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Question about the industrial building pressurization 2
- Thread starter lzh007
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10% of AHU cfm would be a better bet than cfm/person, since the occupancy per sqft is much lower these days in industries, given the high level of mechanisation and computerisation.
However, if a specific pressuristion level is to be maintained, then it's a different calculation and if I remember right, it was widely discussed in one of the thread before in the same forum.
HVAC68
However, if a specific pressuristion level is to be maintained, then it's a different calculation and if I remember right, it was widely discussed in one of the thread before in the same forum.
HVAC68
- Thread starter
- #5
I know around 10% Min OA is a rule for commercial and residential buildings, does it apply to industrial building?
Because the air flow for some industrial room is very big, 10% can still need lots of energy to heat. ( For example the unit provides ventilation in summer only).
BTW, I did a search, but can not find the old discussion about the pressurization.
Because the air flow for some industrial room is very big, 10% can still need lots of energy to heat. ( For example the unit provides ventilation in summer only).
BTW, I did a search, but can not find the old discussion about the pressurization.
Whats the building used for. How big is it. How old is it. What standard is it built to? Whats the occupancy. etc.
If its a leaky old building, then it will already have a good amount of vent.
10% fresh air might be over the top, but really it depends on what is being undertaken inside.
Is the atmosphere polluted by process or is it a store?
Give us a clue
Friar Tuck of Sherwood
If its a leaky old building, then it will already have a good amount of vent.
10% fresh air might be over the top, but really it depends on what is being undertaken inside.
Is the atmosphere polluted by process or is it a store?
Give us a clue
Friar Tuck of Sherwood
- Thread starter
- #8
friartuck (Mechanical), sorry for my late reply.
This is a new industrial building, lots of heat rejection inside but no polltion and people inside. So Ventilation or AC will be needed in summer, Ventilation and heating needed in winter.
lilliput1 (Mechanical)
Rule of the thumb is .05 CFM/SF.
What's the number for SF? Is it the area of the floor space or the outside envelope of the builing?
This is a new industrial building, lots of heat rejection inside but no polltion and people inside. So Ventilation or AC will be needed in summer, Ventilation and heating needed in winter.
lilliput1 (Mechanical)
Rule of the thumb is .05 CFM/SF.
What's the number for SF? Is it the area of the floor space or the outside envelope of the builing?
SF is the floor space area. The 0.05 CFM/SF is for a typical 14' floor to floor height office. You may factor it up based on how high & loose the building is compared to the typical office. It is not practical to pressurize the building completely to prevent winter infiltration. You should have heated vestibules at main entrance & local heaters at doorways, large glass areas.
Often the exhaust volume determines the make-up volume. Certain areas (such as restrooms and kitchens) and processes will cumulatively require some total amount of exhaust air. The minimum make-up air should exceed this value by 5-10 percent, OR should be enough to meet ventilation requirements for people or for the type of space (the higher of the two). When in order to offset the exhaust the make-up volume becomes excessive, you should think about using heat recovery (an enthalpy wheel or a run-around loop).
Exhaust volume or cfm per person should determine the MINIMUM make-up air. High CO2 and enthalpy economizer are functions that may increase ventilating air above minimum... Either high CO2 or economizer will modulate AHU dampers to take in more outside air and return less air. This will not change building pressure, however, for a typical supply/return system.
Hmmmm. Good question.
Typicaly one wants to maintain some minimum overpressurization of building to minimize infiltration. In some cases, folks only consider keeping too cold, or too warm air from "leaking" into building. But in truth one is also concerned with infiltrating humidity and dust/dirt, and other things.
I've seen a number of schemes used. Including rather complicated methods of calculating makeup air requirements to account for losses due to exhaust fans running. ie Total up exhaust fan CFMs. Then add the amount (for instance, 10%)for fresh OA intake. Plus maybe a little extra just to be sure. And arrive at some fixed number of OA.
The above can work. But tends to work poorly.
Also, some schemes use exhaust/relief fan (associated with AHU)modulation based on some fixed number. ie relief fan speed lags supply fan speed by 10%. (And/or same control scheme applied to dampers. Relief damper modulated open based on fixed number which is a lag behind OA damper modulating open.)
Ditto, can work, but not well.
Problem with all the fixed number methods is it requires careful calculation. Assumes exact number of exhaust fans are on and that each is putting out rated CFM. Etc. Etc.
If extra exhausts are started up that were forgotten during initial calculations, or another exhaust fan added later. Building becomes negative.
If exhaust fan (sizeable one) is shut down due to lack of need, or it just plain fails, or it's belt is slipping, or whatever, building gets pressurized too much.
Not good either. You're wasting energy conditioning more OA intake than yah need. Plus doors blow open. Are hard to shut.
My preferred method, and the one that I've seen working best as it accounts for variations that're normal in most buildings. Is a separate control point connected to a building (zone) space static sensor. Two sensing legs on sensor. One sensing pressure in space served by AHU. The other leg tubed to some point outside the building walls. Controls compare the two. Modulate separate releif dampers/relief fan for building (zone)that're independent of AHU. Or, modulate relief fan and dampers tee'd off of return ducting of AHU.
As building pressure, relative to outside, goes up, relief dampers open farther, and at some point (in my experience, somewhere around 60 to 70%) of opening, relief fan starts and starts incremental climb in speed as needed to reduce building pressure. As building (zone) pressure drops, relief fans slows further and further. At some point determined by fan performance curve, or simply by trial and error observation, yah stop the fan. Then rely on relief dampers to control pressure. (Relief dampers lagging relief fan as regards to shutting down.)
Typically, for ordinary buildings, we try for a positive .05 inch pressure. Enough to prevent the "majority" of infiltraion. Tho it won't stop it on especially windy days.
In some installations, setpoint is different. ie We have a customer that's a medical equipment manufacturer with clean rooms, etc. They have their own very definite requirements. A fairly high overpressure for building in general. Plus even higher, incrementally so, in various clean rooms.
As regards CO2. Specs and control schemes I've seen simply monitor either individual room/space, or return air, and nothing is done if CO2 below certain point. Then as CO2 level hits set point, if it does, OA minimum air requirement is incremented, slowly, in an attempt to lower CO2 below set point plus deadband amount.
You say you're in Canada? I'm in Minnesota. If you're gonna use this sort of scheme, best to have pre-heat coil to knock chill out of OA air. Better yet, put in an ERU (energy recovery unit).
Sorry, can't tell yah a lot more. I'm not a system or building designer. Just a controls guy. I have done some design work, but not a lot. Good HVAC designer would know more of the specifics than I would. Most times I just get specs and requirements from designer for what needs to happen. Then I do my thing of designing control scheme that'll accomplish designer's goals.
It's my observation that HVAC system designers usually don't know all that much about specific equipment and controls. But, OTOH, I don't know that much about all the math, design criteria, buildings codes, etc that it takes to design a good system. That's why, in-house, in the company for whom I work, we have different guys doing these things.
Typicaly one wants to maintain some minimum overpressurization of building to minimize infiltration. In some cases, folks only consider keeping too cold, or too warm air from "leaking" into building. But in truth one is also concerned with infiltrating humidity and dust/dirt, and other things.
I've seen a number of schemes used. Including rather complicated methods of calculating makeup air requirements to account for losses due to exhaust fans running. ie Total up exhaust fan CFMs. Then add the amount (for instance, 10%)for fresh OA intake. Plus maybe a little extra just to be sure. And arrive at some fixed number of OA.
The above can work. But tends to work poorly.
Also, some schemes use exhaust/relief fan (associated with AHU)modulation based on some fixed number. ie relief fan speed lags supply fan speed by 10%. (And/or same control scheme applied to dampers. Relief damper modulated open based on fixed number which is a lag behind OA damper modulating open.)
Ditto, can work, but not well.
Problem with all the fixed number methods is it requires careful calculation. Assumes exact number of exhaust fans are on and that each is putting out rated CFM. Etc. Etc.
If extra exhausts are started up that were forgotten during initial calculations, or another exhaust fan added later. Building becomes negative.
If exhaust fan (sizeable one) is shut down due to lack of need, or it just plain fails, or it's belt is slipping, or whatever, building gets pressurized too much.
Not good either. You're wasting energy conditioning more OA intake than yah need. Plus doors blow open. Are hard to shut.
My preferred method, and the one that I've seen working best as it accounts for variations that're normal in most buildings. Is a separate control point connected to a building (zone) space static sensor. Two sensing legs on sensor. One sensing pressure in space served by AHU. The other leg tubed to some point outside the building walls. Controls compare the two. Modulate separate releif dampers/relief fan for building (zone)that're independent of AHU. Or, modulate relief fan and dampers tee'd off of return ducting of AHU.
As building pressure, relative to outside, goes up, relief dampers open farther, and at some point (in my experience, somewhere around 60 to 70%) of opening, relief fan starts and starts incremental climb in speed as needed to reduce building pressure. As building (zone) pressure drops, relief fans slows further and further. At some point determined by fan performance curve, or simply by trial and error observation, yah stop the fan. Then rely on relief dampers to control pressure. (Relief dampers lagging relief fan as regards to shutting down.)
Typically, for ordinary buildings, we try for a positive .05 inch pressure. Enough to prevent the "majority" of infiltraion. Tho it won't stop it on especially windy days.
In some installations, setpoint is different. ie We have a customer that's a medical equipment manufacturer with clean rooms, etc. They have their own very definite requirements. A fairly high overpressure for building in general. Plus even higher, incrementally so, in various clean rooms.
As regards CO2. Specs and control schemes I've seen simply monitor either individual room/space, or return air, and nothing is done if CO2 below certain point. Then as CO2 level hits set point, if it does, OA minimum air requirement is incremented, slowly, in an attempt to lower CO2 below set point plus deadband amount.
You say you're in Canada? I'm in Minnesota. If you're gonna use this sort of scheme, best to have pre-heat coil to knock chill out of OA air. Better yet, put in an ERU (energy recovery unit).
Sorry, can't tell yah a lot more. I'm not a system or building designer. Just a controls guy. I have done some design work, but not a lot. Good HVAC designer would know more of the specifics than I would. Most times I just get specs and requirements from designer for what needs to happen. Then I do my thing of designing control scheme that'll accomplish designer's goals.
It's my observation that HVAC system designers usually don't know all that much about specific equipment and controls. But, OTOH, I don't know that much about all the math, design criteria, buildings codes, etc that it takes to design a good system. That's why, in-house, in the company for whom I work, we have different guys doing these things.
Pharmaceutical projects require pressirization control. Some do use actual space differential pressure static control but location of the reference static pressure sensor is critical. Door openings, wind direction, people traffic would screw it up causing false alarms.
Other method is active tracking of the airflows to the room with airflows out with the cntrol maintaining the differential between the two. Being not affected by door openings, this is the most prevalent system.
However what works in theory often do not work in real life. The system may work initialy but after sensor get dirty (specially airflow sensors on the unfiltered exhaust & return air), the system does not work as good. When nuisance alarms occur, people circumvents the system. Venturi type air valves (Phoenix - former Mitco valves) are less prone to this but they are best for constant volume or (2) position constant volume than airflow tracking control because there is feedback sensing of actual current airflow.
To be a good engineer you have to understand the system and know the math. Pressure is build up in the space by air going in. How much the pressure is depends on the cracks in the space envelpe for air to leak out. Now the pressure loss of air leaking out an orifice or crack is proportional to the square of the air velocity leaking out. Depending on the shape or orifice configuration (bellmouth, sharp edge, etc) there is a factor that can be applied to the velocity squared to come up with the actual pressure drop through the orifice. From this one can (I was the first to do it in my place of work) generate a table of door sizes & typical crack dimensions & come up with CFM correspoding to different space pressurizations (say increments of 0.025'wg.) What coefficient to use? I used about 1.5 based on .5 entry loss + 1.0 exit loss for flanged orifice.
Other method is active tracking of the airflows to the room with airflows out with the cntrol maintaining the differential between the two. Being not affected by door openings, this is the most prevalent system.
However what works in theory often do not work in real life. The system may work initialy but after sensor get dirty (specially airflow sensors on the unfiltered exhaust & return air), the system does not work as good. When nuisance alarms occur, people circumvents the system. Venturi type air valves (Phoenix - former Mitco valves) are less prone to this but they are best for constant volume or (2) position constant volume than airflow tracking control because there is feedback sensing of actual current airflow.
To be a good engineer you have to understand the system and know the math. Pressure is build up in the space by air going in. How much the pressure is depends on the cracks in the space envelpe for air to leak out. Now the pressure loss of air leaking out an orifice or crack is proportional to the square of the air velocity leaking out. Depending on the shape or orifice configuration (bellmouth, sharp edge, etc) there is a factor that can be applied to the velocity squared to come up with the actual pressure drop through the orifice. From this one can (I was the first to do it in my place of work) generate a table of door sizes & typical crack dimensions & come up with CFM correspoding to different space pressurizations (say increments of 0.025'wg.) What coefficient to use? I used about 1.5 based on .5 entry loss + 1.0 exit loss for flanged orifice.
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1
- #15
1HVACEngineer
Mechanical
lilliput1,
My company, specifically NorthEastern US offices, cater almost exclusively to the Pharmaceutical Market. We have tables developed to calculate the differential pressure across doors, similar to what you mentioned "table of door sizes & typical crack dimensions & come up with CFM correspoding to different space pressurizations"; based on 2003 ASHRAE Applications Handbook: Chapter 52, page 52.5, Pressurization (Smoke Control) see Equation below. Most of our clients required indiviual rooms to be set at a pressure level and the adjacent rooms to cascade from the highest (i.e. cleanest to dirtiest). Hope this equation helps you.
Eq(8):
Q = 776•C•A•?(2•?p/?) = 4005•C•A•?(?p)
Q = volumetric airflow rate, cfm
C = flow coefficient (typically 0.6 to 0.7)
A = flow area (leakage area), ft²
?p = pressure difference across flow path, in. of water
? = density of air entering flow path, lbm/ft³
My company, specifically NorthEastern US offices, cater almost exclusively to the Pharmaceutical Market. We have tables developed to calculate the differential pressure across doors, similar to what you mentioned "table of door sizes & typical crack dimensions & come up with CFM correspoding to different space pressurizations"; based on 2003 ASHRAE Applications Handbook: Chapter 52, page 52.5, Pressurization (Smoke Control) see Equation below. Most of our clients required indiviual rooms to be set at a pressure level and the adjacent rooms to cascade from the highest (i.e. cleanest to dirtiest). Hope this equation helps you.
Eq(8):
Q = 776•C•A•?(2•?p/?) = 4005•C•A•?(?p)
Q = volumetric airflow rate, cfm
C = flow coefficient (typically 0.6 to 0.7)
A = flow area (leakage area), ft²
?p = pressure difference across flow path, in. of water
? = density of air entering flow path, lbm/ft³
1HVACEngineer
Mechanical
My apologies.
Eq(8):
Q = 776•C•A•(2•delta_p/Density)^0.5 = 4005•C•A•(delta_p)^0.5
Q = volumetric airflow rate, cfm
C = flow coefficient (typically 0.6 to 0.7)
A = flow area (leakage area), ft²
delat_p = pressure difference across flow path, in. of water
Density = density of air entering flow path, lbm/ft³
Eq(8):
Q = 776•C•A•(2•delta_p/Density)^0.5 = 4005•C•A•(delta_p)^0.5
Q = volumetric airflow rate, cfm
C = flow coefficient (typically 0.6 to 0.7)
A = flow area (leakage area), ft²
delat_p = pressure difference across flow path, in. of water
Density = density of air entering flow path, lbm/ft³
If delta_P = K (FPM/400%)^2
and using 1HVACEngineer's corrected equation
K=(1/C)^2
Using C = .6 to .7
we get K = 2.78 to 2.04
versus the 1.5 I recommended. Or at k=1.5, C=0.8165
Using K =1.5 versus 2.78 or 2.04 therefore would result in
CFM through cracks of 36% to 17% more. This would be more conservative and thus say allow for miscellaneous other losses. It is important to seal the pressurized space. In addition of leaking out through door craks, air could leak through; ceiling tile,lights and through opening in partition above ceiling; receptacles and electrical conduits; gaps between floor and walls; gap around piping penetrations to the room; window cracks; ungasketed access doors & panels; etc.
and using 1HVACEngineer's corrected equation
K=(1/C)^2
Using C = .6 to .7
we get K = 2.78 to 2.04
versus the 1.5 I recommended. Or at k=1.5, C=0.8165
Using K =1.5 versus 2.78 or 2.04 therefore would result in
CFM through cracks of 36% to 17% more. This would be more conservative and thus say allow for miscellaneous other losses. It is important to seal the pressurized space. In addition of leaking out through door craks, air could leak through; ceiling tile,lights and through opening in partition above ceiling; receptacles and electrical conduits; gaps between floor and walls; gap around piping penetrations to the room; window cracks; ungasketed access doors & panels; etc.
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