convective gain.
Step 2[LINK]
The calculation of subzone temperatures follows the method used in the ZoneTempPredictorCorrector module and described in the section Basis for the System and Zone Integration. Namely a third order finite difference expansion of the temperature time derivative is used in updating the subzone temperatures. Otherwise the subzone temperatures are obtained straightforwardly by solving an energy balance equation for each subzone.
\[\begin{array}{cl} T_{oc} & = \frac{(C_{air,oc} \cdot (3 \cdot T_{ - 1,oc} - (3/2) \cdot T_{ - 2,oc} + (1/3) \cdot T_{ - 3,oc}) + \dot Q_{ocz} \cdot Fr_{gains} + HAT_{oc} + MCPT_{tot})} {((11/6) \cdot C_{air,oc} + HA_{oc} + MCP_{tot})} \\ T_{mx} & = \frac{(C_{air,mx} \cdot (3 \cdot T_{ - 1,mx} - (3/2) \cdot T_{ - 2,mx} + (1/3) \cdot T_{ - 3,mx}) + \dot Q_{ocz} \cdot (1 - Fr_{gains}) + \dot Q_{mxz} + HAT_{mx} + T_{oc} \cdot MCP_{tot})} {((11/6) \cdot C_{air,mx} + HA_{mx} + MCP_{tot})} \end{array}\]
Here \({C_{air,oc}}\) and \({C_{air,mx}}\) are the heat capacities of the air volume in each subzone. \({C_{air,mx}}\) is calculated by:
\[{R_{air,mx}} = {V_{mx}} \cdot \left( {\Delta {z_{mx}}/{z_{ceil}}} \right) \cdot {\rho_{air,mx}} \cdot {c_{p,air,mx}} \cdot Mu{l_{cap}}/\left( {\Delta {t_z} \cdot 3600} \right)\]
\[{c_{air,mx}} = {R_{air,mx}} \cdot \Delta {t_z}/\Delta {t_{sys}}\]
The gains fraction \(F{r_{gains}}\) is a user input via a schedule. It is the fraction of the convective gains in the occupied subzone that remain in that subzone.
The other subzone air heat capacities are calculated in the same manner.
Mixed calculation[LINK]
The above iterative procedure assumed that the UFAD zone model was applicable: i.e., conditions were favorable temperature stratification in the zone. Now that this calculation is complete and the subzone temperatures and depths calculated, we check to see if this assumption was justified. If not, zone conditions must be recalculated assuming a well-mixed zone.
If \({T_{mx}} < {T_{oc}}\) or \(MC{P_{tot}} \le 0\) or \({H_{fr}} \cdot {H_{ceil}} < \Delta {z_{occ,\min }}\), then the following mixed calculation will replace the UFAD exterior zone calculation.
Note: \(\Delta {z_{occ,min}}\) is the minimum thickness of occupied subzone. It is set to 0.2 meters.
The mixed calculation iteratively calculates surface convection coefficients and room temperature just like the displacement ventilation calculation described above. In the mixed case however, only one zone temperature T\(_{avg}\) is calculated. The 2 subzone temperatures are then set equal to T\(_{avg}\).
First, Fr\(_{hb}\) is set equal to zero.
Then the code iterates over these steps.
Calculate T\(_{avg}\) using:
\[\begin{array}{rl} T_{avg} &= ({C_{air,z}} \cdot (3 \cdot {T_{ - 1,z}} - (3/2) \cdot {T_{ - 2,z}} + (1/3) \cdot {T_{ - 3,z}}) \\ &+ {{\dot Q}_{tot,conv}} + HA{T_{oc}} + HA{T_{mx}}MCP{T_{tot}})/((11/6) \cdot {C_{air,z}} + H{A_{oc}} + H{A_{mx}} + MC{P_{tot}}) \end{array}\]
\[{T_{mx}} = {T_{avg}}\]
\[{T_{oc}} = {T_{avg}}\]
Call HcUCSDUF to calculate the h\(_{c}\)’s.
Repeat step 1.
Final calculations[LINK]
The UFAD exterior zone calculation finishes by calculating some report variables.
We define heights:
\[{H_{trans}} = F{r_{hb}} \cdot {H_{ceil}}\]
\[{H_{mxavg}} = \left( {{H_{ceil}} + {H_{trans}}} \right)/2\]
\[{H_{ocavg}} = {H_{trans}}/2\]
Using the user defined comfort height we calculate the comfort temperature.
If mixing:
\[{T_{comf}} = {T_{avg}}\]
If UFAD:
If \({H_{comf}} < {H_{ocavg}}\):
\[{T_{comf}} = {T_{occ}}\]
Else if \({H_{comf}} \ge {H_{ocavg}}\) and \({H_{comf}} < {H_{mxavg}}\):
\[{T_{comf}} = \left( {{T_{oc}}\left( {{H_{mxavg}} - {H_{comf}}} \right) + {T_{mx}}\left( {{H_{comf}} - {H_{ocavg}}} \right)} \right)/\left( {{H_{mxavg}} - {H_{ocavg}}} \right)\]
Else if \({H_{comf}} \ge {H_{mxavg}}\) and \({H_{comf}} < {H_{ceil}}\):
\[{T_{comf}} < {T_{mx}}\]
Using the user defined thermostat height, we calculate the temperature at the thermostat.
If mixing:
\[{T_{stat}} = {T_{avg}}\]
If UFAD:
If\({H_{stat}} < {H_{ocavg}}\):
\[{T_{stat}} = {T_{occ}}\]
Else if \({H_{stat}} \ge {H_{ocavg}}\) and \({H_{stat}} < {H_{mcavg}}\):
\[{T_{stat}} = \left( {{T_{oc}}\left( {{H_{mxavg}} - {H_{stat}}} \right) + {T_{mx}}\left( {{H_{stat}} - {H_{ocavg}}} \right)} \right)/\left( {{H_{mxavg}} - {H_{ocavg}}} \right)\]
Else if \({H_{stat}} \ge {H_{mxavg}}\) and \({H_{stat}} < {H_{ceil}}\):
\[{T_{stat}} = {T_{mx}}\]
The average temperature gradient is:
If \({H_{mxavg}} - {H_{occavg}} > 0.1\):
\[Grad{T_{avg}} = ({T_{mx}} - {T_{occ}})/({H_{mxavg}} - {H_{occavg}})\]
else \(Grad{T_{avg}} = 0.0\).
Finally, the zone node temperature is set to T\(_{mx}\).
Other variables that are reported out are \(\Gamma\) and \(\phi\).
\[\phi = ({T_{mx}} - {T_{occ}})/({T_{mx}} - {T_{sup}})\]
where \({T_{sup}}\) is the zone supply air temperature.
Cross Ventilation Room Air Model[LINK]
Overview[LINK]
The input object RoomAirSettings:CrossVentilation provides a simple model, developed by UCSD, for heat transfer and vertical temperature profile prediction in cross ventilated rooms. This model was developed using scaling analysis, experimental correlations, computational fluid dynamics, and approximate solutions of the Navier-Stokes equations. The model distinguishes two regions in the room, the main jet region and the recirculations, and predicts characteristic airflow velocities and average air temperatures. The model consists of a set of analytical expressions that clearly display the first order effects of room geometry and inflow characteristics on Cross Ventilation airflow and heat transfer.
Cross-Ventilation[LINK]
The CV model is developed from analysis of the flow structure inside a cross-ventilated room, which naturally divides into distinct jet and recirculation flow regions. The flows are calculated for different room and opening geometries using CFD. Correlations with room and aperture parameters are then deduced for a characteristic temperature and velocity in the jet and recirculation regions.
The SS 2-opening model is developed from analysis of wind tunnel data for pressure and ventilation rates, correlating the flow rate with the pressure difference between the two openings, with contributions from both the mean and unsteady components. The local velocity estimate is also based on wind tunnel data, and gives the velocity as a function of wind angle.
Approach[LINK]
With the widespread availability of air conditioning, natural ventilation became a rare feature in commercial buildings in the US. The recent resurgence of the use of natural ventilation relies on systems whose performance can be predicted and controlled, often working in conjunction with mechanical ventilation and cooling systems in hybrid ventilation configurations. In this contemporary approach natural ventilation deals with thermal and air renewal requirements in the cold and mild season and often plays a limited role in the warm season (due to the need to control indoor velocities and noise levels). Whereas traditional natural ventilation systems often use large openings (1-4m\(^{2}\)), both recent and expected future examples tend to use smaller openings (0.2-1m\(^{2}\)), often in multiple opening configurations that allow for efficient heat removal with controlled indoor velocities.
The modeling approach used is based on simplified solutions of the Navier Stokes equations that are integrated into scaling formulae that predict relevant flow quantities (characteristic velocities and temperatures) in distinct regions of the flow. The scaling expressions that will be developed use the principle of flow similarity: for a given room geometry and flow regime (either laminar or turbulent), the flow pattern is constant. In the present case, where the goal is to predict internal velocities using as input the inflow velocity (that can be estimated by the airflow rate predicted by the AirflowNetwork model), the principle can be stated as:
(Velocity in a given position in the room) = (Inflow velocity) × (Constant)
The constant is non-dimensional and depends on the position or room zone where the prediction is being made. For a given geometry the constant can be measured experimentally or inferred from a detailed CFD simulation. It is expected that the constant may depend on non-dimensional room geometry parameters that can be inferred from a careful analysis of the flow. The proposed CV model uses a set of CFD simulations, in conjunction with simplified solutions of the Navier Stokes equations, to obtain the constants and their dependence on room geometry features.
The updated CV model proposed in this document was developed using inlet areas A\(_{in}\) and room cross-section areas A\(_{RM}\) in the following range:
0.5% < A\(_{in}\)/A\(_{RM}\) < 5%
whereas the existing model used:
3.2% < A\(_{in}\)/A\(_{RM}\) < 21%
The room lengths for which these models are expected to be applied range between the smallest room size of 4m and the maximum achievable CV room flushing length of 15-18m.
Air flows into the room through an inflow window in the form of an approximately axisymmetric jet, with characteristic diameter L = A\(_{in}\)\(^{1/2}\), that is confined by the room surfaces (see Figure 6). The initial portion of the jet (up to six diameters) is called the potential core and is characterized by shear layer development along the perimeter of the jet. After this phase the jet transitions into a self similar profile whose centerline velocity decays as 1/x’, where x’ = x/D (Awbi, 2003).
For the relevant room length range the geometries used in the two models result in:
New model: 6 < x’ < 60
Existing model: 2 < x’ < 18
The smaller openings used in the new model result in significantly more space for jet development along the room length. In all but the shortest rooms the jet will enter the fully developed stage characterized by a 1/x’ centerline velocity decay rate. For this reason, the updated model uses the average jet velocity profile as opposed to shear layer momentum transfer.
Components of the flow[LINK]
Figure 5 shows the schematic development of a free axisymmetric jet in an unconfined environment, and Figure 6 depicts the confinement effects. The air entrained by the jet is detrained at the outlet, creating two recirculation zones that do not exist in a free jet.
The recirculation flow is similar to the flow in a lid driven cavity, where here the driving force is the confined jet flow. Turbulent lid driven cavity flows have been shown to have self-similar velocity profiles, as shown in Figure 7.
The model is based on the hypothesis that the flow is composed of:
A – Lid driven cavity flows (in the recirculation regions).
B – A confined axisymmetric jet flow.
The goal of the model is to obtain the value and functional form of the constant term shown in the scaling laws for the characteristic average velocities in the two flow regions: jet and recirculation regions.
To estimate the average jet velocity we integrate along the depth of the room (including both the core region and the 1/x’ velocity decay region) to derive an average maximum jet velocity, V\(_{J,m}\):∙
\[{V_{J,m}} = {U_{in}}\sqrt {{A_{in}}} \left\{ {6 + 6.3\ln \left( {{D^*}/6\sqrt {{A_{in}}} } \right)} \right\}/{D^*}\]
(To allow for wind angle effects the room depth D has been replaced by D* - see a later section.) The lid-driven cavity flow profile that characterizes the recirculation region has a maximum velocity that is approximately one half of the driving velocity, which in this case is due to the jet. For this reason we define the separation between the two regions along the virtual surface where the jet reaches 50% of its maximum, centerline, velocity.
The analysis of the flow revealed that the room containment effect leads to higher velocities for cases that combine a large inflow opening with a small room cross-section area. In order to account for this effect we propose to use the following non-dimensional scaling parameter:
\[{C_{RM}} = {\left( {{A_{in}}/{A_{RM}}} \right)^{1/2}}\]
The expression used for the scaling of the average jet velocity is then:
\[{U_J} = {C_1}{C_{RM}}{V_{J,m}}\]
or
\[{U_J} = {C_1}\frac{{{Q_{in}}}}{{\sqrt {{A_{in}}{A_{RM}}} }}\left( {{V_{J,m}}/{U_{in}}} \right)\]
where the constant C\(_{1}\) will be obtained from the results of a set of CFD simulations shown in the next section.
In addition to the jet velocity the characteristic recirculation flow velocity \({U_R}\) is also an interesting parameter that, according to the self-similar flow profile hypothesis shown in Figure 7, should scale with a similar expression to the one shown above for the jet (although with a different correlation constant, C\(_{2}\)). The modeling hypothesis can also be tested by evaluating its capability to predict the maximum recirculation flow rate, which can be predicted by multiplying the average velocity by the room cross-section area:
\[{Q_R} = {C_3}\frac{{{Q_{in}}}}{{{C_{RM}}}}\left( {{V_{J,m}}/{U_{in}}} \right)\]
Because of the self-similar velocity profile shape that occurs in the recirculation, the near wall velocity (used to predict forced convection) can be estimated simply by multiplying the characteristic recirculation velocity by two.
The expressions obtained apply in the turbulent regime (because the CFD simulations on which they are based are for that regime). Both velocity correlations presented below have the functional form:
\[V = aX + b\]
where a is C\(_{1}\), C\(_{2}\) or C\(_{3}\), X is a velocity scale and b is a constant that results from the best fit. The model has a lower limit on the value of X: if X is too small the flow is no longer turbulent. This limit implies that the point X = 0 is never achieved, avoiding the unrealistic prediction of V(0) = b.
In addition to the velocity and airflow rate predictions described above it is also essential to predict the temperature variations in the two zones of the flow. In most cases these variations will be positive due to the internal sensible heat gains. However’, as a result of room surface heat transfer, in some cases the variations may be negative. Analysis of the CFD simulations that are presented in the next section revealed that approximately all the exhaust air comes from the jet regions. For this reason, the heat gains affect the two regions differently:
Heat gains in the jet region only affect this region (no temperature increase in the recirculation regions).
Heat gains in the recirculations affect both jet and recirculation regions (because they must be exhausted by the jets).
In order to predict the temperature increase, heat gains were inserted into the recirculation regions for different room geometries in the set of simulations presented in the next section. The volume containing the heat gains extended over the whole depth of the room, from floor to ceiling, and in the lateral direction from halfway between the window edge and the wall all the way to the wall. Applying energy conservation principles to predict the temperature increase in the jet and recirculation region as a function of the sensible heat gains in each region (the sum of internal gains and internal surface heat transfer) we can obtain the average value of temperature increase in each region:
\[\Delta {T_J} = {C_{T,J}}\frac{{q'{'_{RM}}}}{{\rho {c_p}{Q_{in}}}}\]
\[\Delta {T_R} = {C_{T,R}}\frac{{q'{'_{RM}}}}{{\rho {c_p}{Q_{in}}}}\]
where q”\(_{RM}\) is the total internal heat gain for the room, \(\rho\) and c\(_{p}\) are the density and specific heat capacity of air, respectively, and C\(_{T,J}\), C\(_{T,R}\) are constants.
Results[LINK]
Table 2 lists the cases used to develop the updated CV model, which predicts the output variables defined in Table 3 (see also Figure 8). Figures 9, 10, and 11 show plots of the data derived from the CFD simulations of these test cases, and the lines giving the best fit to the data. Table 4 summarizes the correlation formulae and constants. Finally, Table 5 provides estimates for the error of the correlation in each test case.
See also Figure 8. Note the x-coordinate is along the room, the y-coordinate is across the room and the z-coordinate is vertical.
In each case the error is the percentage difference between the given correlation prediction and the test case result obtained from CFD, i.e. 100*(correlation prediction – CFD)/CFD.
Wind angle effects[LINK]
When the incoming wind is not normal to the façade the jet enters the room at an angle, as shown in Figure 12.
This oblique geometry has two main effects on the flow and consequent correlation predictions:
The inflow area is reduced, creating a larger inflow velocity for the same flow rate.
The room flow path length is increased, from D to D*, resulting in a larger apparent depth.
The first effect increases indoor velocities while the second effect decreases the velocities.
Multiple apertures[LINK]
CFD simulations for rooms with more than one inflow aperture show that the model can be applied to these configurations with minimal adjustments. As a result of increased dissipation in the central recirculation region, where two distinct recirculation flows collide (Figure 13), the airflow velocities are reduced by 5-10%, as shown in the charts in Figure 14.
Heat transfer in recirculating flows
In most cross ventilation flows with recirculations the mixing between recirculation and inflow jet is only partial. Figure 15 shows a schematic representation of the heat transfer process (showing a top view of one half of a symmetrical room). The flow is divided into three distinct streams with connected temperature variations:
The main jet (labeled J in the figure).
The part of the recirculation flow that exchanges heat with the jet (label R).
The wall boundary layer part of the recirculation flow (label W).
The light gray arrows show flow direction. The dark gray arrows show heat transfer in the shear layer. The recirculation region coordinate system is shown in the figure, with coordinate r varying between 0 and L.
Heat transfer in a shear layer[LINK]
It is possible to use a formulation with a heat transfer area, a convection heat transfer coefficient and an average shear layer temperature difference to model the heat exchange between jet and recirculation flows:
\[{Q_{y = 0}} = {A_{SL}}{h_{SL}}\Delta {T_{SL}}\]
\[{A_{SL}} = L P\]
\[{h_{SL}} = \frac{{\rho \,{C_P}\,F}}{{{C_D}\,\,{A_{IN}}\,\,4\,\sigma \,\sqrt \pi }}\]
Locally, the heat transfer process is driven by the variable, local, temperature difference:
\[\rho \,{C_P}\,F\,\frac{{\partial \,{T_J}(x)}}{{\partial \,x}} = - \,\frac{{{A_{SL}}\,{h_{SL}}}}{L}\,\left( {{T_J}(x) - {T_R}(x)} \right) = \frac{{P\,\,\left( {{T_J}(x) - {T_R}(x)} \right)}}{{{C_D}\,{A_{IN}}\,4\,\sigma \,\sqrt \pi }}\]
Predicting air temperature in the recirculation region of the flow[LINK]
An accurate representation of system behavior must consider both the local heat transfer in the shear layer (varying along the x direction, driven by the temperature profiles T\(_{J}\)(x) and T\(_{R}\)(x)) and the heat transfer in the wall boundary layers (with temperature T\(_{W}\)(r)). Both transfers can be modeled using differential equations, resulting in the following system of equations:
\[\begin{array}{rl} \rho C_p F R \frac{\partial T_W \left(r\right)}{\partial r} & = \frac{A_S h_S}{L} \left(T_S - T_W \left(r\right)\right) \\ \rho C_p F R \frac{\partial T_R \left(x\right)}{\partial x} & = \frac{A_{SL} h_{SL}}{L} \left(T_J\left(x\right) - T_R \left(x\right)\right) \\ \rho C_p F \frac{\partial T_J \left(x\right)}{\partial x} & = \frac{A_{SL} h_{SL}}{L} \left(T_R\left(x\right) - T_J \left(x\right)\right) \end{array} \label{eq:DiffEqForTWrTRxTJx}\]
The relevant boundary conditions are:
\[{T_W}({L_{}}) = {T_R}(0)\,,\,{T_W}(0) = {T_R}(L)\,,\,\,{T_J}(0) = {T_{IN}}\]
The first equation in Equation [eq:DiffEqForTWrTRxTJx] models heat transfer in the wall layer. The second models the temperature variation of the portion of the recirculation flow that is in contact with the jet, it differs from the third equation only in a sign (the temperature variation in T\(_{R}\) is symmetric to the variation in T\(_{J}\)) and the flow rate (F\(\cdot\)R for the recirculation and F for the main jet). The temperature in the recirculation region is given by:
\[{T_W}(r) = {T_S} - ({T_S} - {T_{IN}})\,\frac{{{e^{\frac{{{h_S}\,{A_S}}}{{\rho \,{C_P}\,R\,F}}\frac{r}{{{L_{}}}}}}\,({e^{ - \frac{{{A_{SL}}\,{h_{SL}}\,(1 + R)}}{{\rho \,{C_P}\,R\,F}}}} - 1)}}{{{e^{ - \frac{{{A_{SL}}\,{h_{SL}}\,(1 + R)}}{{\rho \,{C_P}\,R\,F}}}} - {e^{\frac{{{h_S}\,{A_S}}}{{\rho \,{C_P}\,R\,F}}}}\,(1 + R) + R}}\]
Heat transfer in recirculating flows with internal gains[LINK]
Whereas surface heat transfer is an important component of room heat transfer, internal gains often dominate the room temperature field. Referring to the previously used subdivision of the room into jet and recirculation regions we can see that gains can occur exclusively in the jet region, exclusively in the recirculation region or in both regions.
Heat gains in the jet region[LINK]
The effect of gains in the jet region on room air temperature distribution is simple to model and, within the first order accuracy goal, can be characterized as follows:
For gains occurring in the jet region (between the inlet and the room outlet) a change in inflow temperature is an adequate, conservative approximation. From energy conservation we conclude that the altered inflow temperature is given by:
\[{T_{ING}} = {T_{IN}} + \frac{{{G_J}}}{{\rho {C_P}F}} \label{eq:AlteredInflowTempEquation}\]
Gains occurring in the jet region, close to the outlet can be ignored in a first order accuracy model.
Clearly the two approaches described above are only exact when: the gains occur at the inlet and perfectly mix with the inflow jet (for the first case), or whenever the gains occur very close to the outlet (in the second case). In all other situations, the first approach provides a conservative approximation. In the model implementation in EnergyPlus the user will not define the exact location of the heat gains within each region of the flow therefore, heat gains in the jet region, are inserted in two equal parts in two separate points of the jet flow (in the room entrance and before the exit, see points r = 0 and r = L\(_{R}\) in Figure 15).
Heat gains in the recirculation region[LINK]
When heat gains occur in the recirculation region significant heat accumulation occurs due to the limited heat transfer ability of the shear layer. Due to this limited ability, the recirculation is partially isolated from the main jet flow and higher temperatures are generated in this region whenever heat gains are present. The concept of a global heat transfer coefficient is not applicable.
In the case of a flow in a room with adiabatic surfaces and heat gains in the recirculation region the indoor air temperatures can be predicted using the solution to the following system of equations:
\[\begin{array}{rl} \rho C_p F R \left(T_R\left(0\right) - T_R\left(L\right)\right) & = G_R \\ \rho C_p F R \frac{\partial T_R \left(x\right)}{\partial x} & = \frac{A_{SL} h_{SL}}{L} \left(T_J\left(x\right) - T_R \left(x\right)\right) \\ \rho C_p F \frac{\partial T_J \left(x\right)}{\partial x} & = \frac{A_{SL} h_{SL}}{L} \left(T_R\left(x\right) - T_J \left(x\right)\right) \end{array} \label{eq:TRoTRLdTRxdTJxSystem}\]
This system of equations differs from the system shown in Equation [eq:DiffEqForTWrTRxTJx] in several ways. Because there is no wall heat transfer there is no need to split the recirculation flow in two parts (R and W). The temperature increase in the recirculation flow is simply defined by the internal gains (first equation in Equation [eq:TRoTRLdTRxdTJxSystem]). The second and third equations in Equation [eq:TRoTRLdTRxdTJxSystem] model shear layer heat transfer and are similar in the two cases (Equation [eq:DiffEqForTWrTRxTJx] and Equation [eq:TRoTRLdTRxdTJxSystem]).
The recirculation flow temperature at x = 0 is given by:
\[{T_R}(0) = {T_{IN}} + \left( {\frac{1}{{{e^{\frac{{{A_{SL}}{h_{SL}}\,(1 + R)}}{{\rho \,{C_P}F\,R}}}} - 1}} + 1} \right)\frac{{{G_R}(1 + R)}}{{\rho \,{C_P}F\,R}}\]
The temperature in the recirculation is inversely proportional to the recirculation flow ratio (R) and the shear layer area (measured by the product: A\(_{SL}\) h\(_{SL}\)). As both of these parameters increase this expression becomes similar to Equation [eq:AlteredInflowTempEquation]. The recirculation flow temperature at x = L is given by:
\[{T_R}(L) = {T_R}(0) + \Delta {T_R}\]
The temperature variations in the recirculation flow is given by:
\[\Delta {T_R} = \frac{{{G_R}}}{{\rho \,{C_P}\,F\,R}}\]
Combined effects of surface heat transfer and internal gains in the recirculation region[LINK]
In this case, the following system of equations must be solved:
\[\begin{array}{rl} \rho C_p F R \frac{\partial T_W \left(r\right)}{\partial r} & = -\frac{A_S h_S}{L} \left(T_S - T_W \left(r\right)\right) - \frac{G_R}{L_R} \\ \rho C_p F R \frac{\partial T_R \left(x\right)}{\partial x} & = \frac{A_{SL} h_{SL}}{L} \left(T_J\left(x\right) - T_R \left(x\right)\right) \\ \rho C_p F \frac{\partial T_J \left(x\right)}{\partial x} & = \frac{A_{SL} h_{SL}}{L} \left(T_R\left(x\right) - T_J \left(x\right)\right) \end{array}\]
With the boundary conditions: \({T_W}(L) = {T_R}(0)\,,\,{T_W}(0) = {T_R}(L)\,,\,\,{T_J}(0) = {T_{IN}}\)
For simplicity the heat gains are considered to be evenly distributed along the recirculation path. The temperature variation in the recirculation region is given by:
\[{T_W}(r) = {T_S} + \frac{{{G_R}}}{{{A_S}\,{h_S}}} - \frac{{{e^{\frac{{{h_S}\,{A_S}}}{{\rho \,{C_P}\,R\,F}}\frac{r}{{{L_R}}}}}\,({e^{ - \frac{{{A_{SL}}\,{h_{SL}}\,(1 + R)}}{{\rho \,{C_P}\,R\,F}}}} - 1)\,({G_R} + {A_S}\,{h_S}\,({T_S} + {T_{IN}}))}}{{{e^{ - \frac{{{A_{SL}}\,{h_{SL}}\,(1 + R)}}{{\rho \,{C_P}\,R\,F}}}} + R - {e^{\frac{{{h_S}\,{A_S}}}{{\rho \,{C_P}\,R\,F}}}}\,(1 + R)}}\]
RoomAirflowNetwork Model[LINK]
Overview[LINK]
The RoomAirflowNetwork integrates the RoomAir model with the AirflowNetwork model. The model allows multiple Room Air nodes to be associated with AirflowNetwork intra zone nodes. The AirflowNetwork intra zone linkages provide links among the intra zone nodes and users may specify airflow components among these links. The AirflowNetwork model calculates airflows of the links. The incoming airflows are part of the zone heat and moisture balances of intra zone nodes. The model also allows users to specify a node to be connected to surfaces to have convective heat transfer and/or moisture transfer between surfaces and the node, portion of internal gains and supply air and return air fractions from zone equipment and AirLoop terminals.
The input object RoomAirSettings:AirflowNetwork lists multiple RoomAirflowNetwork nodes in a zone. The input object of RoomAirflowNetwork:Node defines a list of connections assigned to a particular RoomAirflowNetwork node so that the connections will be used as components of node heat and moisture balance equations. In addition, the input objects AirflowNetwork:IntraZone:Node and AirflowNetwork:IntraZone:Linkage specify airflow nodes, linkages, and associated flow components, so that the AirflowNetwork model is used to calculate intrazone linkage airflows, which will be a part of loads for the RoomAir node balances.
Model description[LINK]
The model calculates each component of the governing equations first. Then heat and moisture balance equations are solved to get the node temperature and humidity ratio at the current time step.
Governing equations for each RoomAirflowNetwork node are described below:
Energy balance equation
\[\begin{split} {F_i}{C_{z,i}}\frac{{d{T_{z,i}}}}{{dt}} =& \sum\limits_{j = 1}^{{N_{i,g}}} {{F_{gain,j}{\dot Q}_{i,j}}} + \sum\limits_{j = 1}^{{N_{i,sur}}} {{h_{i,j}}} {A_j}\left( {{T_{s,j}} - {T_{z,i}}} \right) + \sum\limits_{j = 1}^{{N_{i,AFN}}} {{{\dot m}_{AFN i,j}}} {C_p}\left( {{T_{z,j}} - {T_{z,i}}} \right)\, \\ &+ \sum\limits_{j = 1}^{{N_{i,HVAC}}} {{{F_{HVAC,j}}{\dot m}_{sup,i,j}}} {C_p}\left( {{T_{sup_j}} - {T_{z,i}}} \right) \end{split}\]
where:
F\(_{i}\) = Fraction of zone air volume for the ith node, defined in the Fraction of Zone Air Volume field of the RoomAir:Node:Airflownetwork
C\(_{z,i}\) = Volumetric heat capacity for the ith node: \(\rho_{air,i}\)V C\(_{p}\)C\(_{T}\)
\(\rho_{air,i}\) = Dry air density at the ith zone node
C\(_{p}\) = zone dry air specific heat
C\(_{T}\) = sensible heat capacity multiplier (Detailed description is provided below)
V = Zone air volume
T\(_{z,i}\) = Air temperature at the ith zone RoomAir node
N\(_{i,g}\) = The number of internal gains for the ith node, defined in the RoomAir:Node:Airflownetwork:InternalGains object
F\(_{gain,j}\) = Fraction of internal gain from jth internal gain object defined in the RoomAir:Node:Airflownetwork:InternalGains object
\({\dot Q}_{i,j}\) = The amount of internal sensible gain at the jth internal gain object and ith node.
\({\sum\limits_{j = 1}^{N_{i,g}} F_{gain,j}{\dot Q}_{i,j}}\) = The total amount of internal sensible gains assigned to the ith node
N\(_{i,sur}\) = The number of adjacent surface with convective heat transfer for the ith node, defined in the RoomAir:Node:Airflownetwork:AdjacentSurfaceList object
h\(_{i,j}\) = Heat transfer coefficient between the jth adjacent surface and ith node
A\(_{j}\) = Area of the jth adjacent surface
T\(_{s,j}\) = Inside temperature at the jth adjacent surface
\({\sum\limits_{j = 1}^{N_{i,sur}} h_{i,j} A_j \left( T_{s,j} - T_{z,i}\right)}\) = The total amount of convective gains from adjacent surfaces attached to the ith node
N\(_{i,AFN}\) = The number of RoomAir:Node:Airflownetwork objects connected to the ith node. The links are defined in AirflowNetwork:IntraZone:Linkage objects.
\({\dot m}_{AFN,i,j}\) = Mass flow rate from the jth node to the ith node. The flow rates are calculated in the AirflowNetwork model. It should be pointed out that the outgoing flows from the ith node will not be used in the present energy and moisture balance calculations, because they are a part of energy and moisture balances.
T\(_{z,j}\) = Air temperature at the jth zone RoomAir node
\({\sum\limits_{j = 1}^{N_{i,g}} \dot m_{AFN i,j} {C_p}\left( T_{z,j} - T_{z,i} \right)}\) = The total amount of heat transfer driven by intra zonal flows
N\(_{i,HVAC}\) = The number of RoomAir:Node:Airflownetwork:HVACEquipment objects which provide supply air to the ith node.
F\(_{HVAC,j}\) = Fraction of output or supply air to the ith node from jth HVAC equipment defined in RoomAir:Node:Airflownetwork:HVACEquipment objects
\({\dot m}_{sup,i,j}\) = The supply mass flow rate from the jth HVAC equipment
T\(_{sup,j}\) = Supply air temperature at the jth HVAC equipment
\({\sum\limits_{j = 1}^{N_{i,HVAC}} F_{HVAC,j} \dot m_{sup,i,j} \left( T_{sup,j} - T_{z,i}\right)}\) = The total amount of heat transfer from HVAC equipment to the ith node
Moisture balance equation
\[\begin{split} {F_i}{M_{z,i}}\frac{{d{W_{z,i}}}}{{dt}} =& \sum\limits_{j = 1}^{{N_{i,g}}} {{F_{gain,j}{\dot Q}_{i,j}}} + \sum\limits_{j = 1}^{{N_{i,sur}}} {{h_{M,j}}} {A_j}\left( {{W_{s,j}} - {W_{z,i}}} \right) + \sum\limits_{j = 1}^{{N_{i,AFN}}} {{{\dot m}_{AFN i,j}}} \left( {{W_{z,j}} - {W_{z,i}}} \right)\, \\ &+ \sum\limits_{j = 1}^{{N_{i,HVAC}}} {{{F_{HVAC,j}}{\dot m}_{sup,i,j}}} \left( {{W_{sup_j}} - {W_{z,i}}} \right) \end{split}\]
where:
Fi = Fraction of zone air volume for the ith node, defined in the Fraction of Zone Air Volume of the RoomAir:Node:Airflownetwork
M\(_{z,i}\) = Mass for the ith node: \(\rho_{air,i}\)V\(_{Cp}\)C\(_{T}\)
\(\rho_{air,i}\) = Dry air density at the ith zone RoomAir node
V = Zone air volume
W\(_{z,i}\) = Air humidity ratio at the ith zone RoomAir node
N\(_{i,g}\) = The number of internal gains for the ith node, defined in the
RoomAir:Node:Airflownetwork:InternalGains object
F\(_{gain,j}\) = Fraction of internal gain from jth internal gain object defined in the RoomAir:Node:Airflownetwork:InternalGains object
\({\dot Q}_{i,j}\) = The amount of internal latent gain at the jth internal gain object and ith node.
\({\sum\limits_{j = 1}^{N_{i,g}} F_{gain,j}{\dot Q}_{i,j}}\) = The total amount of internal latent gains attached to the ith node
N\(_{i,sur}\) = The number of adjacent surface with convective heat transfer for the ith node, defined in the RoomAir:Node:Airflownetwork:AdjacentSurfaceList object
h\(_{M,j}\) = Mass transfer coefficient between the jth adjacent surface and ith node
A\(_{j}\) = Area of the jth adjacent surface
W\(_{s,j}\) = Inside humidity ratio at the jth adjacent surface
\({\sum\limits_{j = 1}^{N_{i,sur}} h_{M,j} A_j \left( W_{s,j} - W_{z,i}\right)}\) = The total amount of convective moisture gains from adjacent surfaces attached to the ith node
N\(_{i,AFN}\) = The number of RoomAir:Node:Airflownetwork objects connected to the ith node. The links are defined in AirflowNetwork:IntraZone:Linkage objects.
\({\dot m}_{AFN,i,j}\) = Mass flow rate from the jth node to the ith node. The flow rates are calculated in the AirflowNetwork model. It should be pointed out that the outgoing flow from the ith node will not be used in the present calculation, because they are a part of energy balance.
W\(_{z,j}\) = Air humidity ratio at the jth zone RoomAir node
\({\sum\limits_{j = 1}^{N_{i,AFN}} \dot m_{AFN,i,j} \left( W_{z,j} - W_{z,i}\right)}\) = The total amount of moisture transfer driven by intra zonal flows
N\(_{i,HVAC}\) = The number of RoomAir:Node:Airflownetwork:HVACEquipment objects which provide supply air to the ith node.
F\(_{HVAC,j}\) = Fraction of output or supply air to the ith node from jth HVAC equipment defined in RoomAir:Node:Airflownetwork:HVACEquipment objects
\({\dot m}_{sup,i,j}\) = The supply mass flow rate from the jth HVAC equipment
W\(_{sup,j}\) = Supply air humidity ratio at the jth HVAC equipment
\({\sum\limits_{j = 1}^{N_{i,HVAC}} F_{HVAC,j} \dot m_{sup,i,j} \left( W_{sup,j} - W_{z,i}\right)}\) = The total amount of moisture transfer from HVAC equipment to the ith node
Calculation procedures[LINK]
The AirflowNetwork model is called first to calculate intrazone linkage air flows based on node temperatures and humidity ratios at the previous time step. The linkage loads are summarized based on each node with all incoming airflow, while any outgoing airflows from the node of the interest are excluded in the node energy and moisture balances. When the RoomAirModelAirflowNetowrk module is called, it summarizes internal gains at each node based on specification provided in the RoomAirflowNetwork:Node:InternalGains object, convective heat transfer from adjacent walls specified in the RoomAirflowNetwork:Node:AdjacentSurfaceList object. These load components are used in the predictor to predict system loads based on thermostat setpoint and supply fraction in the controlled zone node. In the corrector, the supply air conditions defined in the RoomAirflowNetwork:Node:HVACEquipment object are added. After all load components are assembled together, the node energy and moisture balance equations are solved. The node temperature and humidity ratio are calculated at the current time step.
References[LINK]
Altmayer, E.F., Gadgil, A.J., Bauman, F.S., Kammerud, R.C. Correlations for convective heat transfer from room surfaces. ASHRAE Transactions, No. 2764.
Awbi, H.B. 2003. Ventilation of Buildings (2nd Edition). Routledge.
Awbi, H.B. and Hatton, A. 1999. Natural convection from heated room surfaces, Energy and Buildings, 30, 233-244.
Awbi, H.B. and Hatton, A. 2000. Mixed convection from heated room surfaces, Energy and Buildings, 32, 153-166.
Aynsley, R.M., Melbourne, W., Vickery, B.J. 1977. Architectural Aerodynamics. Applied Science London.
Baturin, V. V. , Billington, N. S. 1972. Fundamentals of Industrial Ventilation, Franklin Book Company 1972, pp. 174-179.
Bejan, A. 1994. Convection Heat Transfer 2nd ed, Wiley, USA.
Carrilho da Graça, G. 2003. Simplified models for heat transfer in rooms. Ph. D. Thesis, University of California, San Diego.
Chandra, S., Kerestecioglu, A.A. 1984. Heat transfer in naturally ventilated rooms data from full-scale measurements, ASHRAE Transactions, Volume 90, part 1b 211-224.
Chen, Q., and B. Griffith. 2002. Incorporating Nodal Room Air Models into Building Energy Calculation Procedures. ASHRAE RP-1222 Final Report.
Cooper, P. and P.F. Linden. 1996. Natural ventilation of an enclosure containing two buoyancy sources. Journal of Fluid Mechanics, Vol. 311, pp. 153-176.
Dominique, M. and P. Guitton. 1997. Validation of displacement ventilation simplified models. Proc. of Building Simulation.
Griffith, B. and Q. Chen. 2004. Framework for coupling room air models to heat balance load and energy calculations (RP-1222). International Journal of Heating, Ventilating, Air-conditioning and Refrigerating Research. ASHRAE, Atlanta GA. Vol 10. No 2. April 2004.
Holford, J.M., G.R. Hunt and P.F. Linden. 2002. Competition between heat sources in a ventilated space. Proceedings of RoomVent 2002, pp. 577-580.
Hunt, G.R., J.M. Holford and P.F. Linden. 2002. Characterization of the flow driven by a finite area heat source in a ventilated enclosure. Proceedings of RoomVent 2002, pp. 581-584.
Hunt, G.R. and P.F. Linden. 2001. Steady-state flows in an enclosure ventilated by buoyancy forces assisted by wind. . Journal of Fluid Mechanics, Vol. 426, pp. 355-386.
Kalema, T., Haapala, T. ’995. Effect of interior heat transfer coefficients on thermal dynamics and energy consumption, Energy and Buildings 22 (2) pp. 101-113.
Kaye, K.N. and P.F. Linden. 2004. Coalescing axisymmetric turbulent plumes. Journal of Fluid Mechanics, Vol. 502, pp. 41–63.
Lin, Y.J. and P.F. Linden. 2005. A model for an under floor air distribution system. Energy&Building, Vol. 37, pp. 399-409.
Linden, P.F., G.F. Lane-Serff and D.A. Smeed. 1990. Emptying filling boxes: the fluid mechanics of natural ventilation. Journal of Fluid Mechanics, Vol. 212, pp. 309-335.
Linden, P.F. and P. Cooper. 1996. Multiple sources of buoyancy in a naturally ventilated enclosure. Journal of Fluid Mechanics, Vol. 311, pp. 177-192.
Liu, Q. 2006. The Fluid Dynamics of an Underfloor Air Distribution System. A PhD dissertation in Engineering Sciences (Systems Science) at UC San Diego.
Morton, B.R., G.I. Taylor andJ.S. Turner. 1956. Turbulent gravitational convection from maintained and instantaneous sources. Proceedings of the Royal Society of London, Vol A234, pp. 1-23.
Mundt, E. 1996. The Performance of Displacement Ventilation Systems – Experimental and Theoretical Studies, Ph. D. Thesis, Bulletin N38, Building Services Engineering KTH, Stockholm.
Neiswanger, L., Johnson, G.A., Carey, V.P. 1987. An experimental study of high Raleigh number mixed convection in a rectangular enclosure with restricted inlet and outlet openings. Transactions of ASME, Vol. 109, 446-453.
Ohba, M., Irie, K., Kurabuchi, T. 2001. Study on airflow characteristics inside and outside a CVmodel, and ventilation flow rate using wind tunnel experiments. Journal of Wind Engineering and Industrial Aerodynamics, in press.
Rees, S.J., and P. Haves. 2001. A nodal model for displacement ventilation and chilled ceiling systems in office spaces. Building and Environment, Vol. 26, pp. 753-762.
Siebers, D. L., Schwind, R. G., Moffat, R. J. 1983. Experimental Mixed Convection Heat Transfer From a Large Vertical Surface in a Horizontal Flow. SAND 83-8225, Sandia National Laboratories, Livermore CA.
Spitler, J. D., Pedersen, C.O., Fisher, D.E. 1991. Interior Convective Heat Transfer in Buildings with Large Ventilative Flow Rates. ASHRAE Transactions, Vol. 97, Pt.1, pp. 505-515.
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