A voiding 101 Serious HVAC System Design Mistakes

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AVOIDING 101 SERIOUS HVAC SYSTEM DESIGN MISTAKES

A voiding 101 Serious HVA C System Design Mistakes

Written By: Hal Finkelstein

Other Books Written By The Author: Published By The National Resource Center Variable Air Volume System Operation: A Guide to Engineering, Design & Operations, Pub #37 Contamination Control Ventilation: For Healthcare, Industrial, And Commercial Facilities, PUB #31 Ventilation For Ethylene Oxide Sterilization Systems, PUB #96 Steam Distribution And Flow, PUB #88 Avoiding 101 Serious Electrical System Design And Installation Mistakes, PUB #79 HVAC Systems For Bioterrorism Protection, PUB #33B

Published By John Wiley & Sons, Inc. Hazardous Substances In Buildings

Published By: The National Resource Center, Inc Publications Department Suite 167 1440 Coral Ridge Drive Coral Springs, Florida, 33071 Email: [email protected]

Pub #84

February, 2009

Copyright ©2009, The National Resource Center.,lnc All rights reserved. No part of the contents of this publication may be reproduced or transmitted in any form or by any means. The National Resource Center accepts no advertising, and its publications may not in part, in whole, or in any manner be used or referenced to advertise any product. The information contained herein is for education and information purposes . You are advised to seek out proper legal and engineering consult before utilizing the information herein for any specific use. The National Resource Center, its divisions, employees and officers are not responsible for quality or validity of information herein or for any adverse consequences of acting on such information.

II

Table Of Contents Chapter 1: Air Distribution 1.1. Max and Min VAV air quantities not shown on drawings. 1.2. Linear diffusers create drifts and comfort problems. 1.3. Flexible duct installation problems 1.4. Un-balanced duct branches. 1.5. Interconnecting duct loop arrangements. 1 .6. Ceiling Return problems. 1.7. Diffusers dumping air at low flow. 1.8. Light troffers causing drafts and comfort problems. 1.9. Pressure dependent VAV system problems. 1.10. VAV terminal box operates with reduced capacity. 1.11 . VAV terminal box operates with reduced capacity. 1.12. Flex duct installation problems. 1.13. Flex duct installation problems. 1.14. Diffusers create drafts and comfort problems. 1.14. Diffusers create drafts and comfort problems. 1.16. Excessive duct leakage causes problems. 1 . 17. Insufficient air flow from a VAV terminal box. 1.18. Terminal boxes do not open. 1.19. Terminal boxes do not close . 1.20. Steam Humidifiers soak ducts. 1.21 . Duct Humidifiers cause water to collect in ducts. 1.22. Smudging occurs around ceiling diffusers. 1.23. Terminal boxes and diffusers generate noise. 1.24. High humidity occurs in courthouse occupancy. 1.25. Improperly sized diffusers. 1.26. Insufficient smoke control in VAV systems.

1

3

3 4 4 5 5 8 8 8 9 9 10 12 12 14 14 15 15 15 16 21

22 23 24 24

Chapter 2: Air Movement 2.1. 2.2 . 2.3. 2.4 2.5. 2.6 2.7. 2.8. 2.9. 2.10. 2.11 . 2.12.

Return fan usage. Forward curved fans usage. Parallel fan arrangement problems. Airfoil blade type fans . Centrifugal fan usage problems. Variable pitch controls and vane-axial fans. Fan vibration , pulsating and noise. Multi-branch system and VAV fan surging. VAV supply fan surges and hunts. Filter bank diffuser plate installation problems. Vane-axial fan installation causes noise and vibration . Noisy fan operation and low air flow.

III

26 27 27 28 28 29 29

30 30 30 31 32

2 .13 2.14. 2 .15. 2.16. 2 .17. 2.18. 2.19. 2.20 . 2.21. 2.22.

Outdoor air mixing box stratification. Fan in plenum has reduced flow output. New fan installation using excessive energy. A reduction of airflow is observed in a ducted system. Fan bearings constantly fail and have a short life span. Fan motor drives constantly fail. Overall system noise. Large amounts of fan noise. Fan system can not put out design airflow. System effects cause deterioration in fan operation.

33 35 36 37 38

39 40 40 41 42

Chapter 3: Controls 3.1. 3.2. 3.3. 3.4. 3.5. 3.6. 3.7. 3.8.

3.9. 3.10. 3.11 . 3.12. 3.13. 3.14. 3.15. 3.16. 3.17.

Improperly installed airflow stations. Proportional integral control use. Velocity sensors improperly read. VAV box furthest from fan short of air. Duct static pressure controller can not adjust properly. Fan over pressurizes duct system causing damage. Supply-return fan tracking system malfunctions. Building pressure constantly varies. Return fan feedback problems create building pressure problems. Improper building pressure sensor location. VAV terminal sensor has insufficient velocity and airflow. VAV box malfunctions on minimum airflow. Pneumatic controllers have chaotic operation. Fan adjustment control in a VAV system constantly hunts. Terminal boxes continuously hunt. Control configurations affect VAV system operations. Velocity pressure signal control creates fan operation problems.

46 46 46

47 48 48 48 49

50 52 52 53

55 56 56

58 59

Chapter 4: Equipment (Boilers, Chillers, Cooling Towers, etc.) 4.1 . 4.2. 4.3. 4.4. 4.5. 4.6. 4.7. 4.8. 4.9.

Steam absorption machine continuously hunts. Steam turbine chiller utilizes excessive amounts of steam. Rooftop unit causes excessive vibration noise. Hot water and condensate pumps produce noise and vibration. Boiler shows signs of freezing damage. Boiler cast iron sections are cracking. Unstable water level in boilers. Low pressure heating boiler shows signs of pitting corrosion. Centrifugal refrigeration compressor has operating problems.

IV

61 62 63 65 66 67 69 70 71

l

Chapter 5: Hydronics 5.1. 5.2. 5.3. 5.4. 5.5. 5.6. 5. 7.

Continuous check valve slam occurs. Severe pitting occurring to chillers. Corrosion occurring throughout entire hydronic system. Corrosion occurring to cooling tower piping. Steam traps blow by steam. Malfunctioning steam pressure reducing stations. Balancing valves do not facilitate system balancing.

73

74 75 77

79 83 83

Chapter 6: Interior Air Quality And Air Borne Contamination Control. 6.1. 6.2. 6.3. 6.4. 6.5. 6 .6. 6 .7. 6 .8.

6.9. 6 .10. 6 .11. 6.12. 6.13.

Insufficient outdoor air creates IAQ problem. System has insufficient total air supply. VAV box minimum settings create IAQ problems. Humidity levels climb above 60%. Perimeter zones have IAQ problems. Partitioned offices have air distribution problems. Interior open office ventilation creates IAQ problems. Half height partitions with closed bottoms create problems. Wet air handler final filters cause serious problems. Problems with dirty reheat coils. Exhaust discharge works its way back into the building . Nitrous Oxide buildup in Hospital Operating rooms Bag In/Bag Out filtering system.

85 86 86 86

87 87 87 88 89

91 91 93 93

Chapter 7: Laboratories 7 .1 . 7.2. 7.3. 7.4. 7.5. 7.6 . 7.7

Pressure dependent systems. Fume hoods can not get sufficient makeup air. Contaminated air escapes from Lab to surrounding areas. Negative room can not get makeup air via normal crackage. Conventional VAV fume hoods can create room imbalances. Fume hoods don't remove particulate matter in hood. Fumes hoods don't exhaust all fumes within the hood.

v

97 97 98 100 101 102 103

About The Author: Hal Finkelstein has more than 40 years experience in the HVAC field. He is one of the founders of the Empire Consulting Group, a twenty year old company specializing in consulting to Design Firms, Healthcare, Commercial and Industrial Facilities. Their specialties included contamination control and HVAC system problem mitigation. Mr. Finkelstein is the author of more than ten books and publications and many more papers on ventilation, HVAC systems and contamination control. He is an educator in the field of Heating, Ventilation and Air Conditioning and presents Seminars throughout the United States.

VI

Chapter 1 Air Distribution

1.1.

Maximum and minimum supply air quantities for each VAV terminal unit were never listed on the design drawings.

Wherever the max and min values are not listed on the construction drawings, it is expected that the contractor would know that the maximum value had to be equal to the total of the cfm values listed for the diffusers connected to the applicable VAV terminal box.

Though this may work for the

maximum value, many different procedures are utilized to determine what the applicable minimum value should be. One method is to set the minimum values at 30 to 33% of the maximum value.

Though this method may be close to the

required value, it may fall short in providing the amount of air required for good Interior Air Quality and/or good humidity control and low load.

For this reason, the maximum and

minimum terminal box values should always be selected by the design engineer and shown on the design drawings. several

situations

reported

to

the

NRC

by

In

SMACNA

contractors, where the minimum terminal box was not shown on the drawings, the contractors elected to set the terminal boxes in many areas at 30% of the maximum value. Only after

comfort problems developed did the contractors find out that the actual design was based on minimums in many of the rooms to be set at 50% of the maximum value. Had this value been shown on the drawings at the design stage, much fighting and money could have been saved rather than having to make the corrections after the fact. In one major high rise building reported to the NRC, close to 500 perimeter boxes were set up by the contractor at 30% of the maximum value since the minimum value was not shown on the drawings and the contractor was in the middle of commissioning the HVAC systems to meet a move-in dead line. After the units were set up, the engineers came forth to notify all involved that the terminal boxes should have been set with a zero minimum flow value. After the terminal boxes were then set up at the zero flow minimum, interior air problems were reported by the majority of occupants. Many dollars later, and in the middle of a litigation, the contractor was paid to go back and set the terminal boxes back to 30% of the maximum value.

To

accommodate the increase in minimum flow, the supply temperature in these areas had to be reset to a higher value which in turn created humidifies exceeding 60%. presented both additional comfort and health problems.

2

This

1.2.

Linear diffusers utilized in a VAV system create drifts and are causing major comfort problems.

Linear strip diffusers must be utilized with deflection blades and the defection blades must be adjusted during the balancing process. And we stress that they must be adjusted. Additionally,

it

is

advisable

to

require

multiple

duct

connections for every linear diffuser exceeding 4 feet in length. One connection for each four foot section appears to work well. On long lengths of linear diffusers, providing these multiple duct connections assures even distribution of air into the linear diffuser pfennums.

1.3.

SMACNA-recommended straight duct lengths for flexible duct between fittings when not adhered to, cause reduced air flow from terminal boxes and diffusers.

To many times when flexible duct is utilized there is a complete disregard for guidelines and standards that specify that the flexible duct sag should not exceed 112", and that the maximum length should not be over 10 feet. To many times flexible duct is utilized strictly to snake through and around obstacles in the ceiling space. In doing this usually excessive lengths of flex duct are normally utilized with excessive sag creating large resistance to air flow.

This in turn can severely

effect the out put from the terminal boxes. 3

1.4.

Duct layouts to avoid large pressure differentials between branches were not utilized, thereby generating an out of balance system with large differences in branch pressure drops and air flows, especially in VAV systems under partial load modes.

Ducts should be designed and installed to create a total system with as near as possible symmetrical branches or at least with equal pressure drops.

This type of system is easy

to balance and keep balanced as loads within the system vary along with associated air flows.

Duct systems with branches

that have a very wide difference in pressure differentials will not stay balanced and will create large variations in cfm flows from branch to branch.

1.5

Interconnecting (or loop) arrangements when helpful to reduce large pressure differences between branches, were ignored, creating the same type of situation as discussed in# 1.4.

Where it is not practical to design and install a duct system with branches containing equal or near equal pressure drops the ends of the branches are normally looped so as to equalize the pressure differentials, if this is not done you end up with the same condition as described in # 1.4 above.

4

1.6.

Systems utilizing large ceiling return plenums demonstrate uneven air return characteristics with resulting Interior Air Quality problems and comfort problems.

When using ceiling space as a return plenum, return air collecting inlets above ceilings must be evenly distributed and not more then 4,000 CFM should be collected at any one location.

1.7.

Diffusers dumping air at low flow create drafts and other comfort problems.

Diffusers having high entrainment characteristics at low air flow should be used. Many low entrainment type diffusers have a tendency to dump air directly downwards when VAV systems operate at low

air flow.

Figure 1. 1

shows what

happens to each of the two most popular types of diffusers utilized in HVAC systems as the air flow is reduced from maximum flow.

When a diffuser is sized based on the

maximum air flow rate it will have a sharp drop in discharge velocity as the air flow is reduced as it is in a VAV system.

The reduction in discharge velocity also causes a decrease in the throw of the diffuser. In this discussion we are taking the throw at the distance in which the terminal velocity will be 50 fpm. 5

Discharge Velocity vs Capacity 1,400 1,200

-

1,000

c

"(3

_Q Q)

>

800

Q)

ei

ro

..c (_)

600

(J)

Ci

400 200

0

10

20

40

30

50

60

70

80

90

100

Percent Of Full Capacity

Fig, 1.1

This distance is very important to proper comfort levels. Figure 1.2 shows how the throw is reduced as the air flow is reduced for the same types of diffusers. The reason the throw drops off rapidly is because the throw is dependent on the velocity of the air stream and the overall momentum which is dependent on the mass of the air at any given time. The mass is decreased because the VAV terminal is decreasing the flow.

6

As the throw drops off it will reach a point where in effect the air is falling down through the occupants comfort zone,

the

effect known as "dumping".

Throw vs Capacity

10

20

30

40

50

60

70

Percent Of Full Capacity End Of Throw Indicates A Velocity Of 50 fpm.

Fig. 1.2

7

80

90

100

1.8.

System utilizing light trotters experiences areas with drafts and comfort problems.

Supply Light troffers must be installed

with horizontal air

deflecting blades. It is import when light trotters are installed in ceilings below 12 feet that they be provided with horizontal deflecting blades and that these blades be adjusted as part of the balancing procedure to assure that they will not create drafts. The air discharge from the light troffer slots is in the form of slip streams and rely on the deflecting blades to create

a directional air pattern that will prevent high terminal velocities in the comfort zone, which in turn creates drafts.

1.9.

Utilizing a pressure dependent VAV system in a large facility, causes large variations in the air flow provided to spaces.

Where duct pressure variations are greater than 1" H10. and especially where the supply system will exceed 10,000 CFM peak capacity, a pressure-independent

system must be

utilized to assure proper control and operation.

1.10. In an attempt to reduce installation costs the contractor makes the diameter of the inlet duct to the terminal box smaller than the VAV terminal box inlet,.

This then creates turbulence at the terminal's

entrance.

8

This type of turbulence reduces the capacity of the terminal box, and creates control problems for the velocity sensor. Make the inlet duct the same size as the terminal box's inlet.

1.11. Neglecting to install a straight run of 1% times the inlet duct diameter as rigid duct preceding the terminals inlet, creates uneven entrance flow conditions.

#1.10 and #1.11 are connected in that they affect the inlet conditions to the terminal box. The terminal box operates a great deal like an orifice. In this sense, the inlet condition to the box greatly affects the output from the box at the selected pressure differentials.

If you expect the box to operate as

indicated in the manufacturers literature and in accordance with the performance curve then you must pay attention to these types of requirements. Utilizing 1 112 duct diameters of straight duct at the terminal box entrance assures that the immediate entrance is perfectly straight, reducing the chance of creating inlet turbulence caused by any fitting.

1.12. Installing flexible duct runs exceeding 5 feet, or worse, greater than 10 feet.

When flex duct is utilized from terminal boxes to supply diffusers, or from main branches to the inlet of terminal boxes, the degree of sag which occurs in the run, is critical to the 9

proper operation of the terminal box. The greater the amount of sag, the greater the pressure drop in the flexible duct connection.

Follow SMACNA installation requirements for

flexible ducts.

1.13. Neglecting to Install flexible duct, straight and properly supported (see SMACNA design and installation standards).

SMACNA has some excellent installation and design manuals and we highly recommend them.

One of SMA CNA 's basic

installation recommendations, which reduces the chance of operating problems is to use flex duct in lengths not longer than ten feet, with a maximum sag of 112" in any 10 foot run. Sag amounts greater than this will cause a significant increase in branch pressure drop, with an equal significant decrease in cfm flow, creating a serious air flow problem.

For example. A 10 foot length with a 4" sag can have a friction loss of .08" H20. If the sag is increased to 8" as it might very well be for snaking around obstacles, the pressure drop increases to .43"H20, with a resultant significant decrease in air flow from the related branch. (see fig. 1.3 )

10

Flex Duct Pressure Loss vs. Velocity 10

r- -

------r·--- -- ·

t

8

2

i ｾ＠

ii'.

§

ｾ＠

Qj

6

r+- --

-

- --- -

---j -

---- ---

t- -- - -i - -- -- Ｍ Ｍ ｾ＠ --Ｍ ｾ＠ j

t

!

t

IPD Due To Sag

I

41-- - ---- - -1 - . + -

-

- -

-

-

--

-1

ｾ＠

--- I

I

--

I

I

I

I

___ j ___ --- - -

-

- - -- -

I

I

--+ -

-- -- ｾ＠

I -

-

I

- - ---;--- --· -

I

2

-,.

I

-

!/

+

-

- .

i

-i

I 0

1200

600

Velocity - FPM

Fig.1.3

II

1750

2800

1.14. If the minimum air flow rate is below 50% of maximum, neglecting to check the ADPI (Air Diffusion Performance Index) for cooling at both the maximum and minimum points will cause significant comfort problems.

1.15. In a VAV system, as the flow is decreased to below 50%, the air does not diffuse into the room. Instead, it simply drops to the floor, creating drafts and many comfort problems.

#1.14 and #1.15 are related problems and stem from neglecting a diffuser's ADP/. The ADP/ for cooling should not be below 80% at 100 fpm terminal velocity for ceiling slot diffusers and at a 50 fpm terminal velocity for other diffusers (see Chapter 31, 1989 Fundamentals) to prevent comfort problems due to air diffusion characteristics.

The ADP/ is known as the Air

Diffusion Performance Index.

Studies show that ADP/ is a

function of the type of diffuser, the load, the cfm flow rate and the room's aspect ratio. The higher the ADP/, the greater the diffuser's diffusion rate.

Note from figure 1.4, that with the two most popular types of diffusers, as you go to 75% of maximum capacity, the ADP/ falls off very rapidly. Of course once the ADP/ falls below 80% you can expect comfort problems to develop. At ADP/ levels far below 80%, as you may experience when a VAV terminal starts to approach 50% of maximum flow and lower, with most types of diffusers you are guaranteed of generating a situation

12

where you will get very poor diffusion and mixing, which will result in drafts and discomfort.

ADPI vs Capacity -------

,--

100

-

---· -

--- -

80

60 - + - - - +

a:0


tt:. > tt:.

4

Relief Damper

l

Branch 2

C0

Rel.

High Limit

_ I

Damper

ｾ

ｾ＠

Controller

'

Return Fan

Control Of Supply And Return Fans From A Single Controller

Fig. 3.2

3.10. Locate the building pressure sensor to represent average building pressure.

Do not locate sensor close to entrance doors or other locations where the pressure fluctuates. Atrium above first floor is generally a good location.

3.11. VAV terminal multi-point sensors must have a velocity of not less than 600 fpm for proper operation of the terminal controller at minimum position.

Even though they may be factory installed, most single or multipoint pressure sensors utilized in terminal boxes are of the differential pressure type.

Where they are, the minimum

air velocity at the sensor(s) must be 600 fpm to allow for

52

accurate controllability. A multipoint sensor is more accurate than a single point sensor but utilizing a multipoint sensor does not alleviate the 600 fpm minimum velocity requirement.

3.12. VAV terminal box with DDC controllers malfunctions wh enever the system operates on minimum air flow.

Setting of the DOC controller can be an involved and time consuming process.

Most controllers, which have to be

adjusted in the field at the terminal box, are adjusted utilizing a portable operator's terminal or laptop computer while verifying the

flow

with

traverses

and

verifying

pressures

with

magnehelic gauges. Then the controller is adjusted to make up for the difference existing between the readings indicated on the magnehelic gauges and the DOC transmitter. And, then the unit is put through all cycles to verify that the settings did not change while other adjustments were made.

Operating most controllers below .03 LJP, "H20 is risky. And, in many cases, creates operating problems. Many designs call for a minimum primary cfm air flow value which develops a pressure differential at values far below .03 LJP, " H20.

Many

controllers set between .01 and .03 "H20 have a tendency to chase the set point, causing serious variations in flow. Other controllers set at those values may not control at all.

To

enable the controllers to control the terminals they must be set at a higher minimum cfm value or have major sheet metal 53

modifications made to the terminal inlets.

Or,

a correction

factor will have to be developed and input into the building computer to allow the boxes to be adjusted near operable values.

The problem becomes even worse when terminal boxes selected are over-sized in order to assure a very low noise level.

Many times, the only way to increase the operating differential pressure is to increase the velocity pressure which requires an increase in the air velocity at minimum cfm flows.

In an

attempt to increase the terminal inlet velocity at the terminal's sensor, some designers try to utilize a circular device known as a doughnut.

This device is installed at the inlet to the

terminal box. However, in order to assure that the device does not generate turbulence at the box inlet, it must be placed at least 2 112 duct diameters upstream from the inlet. doughnut must also be fabricated

The

with a smooth round

beveled hole and the device must be stiff enough so as to cause no bending or vibration under maximum cfm flows.

The following table in figure 3.3 shows how problems can be under minimum air flows:

54

extensive the

Terminal Size 6" 8" 12" 14"

Design Max CFM 80 105 420 550

Design Min CFM 25 35 126 165

Min CFM L1P .003" .002" .005" .004"

Min CFM Velocity 220 fpm 180 fpm 280 fpm 259 fpm

Fig 3.3

3.13. Pneumatic controllers do not operate properly or their operation is chaotic and unpredictable.

Pneumatic controllers require a minimum of 18 psig to operate properly.

Some systems may have long tubing runs with

undersized tubing. If too many controllers operate off of the same line, and that line is undersized, then at times of high load the pressure at various controllers could fall below the required 18 psig. To make matters worse, it may not be the same controllers all of the time. Depending upon how the load is distributed, it could be different controllers at different times which malfunction. If this appears to be happening, check the pressure of the air supplied to the controllers at different load variations, including full branch load. It is not uncommon to figure a maximum pressure drop of 2 psig for the air line with the longest effective length. In this case if you require 18 psig at the control furthest from the compressor, you would have to provide air at a pressure of at least 20 psig. Additionally, allow approximately 5 cfm per 100 air consuming instruments as a ball park guide.

55

3.14. Fan adjusting control in a VAV system is constantly hunting.

Check to see if the static pressure controller is oversized. The static pressure controller should be specified by size and sensitivity.

If the installed controller is oversized it could

create large fluctuations in the signal which it sends back to the fan. If this is the case, check the sizing and sensitivity of the installed static pressure probe against the manufacturer's recommendations.

3.15. Terminal boxes continuously hunt, attempting to control the required cfm. Static pressures are above the minimum required.

First check static pressure readings and entrance and exit duct conditions. If these are proper, the hunting problem may be a result of a static pressure controller span discrepancy. It is important to note that the span of the static pressure transducer must be small enough to reduce the total amount of random control error. The total random control error for a transducer with a range of 0 to 1" is shown in figure 3.4. Note that this error is random and how and when it occurs is unpredictable. If you have to control at about the .02" range, note the degree of error at such a range. Now compare that to the total random error that would occur if we narrowed the control span to 0-. 1 ". The difference in the total random error that would occur is very significant. There would be virtually 56

no control problem from the transducers due to the random error with the tight control range of 0-.1", as shown in figure 3.5 when you are controlling small pressure differences.

% Control Error vs. Pressure Sensor Reading For Pressure Sensor Range O - 1

11

100% -.------

-

--------------------.,

I

I

"0

50%

"w

I

I

I

I

I

I

I

I

I

I

!:;

§

I

1- - 1- -1- -1- -1- --1- -i

"-

0

I

0% -;.-----------------------------------

! _50%L-1-Ｍ Ｑ ＭＺ Ｍ Ｚ Ｍ

ｾ Ｍ ｾ＠ -100J '":"'.I ": "': "':" ':" _.,. - ,-/ 0.005 0.015 0 .025 0.035 0.045 0.055 0.065 0.075 Differential Pressure - H2 0 "

Fig. 3.4

57

Pv

% Control Error vs. VAV Box Sensor Pv For Pressure Transducer Range 0 - .1" 100%

r

! -- :

ｾ＠

I

T---r

i

:- 1

I

I

4- --ｾ＠ - r- ---1- t- - 1- -+ ---t- -t .

50%

% Total Control Error

-50%

-1 00%

I

ｾｬＮ＠

l

0.005

I

I

I

I

I

I I I I I I I --i - _j_ - -l -- -1- _, _ _J_ - !-- -! I I I I I I I ' ' ' I' ' ' '.i rＬＭｾｬｔ Ｍ ' ' I' ' ' ' I' ' ' i ' ｾｲｌｬ＠ 0.015

0.025

0.035

0.045

0.055

Differential Pressure -

0.065

0.075

0.085

HP"

Transducer With Aulozeroing

Fig. 3.5

3.16. Control configurations can adversely affect the operation of VAV terminal units and fan operations.

The three basic control configurations that have been utilized in HVAC systems are known as proportional (P), proportional integral (Pl), and proportional-integral-derivative (PIO).

With

proportional controls, an inherent offset is present when the control is in action.

Recycling is often experienced after a

variable change. A proportional-integral (Pl) controller resets

58

the

control responses

and

eliminates

this

offset.

A

proportional-integral-derivative (PID) controller senses the rate of change of the error and brings the parameter to the set point without much recycling after a variable change. For example.

In sensing an HVAC system's static pressure and

regulating the fan from same, this offset injects an error into the level of the duct static pressure. And in the case of a VAV system, the error tends to cause a higher pressure than the set point causing a substantial energy waste at off-load conditions. Utilizing Pl and preferably PID for the control of all supply fans in VAV systems will prevent the over shoot in static pressure at reduced loads.

3.17. Utilizing the velocity pressure signal from a duct directly to regulate air flow creates air flow control problems.

Conversion

of velocity pressure into air flow indications is

very important for proper air flow control.

This is especially

critical when the return fan and supply fan must track in order to maintain a proper building pressure, or outdoor air control. This

conversion

can be readily performed by software

programs in direct digital control systems. For pneumatically controlled systems, a square root extracting device is utilized. This device converts the velocity pressure signal into a velocity signal, and in turn the signal is changed into airflow rates.

Where a specific process of conversion is not

specified, some control contractors will usually attempt to 59

utilize a less expensive simplified design short cut to approximate the conversion by utilizing the velocity pressure signal directly on the control transmitter. This requires very careful selection and a lot of fine tuning of the control system after installation.. .. which is rarely performed. Additionally, it takes a lot of fine tuning and maintenance to assure that this type of simplistic installation works over the long run.

60

Chapter4 Equipment

4.1.

A steam absorption machine is continuously hunting in an attempt to produce the design leaving chilled water temperature.

The temperature and pressure conditions within the absorber are influenced by the temperature of the system's cooling tower water. Therefore, the cooling tower water temperature must be closely controlled to maintain stable system operation in a single stage older type absorption machine.

If the

temperature of the cooling tower water should drop, the absorber temperature will drop accordingly.

At this lower

temperature the ability of the solution to absorb refrigerant is increased.

This, in turn, drops the evaporator pressure and

temperature, reducing the temperature of the leaving chilled water.

Sensing this, the chilled water temperature controller throttles the steam valve, reducing the rate at which refrigerant vapor is liberated in the concentrator.

This causes a more dilute

solution to be returned to the absorber, raising the absorber and therefore the evaporator pressure and temperature, in an effort to

re-establish

the

design

leaving

chilled water

temperature. In other words, to compensate for a fluctuating 61

cooling tower water temperature, the absorption machine must hunt continuously in an attempt to find a state of equilibrium between solution concentration and the changing cooling tower water temperature.

4.2.

Steam turbine chiller is utilizing excessive amounts of steam.

In general, steam driven centrifugal chillers are extremely inefficient.

They become even more inefficient than most

other types of chillers as load requirement is reduced. Check the chillers load. Has it changed recently? This type of chiller has a maximum COP value of approximately .80. When the steam turbine chiller is fed from a central distribution system, as in a campus type layout, there is no way to efficiently reduce steam flow to the turbine, except by throttling the steam flow via a throttling valve.

The throttling valve is an

inefficient way to reduce steam flow requirements since the throttling process itself is an energy loss process. If you now find that the steam flow has to be continuously throttled, it would be best to utilize a let down turbine in place of the throttling valve as long as you have use for the mechanical energy generated by the let down turbine.

For the steam

turbine chiller to operate more efficiently it must always be run at close to maximum load in order to utilize the least amount of steam.

62

4 .3.

No matter how many vibration eliminators are utilized on a rooftop HVAC unit installation, the unit causes considerable vibration throughout the building.

Fiqure 4.1 shows an improper way to install a roof top unit on a building.

It is not uncommon to find that such an

installation, especially when there is more then one unit involved,

causes considerable transmission of operating

vibration

throughout

the

building.

It

is

found

that

concentrating the HVAC unit's weight between the buildings beams will cause considerable roof slab deflection and the transmission of noise and vibration, no matter what type of vibration eliminators the installation utilizes.

Where this is

happening, the best way to eliminate the vibration noise problem is to install the unit as shown in figure 4.2 .

This

figure shows the unit installed on a steel frame supported by column extensions. This type of installation keeps practically all of the vibration out of the roof slab.

A comprise method

would be figure 4.3. This method adds an extra stiffening slab under the unit.

This slab adds extra mass and stiffness and

reduces the chance that significant amounts of vibration will be transmitted through the building.

63

HVAC Unit Roof Slab

Fig. 4.1

HVAC Unit

Roof Slab

Fig. 4.2

64

HVAC Unit

Roof Slab

Fig. 4.3

4.4.

Hot water heating pumps or condenser water pumps produce noise and cause acute vibration.

Pumps which are called upon to circulate hot water are affected to a greater degree by improper installations than pumps which circulate chilled water.

Cavitation can quickly

occur in hot water circulating pumps, generating noise, vibration and severely shortening the life of bearings and impellers.

A great deal of the installation problems occur in

the suction piping directly proceeding the pump. Additionally, it is important to remember that the greater the velocity of the water within the suction piping as it enters the inlet of the pump, the greater can be the problems caused by improper

65

installations. When suction piping is the same size as the size of the suction inlet,

there should be no elbows or valves

placed any closer to the suction inlet, then five times the pipe diameter.

If the suction piping is larger then the suction

opening to the pump and the pipe utilizes an offset reduction piece to connect the piping to the suction inlet, then there should be no elbows or valves placed any closer then five times the pipe diameter from the reduction piece to the first elbow or valve.

Installing elbows or valves any closer than

these distances will cause an increase in turbulence with resulting vibration, noise and reduction in capacity and the life of the bearings and impellers.

4.5.

A Boiler ta ken off line and isolated shows signs of freezing damage to its interi or.

Facility boiler plants that use a standby boiler as a rotating online boiler, to be able to take different boilers off line for maintenance and repairs, besides having their water lines and steam lines isolated from the other boilers, must also have its stack isolated. If this is not done in a winter climate, cold air, reversing itself down the boilers individual stack, will cause freezing damage to the boiler's heating surfaces. The design should include a method for closing a stack isolation damper remotely, through the computer panel for each boiler that is

66

taken off line for maintenance and repair work. This of course is of main concern during winter months in northern climates.

4.6.

Cast iron sections are cracking at the section roots above the combustion chamber in oil fired cast iron modular boilers.

Cast iron modular type boilers normally utilize a flame retention type oil burner or other type of high efficiency oil burner.

The amount of excess air provided to the burner

creates a high degree of fuel wastage and an elevated and elongated flame. The flame wil/ have a tendency to hit the cast iron sections, usually in the vicinity of the back part of the sections and at the upper part of the section towards the casting root. The sections will crack at that point due to the high temperature of the flame impingement on the cast iron surface. This is known as flame impingement failure. You will see the discoloration in the vicinity of the cracks that will be related to the specifics of flame impingement.

Flame

impingement is usually caused by providing too much excess air.

If the boiler uses a burner with turn down, it can vary its load in proportion to the actual load on the building. It is equally important that when you use turn down, that the combustion system be capable of varying the amount of excess air as it varies the firing rate in proportion to the actual system load. A typical excess air variation is shown in figure 4.4. 67

Exc ess Air vs B ur ner L oad(%) -

10

20

40

30

60

50

80

70

90

1 00

Burner L oad(%)

Fig. 4-4

Fuel Wasted ( 0/o) vs Excess Air (o/o) Number #2 Fuel Oil 50

....

ｾ＠

.... ->-

....

40 ,-..

cf2.

.......... "O Q)

-

30

....... Cf) cu

s

-ｾ＠

ｾ＠

__......

ｾ＠

ｾ＠ ｾ＠

ｾ＠

Ｍｾ＠

ｾ＠

20

Q)

::::l

LL

10

0

10

20

30

40

50

60

Excess Air(%) Fig. 4.5

68

70

80

90

100

Fuel Wasted (o/o) vs Excess Air (o/o) Natural Gas 60

50

-ｾＴＰ＠

-s

----

'"O

,_ ,_

(l)

ｾ＠

30

ｾ＠

20

ｾ＠

ｾ＠

ｾ

ｉ ｾ＠

/

ｩ＠

ｾ＠

LL ｾ＠

ｾ＠

10

-

-

0 10

20

30

50

40

60

70

80

100

90

Excess Air(%)

Fig. 4.6

4 .7.

Unstable water level in boiler causes water level controls to malfunction and also causes water hammer throughout the stream piping system.

One of the most common causes of an unstable water level in a steam boiler is a buildup of impurities in the water.

These

impurities can be in the form of oil, solids and excess treatment chemicals.

A high impurity level causes constant

69

churning of the water level and in its extreme, actual foaming within the boiler water. The churning causes the water level to rise and fall by large amounts.

In its worst situation, the

churning on the down side could cause the water makeup controls to feed excessive water into the boiler.

This then

causes an excessively high water level, which in turn can cause water to carry over into the steam system and generate damaging water hammer.

4.8.

Low Pressure heating boiler shows signs of a pitting type of corrosion.

This type of corrosion in low pressure boilers is generally caused by large amounts of dissolved gases in the boiler water.

The gases which play the biggest part in pitting type

corrosion are Carbon Dioxide and Oxygen, with Oxygen being the largest offender. Carbon Dioxide acts as a type of accelerator, speeding up the corrosion that may be occurring. Therefore, if we can get rid of the excess amounts of oxygen, we can reduce or eliminate the pitting a/together.

You can

control the amount of dissolved gases by adding alkali to the boiler water thereby increasing the alkalinity of the water.

To

neutralize all of the Carbon Dioxide in the boiler water you want to produce about 250 ppm hydroxyl alkalinity as CaC03 . You can accomplish this by adding a sufficient

amount of

alkali chemical to maintain a boiler water alkalinity of 400 ppm expressed as CaC03_ Under these conditions, you should have a phenolphthalein alkalinity of approximately 70% of the 70

total alkalinity.

You can protect against the Oxygen in

systems that have large amounts of absorbed Oxygen and Carbon

Dioxide through

the use of Sodium

Chromate

chemicals or other Oxygen reducing chemicals approved for usage by the EPA. Where Sodium Chromate is utilized as a corrosion inhibitor, maintain approximately 800 ppm as sodium chromate in the boiler water at all times with a pH at 8 or slightly higher. For all chemicals follow the manufacturer's instructions in detail.

4.9.

A centrifugal refrigeration compressor without variable speed but utilizing a simple suction valve to throttle gas flow for capacity control can create efficiency problems and utilize excessive electricity or steam to create its refrigeration.

Where variable speed control is not utilized for capacity control, a form of suction gas control would have to be utilized to initiate capacity control.

Two methods utilized are inlet

guide vane control and suction valve control.

Adjusting the

inlet vanes controls the compressors capacity by reducing suction pressure, and in turn, flow and also reduces the discharge pressure.

The inlet vane control method is very

efficient because it imparts a pre-rotation to the gas. The simple suction valve does not impart pre-rotation and is therefore much less efficient.

Although the suction valve

method may be less expensive than inlet vane control, it is less efficient when it comes to operation. 71

You, therefore, end

up utilizing additional steam or electricity to operate the centrifugal refrigeration compressor. Depending upon the size of the machine, the increase in operating cost will far exceed the installation savings obtained by utilizing the simple suction valve control method.

72

Chapter 5

Hydronics

5.1. Continuous check valve slam occurs in a hydronic system.

Check valve slam is an occurrence that is due to the noise which is made when the moving portion of the check valve still has enough momentum to make noise and cause movement of the valve body when it contacts its seat. When a pump is shut down, the pump may still have some rotating movement. This movement causes the pump to still generate head and deliver flow for a short amount of time.

But when the head finally

drops below the system head, flow will reverse, causing the check valve to close and slam.

For example. When the flow

reverses before the check valve is completely closed and the return flow catches the check valve, it moves the element off the seat and slams it closed.

Where check valve slam is a

problem, the valve which should be used to prevent this occurrence is a center guided spring loaded check valve.

In

this type of check valve the disk is only required to move a very short distance from open to shut. In this way, the valve will close before the reverse flow develops, and slam will be eliminated.

73

5.2.

Severe pitting of newly installed chillers in the absorber and condenser tubes occurred in a system utilizing 4" to 6" steel piping with an open cooling tower.

It is important to note that badly maintained water circuits in a chiller system can cause very rapid corrosion of chiller absorber and condenser tubing. In fact, it is not uncommon that

even

in

new

installations,

such

corrosion

completely destroy a chiller in one to three years.

could

One very

common scenario is as follows:

Algae, spores, sand and other debris builds up to cause fouling of the absorber and condenser tubes. As the fouling progresses, localized, differential corrosion cells begin to form.

The differential corrosion cells cause pitting corrosion

and attack the internal areas of the absorber and condenser tubes.

As the pitting corrosion progresses, it causes wastage of copper material.

Copper then plates out on the interior

surfaces of the steel piping material. A combination of low pH values (6-7) along with the warm condenser water and the copper plating out on the steel causes corrosion to then progress to the steel piping. This corrosion process develops tubercle material which, due to the water velocity, periodically breaks off and ends up in the cooling tower water circuit.

74

The iron oxide that then enters the cooling tower water circuit begins to work its way to the absorber and condenser tubes, accumulate with the algae, spores, sand, etc. and begins to cause increased fouling of the tubes.

Once fouling rapidly

accelerates, cleaning brushes utilized in the chillers usually becomes ineffective. Water flows and water velocities begin to severely decrease and tends to decrease the effectiveness of cleaning brushes which may be installed.

Therefore, don't

depend upon them in this type of situation.

This type of corrosion cycle becomes self feeding. The more corrosion, the more you get increased fouling which then creates increased corrosion and so on and so on. To stop the process, you must start with a complete system cleaning, keeping the cooling water clean and making sure that the water pH stays in the 8 to 8.5 range or at the pH level required, depending on the water treatment chemicals utilized in the system.

5.3.

Corrosion occurred th roughout an entire hydronic heating system. The system utilizes copper tubing, and operates as a micro-bore system utilizing supply water temperatures of 220 to 2300 F. This system also utilizes a 600 delta temperature drop through heat transfer equipment. The system is experiencing a great deal of corrosion and gas generation.

In systems as above, water cleanliness and the pH of the water are critical in preventing and if required, correcting the above 75

situation. Operating a hydronic system at temperatures above 1800 F

calls for a great deal of attention to housekeeping

applicable to the water within the system.

First, it is

imperative to keep the water perfectly clean, mill scale, mud, sand and other such items must not be allowed to enter the system.

At the time of system startup the interior of the

system must be cleaned with any of the commercially available cleaning products before the water is allowed to enter the system. Periodically, during operation, the pH of the water must be checked and maintained at no less than 8. If the pH is allowed to go lower, corrosion will occur and accelerate. Air must also be properly vented from the system. It is not uncommon that methane gas forms in the system due to the debris, air and low pH of the water. Proper air venting of the system must be through commercially available venting equipment

but not through the use of many automatic air

vents. Many automatic air vents can suck more air in to the system then they vent out. The best way to vent is to utilize commercially available centrifugal air eliminators in

the

mechanical room and manual air vents at the high points in the system.

76

5.4.

Continuous undetected corrosion occurs within the cooling tower piping systems.

The

cooling tower water circuit and any other open piping

system, besides having a good way to monitor the water treatment, must have corrosion coupon stations in order to properly monitor the extent of corrosion and the rate of corrosion.

In this way, you can gather the information

required to determine how efficient the water treatment actually is.

The corrosion coupon's main purpose is to measure corrosion in the form of pitting so as to allow the operator to apply factual numbers to the determination of water treatment efficiency.

The coupon is actually a strip of metal that is

placed in a type of holder and placed within the flowing fluid. In order to obtain a fairly constant result the coupons must be prepared in a standard way. The American Society for Testing Materials has a standard which gives details for cutting and polishing the coupon strips of many different types of metals. You can also purchase the coupons from a water treatment supply firm.

After the strip is prepared it is placed in the center of the pipe in which the water is flowing.

At the beginning, the corrosion

on the strips will have a high corrosion rate. . After a short

77

amount of time the rate will begin to fall off steadily, eventually stabilizing at some constant rate.

The strip coupon is weighed both before and after the exposure. The pits are measured with a dial pitch depth gauge and their number, size, shape and distribution are then recorded.

The average corrosion rate in mils per year is calculated from the loss in weight of the coupon. Following is the ASTM formula: Average penetration, mpy

where:

=(w) (365) 103 I (2.54)3(A) {d) (t)

w = weight loss in grams A

=surface area of coupon in square inches

d = density of metal in grams/square centimeter t =time of exposure in days mpy = mils per year

(By comparing the mpy over time you can track the efficiency of the water treatment process)

78

5.5.

Steam traps blow by steam even though they appear to be installed and operating properly.

Besides faulty installation one of the prime reasons for steam traps blowing by steam is improper sizing.

The following is a shortcut guide to sizing steam traps: You must know the following:

1. The net final pressure at the steam trap (after all valve pressure drops, etc.),

2. If there is a rise or lift after the trap, this must be accounted for. Figure every two feet of rise equals about 1 psi of backpressure against the trap.

3. All other possible causes of back pressure (e.g. discharging into a pressurized line, tank, system, etc.)

4. Pounds per hour of condensate to be discharged. Consider warm up condensate, etc.

5. Safety factors are usually utilized in selecting the final size of a steam trap.

The differential pressure which the trap will act against must be determined. This will usually be the steam pressure at the

79

trap minus all back pressure values and differential caused by lift.

Example:

If we calculated that the condensate load for the trap is 50 pounds per hour, with a steam supply pressure of 150 psig with a lift of 30 feet, the calculations for steam trap size will be as follows:

Supply Pressure = 150 psig Lift= (30 ft.) =

15 psig

Differential pressure Load

=135 psig

=50 lb./hr

Safety factor = 2 Trap Size= 100 lbs/hr with a differential pressure of 135 psig. In the previous example we assumed the initial trap load. In selecting the actual trap load, depending upon the type of equipment on which the trap is installed and the type of trap, one of the factors you must consider is any warm up that will add to the trap load.

When calculating the warm up load there are two basic equations utilized: one pertains to insulated pipe and the other pertains to uninsulated pipe sections.

80

The basic loss that creates the warm up load has to do with radiation losses. For insulated pipe sections the formula is as follows:

#/hr = Ax U x (ts - ta) x E. HL #/hr= Trap Steam Load A = External Pipe Area U = Insulation U Value. ts = Steam Temperature ta = Air Temperature E = 1 - Insulation Eff.

For sections of pipe without insulation the following formula applies:

#/hr = W x (t1 - t 2) x 5 h HL #/hr W T1 T2 Sh HL

= Load on Steam Trap =Weight of Total Pipe Length =Final Pipe Temperature =Initial Pipe Temperature = Heat Of Pipe Material = Steam Latent Heat

Some common values utilized for pipe weight, specific heat and other parameters are as follows.

81

Pipe Weights Per Foot In Pounds

Pipe Size 1 1/4" 2" 4" 6"

Sch.40 2.27 3.65 10.79 18.97

Sch.80 3.00 5.02 14.98 28.57

The specific heat of steel pipe can be taken as . 12. Insulation efficiency is usually taken as .75 to .85, depending upon the type of insulation and how the insulation is installed.

There are tables available from Armstrong Steam Trap Company (616) 273-1415

and other steam trap companies

which solve the equations previously indicated on Page 10. The equations do have to be utilized for conditions not covered by manufacturer's catalog information.

Safety Factors Use Control

General

Mains Storage Heaters Space Heater Exchangers Air Heater Submerged Coils (low level drain)

Submerged Coils (siphon drain) Tracing Lines Presses

2 2 2 2 2 2 2 2

*Safety Factor Information From Sarco Steam Hookup Book

82

With Temp

3 4

5.6.

Malfunctioning steam pressure reducing stations, result in severe valve chattering, water hammer and other such mechanical failures.

It must be remembered that a steam pressure reduction valve is designed to operate on steam, not a combination of water and steam or water, steam and dirt.

If it is suspected that you

are dealing with wet steam you must utilize a steam separator to separate out the water and dirt, and send only the clean steam to the pressure reduction valve. Additionally, check the size of the pressure reduction valve. As a hint, the valve size will usually be less than the size of the pipe.

If in fact the

valve size is the same as the pipe size, you can suspect that the pressure reduction valve is sized too large.

5. 7.

An operator may find that balancing valves have very little effect in aiding the balance of hot water and chilled water systems.

All too often balancing valves are selected to match the size of the pipe in which they are being installed.

This is not the

proper way to select these type of valves. A valve sized in this way will have very little effect on flow adjustment until you get to a position where the valve is almost tightly closed. Using the valve this way will also cause damage to the valve seat and reduce the useful life of the valve. In addition, a problem such as wire drawing occurs.

Wire drawing causes a channel to

form in the valve's seat. This in turn, causes the valve to allow increases in water flow over time. 83

The balancing valve must

be selected by calculating the valve's flow coefficient. Where the pressure drop of the valve and the system will be determined in feet of the water the flow, coefficient for selecting such balancing valves comes to :

Cv = 1.5Q(spAh)-5

C v = Flow coefficient at 1 psi drop Q

= Design gpm flow

Ah = Pressure drop, ft of water Sf

= Specific gravity of fluid.

Selecting a balancing or control valve through the "coefficient of flow" method allows you to select a valve that will give smooth proportional flow adjustment.

Otherwise what you

wind up with, is a valve that does not control until it gets slightly off of the seat, causing wire drawing and valve failure along with improper flow control.

84

Chapter 6

Interior Air Quality And Airborne Contamination Control

6.1 .

System not designed for best outdoor air criteria thereby creating IAQ and infection control problems.

Outside air is critical for proper interior air quality and good infection control considerations.

Do not provide less then

approximately 113 of maximum air supply as the minimum turndown value and definitely not less than the required amount of outside air to the occupants at the breathing zone, per code and industry standards. Engineers and contractors are being held responsible for Interior Air Quality and Infection Control problems despite complying with local codes.

The

deciding factor is...... did you provide that which would be considered "Good Engineering Practice" utilizing the "Best Available Technology". Beware of designs which provide zero percent minimum air values in VAV systems.

These types of

designs have played a significant role in sick building syndrome litigations.

85

6.2.

System not designed for room peak total supply air flow or system designed for less than .9 CFM per square foot of floor area.

Tests have shown that you should never circulate a total air supply of less than .9 CFM per square foot in occupied buildings. This is the minimum total air that should be circulated in order to allow for proper air mixing and good interior air quality.

6.3.

Facility experiences a large amount of Interior Air Quality complaints when ever the VAV boxes are operating at minimum settings.

Systems designed for less then a minimum supply air of 30% of peak supply air. (see # 6.1 above). Experience shows that VAV box minimums should not be set for less then 30% of peak flows. A check must always be made to assure that the minimum expected loads can be handled without reducing the minimum CFM below 30%.

6.4.

System not designed for room loads at minimum air supply, causing humidity levels to climb above 60%.

The humidity should never be allowed to exceed 60% in occupied buildings.

Humidity levels above

86

this

value

accelerate the growth of bacteria, fungi, molds and many other micro-organisms that affect health.

6.5.

Perimeter zones designed for over 1,000 ft2 floor area create control and comfort problems.

a. Suggested maximum floor area per zone to promote adequate VA V system operation: 1000-1500 sq. ft.

b. Maximum CFM per zone: 1200-2000 CFM per zone is the range that should be considered. 1500 CFM has been found to be a good manageable value for a perimeter zone.

6.6.

Offices partitioned should be limited to control areas of 1000 to 1500 square feet in area to reduce the chance of comfort problems due to control.

Partitioned offices, regardless of partition height, have their share of stratification and dead spots. Even 4 112 foot high partitions will create dead spots at occupant breathing levels. Floor square footage of 1000-1500 sq. ft. per zone, has proven to be a good workable solution in preventing comfort problems due to attempting to control air flow from single control points, in large occupied areas.

87

6.7.

Interior open office ventilation zones designed for over 2,500 ft2 of floor area. (Obviousl y, client needs take precedence.), create comfort and control problems.

Interior open spaces above 2500 sq. ft in area could require as much as 2500 to 3000 CFM or greater, depending on the actual use. This can cause significant comfort and control problems. Attempting to satisfy the number of people who may be in such large areas has proven to be vety difficult from single control points. Though the 2500 sq. ft. maximum is not written in concrete it should be adhered to as closely as possible. There is of course a cost penalty in keeping these zones appreciably below 2500 sq. ft., so a trade off of possible comfort control problems versus cost has to be evaluated.

6.8.

Do not use half-height partitions with closed bottoms.

Coordinate with architects or interior designers to specify open- bottom panels to assist air circulation.

There is a

misconception that if half partitions are utilized then proper air flow within the partitioned space is guaranteed. This is not the case. Actual field tests have shown that even within the so-called half partitions, you must have an open bottom (2" to 4") in order to facilitate proper air mixing.

88

6.9.

Final filters on the discharge side of the cooling coil on a blow through unit constantly get wet and cause the growth of fungi and mold which in turn creates serious IAQ problems.

If the final filters are less then 10 to 15 duct diameters downstream of the coils it is possible that condensate blow by could deposit on the final filters before the droplets have a chance to re-evaporate.

To cure this problem a moisture

eliminator must be installed before the filter bank.

In some

cases this moisture eliminator can also be a combination eliminator

and

diffusion

plate

and

in

this

case

the

requirements for a diffuser plate installation must also be complied with.

89

Following is a moisture eliminator curve:

Moisture Separator Efficiency 100

n-·

-f--T I

' I

---

". (.)

c Q) ·0

I

j! 't-1 ,_Jl

80

60

I

t

ｾ＠

Ll.J

--1-

c 0

ｾ＠

u

40

I

0

(_)

j _ ___i__

20

I

I

I___J_

0 2

3

4

5

6

7

---i

!

I

I

I __ j

--- --8

9

10

11

12

13

14

15

Droplet Size - Microns 500 FPM Face Velocity

Fig. 6.1 Note that the size of the moisture droplets coming off of the cooling coil will average around 10 microns.

In this range,

such an eliminator as above will have a fairly good efficiency. If eliminators are utilized by themselves you must still limit their approach velocity to 500 fpm, for maximum collection efficiency.

90

16

6.10. Dirt build-up on reheat coils within terminal boxes creates serious interior air quality problems.

Many terminal boxes are installed without provisions for cleaning the reheat coils.

In these cases you must install

access doors, both upstream and down stream of the reheat coils.

The reheat coils should be inspected and cleaned at

least two times a year, depending upon the type of installation in which they are installed.

6.11 . Contamination from exhaust fan discharges work their way back into the building in violation of EPA and OSHA requirements .

Of course there are many causes of exhaust streams recirculating requirements

back

into

a

stipulate

that

building.

OSHA

contaminated

and

air must

EPA be

discharged outside of the building's recirculating zone. One of the main items which must be checked in assuring that a discharge is outside of the building's recirculation zone is the discharge air stream's momentum flow and what is called the effective stack height.

There are many software programs

available to measure these items. One of the basic equations that can be utilized to calculate these values are shown in Fig. 6.2.

91

Effective Stack Height For Vertical Discharges Only Momentum Flow:

Ta

F m

T

Effective Stack Height:

T = Exit Temperature

=R0

T8 =Ambient Temperature V

=R0

=Exit Velocity =fps

R = Stack Outlet Radius = ft he h8

=Effective Height =Stack Height

hb = Building Height

Fig. 6.2 As a ballpark guide, if the effective stack height is greater then the height of the recirculating zone then the exhaust air will be discharged outside of the zone.

92

If it is less then the

recirculating zone, then the discharged air will be within the zone and in violation of State, CDC and OSHA regulations.

6.12. Excessive buildup of Nitrous Oxide in a Hospital's Operating room results in serious occupant health problems.

Recent studies have shown that all too often when Nitrous oxide is used as the anesthetic gas, the ambient concentration builds up far in excess of

allowable limits.

recommended limit at this printing is 25 ppm.

The NIOSH For proper

control of anesthetic gases in an operating room, the room must be provided with a proper scavenging system. If this is a room utilized mostly for oral surgery, in addition to a scavenging system, an auxiliary exhaust system must be provided. (see NRC, PUB #94). The auxiliary exhaust system is normally set to activate if the maximum concentration of Nitrous Oxide exceeds 25 ppm.

The system can also be

manually activated to reduce concentration levels.

6.13. HVAC systems which exhaust contaminated air often require a Bag In/Bag Out filtering system. Besides many of these systems fail ing to actually contain contamination , studies also show that in one survey, 50 % of the Bag In/Bag Out systems were found to never have had the bag installed. When operators went to change the original bag, they were shocked,

93

when they opened the containment cover, to find that there was no bag installed.

1. When a Bag In/Bag Out system is utilized it is important to specify it properly and to specify the testing of the unit with proper documentation.

And, in some cases, witnessing

should take place.

2. A generic specification for a Bag In/Bag Out unit and installation at has proven to be reliable is as follows:

Filter Housing: The filter housing shall be a Bag In/Bag Out type to allow replacement of contaminated filters while minimizing exposure to maintenance personnel. Housing shall be manufactured from 11 and 14 gauge T-304, unpainted stainless steel (or other approved material suitable for the environment in which the housing will be utilized) and shall be adequately reinforced to withstand a negative pressure of one and a half ( 1.5) times the expected operating pressure or IO" water gage if the exact operating pressure is unobtainable. The housing shall be side access for filter installation/change-out and shall accommodate standard gasket filters that do not require special attachments or devices to function properly in the housing. All pressure retaining weld joints and seams shall be continuous ly welded. All welds shall be wire brushed and/or buffed to remove heat discoloration, burrs, and sharp edges. Welds that are a portion of gasket sealing surfaces shall be ground smooth and flush with adjacent base metal. All welding procedures, welders, etc. shall be ASME certified and qualified. All hardware on the housing and mechanical components of the filter clamping mechanism shall be stainless steel except for the cast aluminum access door knobs and the brass threaded blocks in the filter clamping mechanism. The crank operated filter clamping mechanism shall include pressure bars with pre-loaded springs that

94

exert a sealing force of 1400 pounds per 24" x 24" filter which is applied as even, uniform load, along at least 80% of the top and bottom of each filter frame.

Bagging: The housing shall have a bagging ring around each filter access port. The bagging ring shall have two continuous ribs to secure the plastic change out bag and the ring shall be hemmed on its outer edge to prevent the bag from tearing. One PVC change out bag shall be supplied with each filter access port. The bags shall be PVC and shall be at least 8 mil thick. The bags shall have a 114" diameter elastic shock cord hemmed into the open end so when stretched around the bagging ring, it provides a secure fit. Bags shall include approximately 12" of transparent PVC at the open end and three glove sleeves built into the body to assist in filter change out. To prevent bag detachment during the change out operation, one nylon security strap shall be provided with each access port. Each filter access port shall be sealed by a removable access door which is designed to seal on the filter housing face and not the bagging ring. Access doors shall have a perimeter edge extruded silicone gasket. Prefilter: Prefilter size and capacity shall be as shown on the contract drawings. Average efficiency shall be 80-85% per ASHRAE 52-76 and approximate initial pressure drop shall be .12" water gage at rated flow. Prefilter shall be UL class 2. HEPA Filter: HEP A filter size and capacity shall be as shown on the contract drawings. HEP A filter shall be individually tested, labeled and certified to have a minimum efficiency of 99.97% on .3 micron DOP when tested in accordance with MIL-STD-282. HEP A filter shall have all glass water proof media, aluminum separators, polyurethane sealer, and a 16 gage galvanized frame (all metal materials shall be as required to prevent corrosion in the internal or external environment the filter will be utilized in) with a 1/4" thick closed cell neoprene gasket on one face. The HEP A filter shall meet UL 586, UL 900 class 2. HEPA filter shall have an approximate pressure drop of 1" water gage at rated air flow. Factory Quality Assurance and Testing: The filter system shall be designed, manufactured, and tested under a Quality Assurance Program that complies with the requirements of ANSI/ASME NQA-1, Prior to leaving the factory, the completed, ready to ship unit's housing shall be tested as follows:

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1. Each filter sealing surface of each housing shall be checked with a flatness gage to guarantee the flatness tolerance for filter mounting frames as recommended in ERDA 76-21, "N uclear Air Cleaning Handbook, Table 4-2". 2. Each housing shall be tested for filter fit using a test fixture that is the maximum size allowed by the tolerances given in the standards governing the actual filters that will used in the housing. With this test fixture installed in the housing , the filter clamping mechanism shall be checked to ensure that all components are operating properly. 3. Use an air tight gasket fixture that simulates an actual filter so the filter sealing surfaces can be tested for leak tightness. With test fixture in place, the filter sealing surface shall be evacuated to negative 1O" water gage (i.e. pressure is such that it tries to pull the fixture away from the sealing surface). The maximum allowable leak rate shall be .. 0005 CFM per cubic foot of tested volume. 4. The inlet and outlet of each filter housing shall be blanked off to allow the entire pressure boundary of the housing to be leak tested. The maximum allowable leak rate shall be ..0005 CFM per cubic foot of tested volume. 5. All above testing shall be in accordance with ASME N5 l 0.

Field Testing: To verify system integrity, and to meet CDC Guidelines, an in-place DOP test shall be performed on the installed system. Test shall be in accordance with ASME N510. Complete documentation shall be provided to the owner. The Testing And Balancing Agency shall confirm that the unit is in proper operating condition and that the bag has been installed in accordance with the manufacturer's recommendations. The design engineers shall witness all tests and the entire installation at their discretion. All required documentation shall be submitted to the engineer for approval.

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Chapter 7 Laboratories

7.1

Utilizing a pressure dependent VAV system in a large facility, causes large variations in the air flow provided to spaces.

Where duct pressure variations are greater than 1" H20. and especially where the supply system will exceed 10,000 CFM peak capacity, a pressure-independent

system must be

utilized to assure proper control and operation.

7.2.

Fume hoods in a laboratory can not get sufficient makeup air and/or they can't maintain proper face velocity during actual laboratory use.

It is advisable when designing a laboratory that the number of fume hoods installed in the room not exceed 2 to 3 hoods for every 500 square feet. When laboratories are designed with a much greater hood density than this, it is difficult to maintain proper air flow in the room without exceeding the 50% terminal air velocity guide. That is, the terminal air velocity within the room, in the vicinity of the fume hood face, should not be greater then 50% of the actual hood face velocity.

If the 50%

guide is violated you often find that there are competing air flows within the room, which cause the hoods to be starved for

97

make-up air; even though calculations might theoretically show that there is good balance

7.3.

Although the laboratory is under negative pressure as related to its surroundings, contaminated air leaves the room whenever the door is opened .

It is a mistake to think that just because a room is deigned for negative pressure,

it will always keep contamination from

leaving the room. To keep contamination from escaping from a room which is maintained under negative pressure, it is important to maintain the proper inrush velocity.

Under

normal conditions, bi-polar flow is what causes air to escape from

a room

which

is

maintained as

negative to

its

surroundings. (see Fig. 7.1) At the time of this writing, ANSI Z9.5 recommends that in order to assure that a room stays negative, even when doors are opened or when a pass through sliding window is opened, the inrush velocity through these openings must be maintained at no less then 50 fpm. It further goes on to say that 100 fpm would be even better.

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Bi-Polar Flow Across A Doorway

t

I

30FPM

9.--

Doorway

Negative Height Of oorway \

. 25FPM :

-,

Positive

ｾＧ＠

Side

I

ｉｾ＠

)

:

Side

I

25FPM

- - - - l 12 FPM

I Outward

Inward

Bi-Polar flow across a doorway separating a negative room from its surroundings. Numbers shown are for example purposes only.

Fig. 7.1

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of variable frequency drives on the fume hood exhaust fans. The lower the sash is positioned, the less

the air flow.

In

some cases when the sash is totally closed, the air flow will be zero.

In most cases, unless the supply air flow is also

adjusted, the building or at least the laboratory, will experience large air imbalances. When the damper method is utilized to control the air, it is rare that the timing could be such that the supply is adjusted in time to prevent the air imbalance.

In

cases where the damper timing problem is creating the air imbalance problem, a more sensitive type of control could be utilized such as the TS/ or Modus type of control, or the hoods involved could each be supplied with their own individual fan. In this way the individual fans can be controlled relative to the sash position of their respective hood. In extreme cases, the hoods could be replaced with bypass hoods and separate heat recovery in the exhaust.

7. 6.

Fume hoods cannot remove parti culate matter generated by the process taking pl ace within the hood .

Fume hoods and the exhaust duct from the fume hood, are usually not designed to handle particulate matter from certain types of experiments taking place within a fume hood. Fume hoods utilize basically low face and entrant velocities (100 fpm., Such velocities cannot remove large quantities of low particle size particulate matter or even the smallest amount of

102

large particles, such as 100 micrometer particles.

The duct

exhausting the air from a fume hood is rarely designed for velocities higher then 2000 fpm. Though such a velocity can entrain small amounts of small size particles, it can not entrain particles greater then 100 micrometers. Therefore, if it should turn out that the installed fume hoods do not work as they should for certain types of processes used in the hoods, the type of process should be reviewed to see if the process gives off excess amounts of very small particles or even small amounts of large particles.

7.7.

Fume hoods which appear to be designed and installed properly, periodically malfunction by not exhausting th e fumes within the hood.

Hoods depend a great deal upon what is happening around them to assure that they will operate properly.

It is very

important that the terminal velocity in the vicinity of the hood not be greater then 112 of the hood's operating face velocity. For best operation, the terminal velocity in the vicinity of the hood should actually be closer to 20% of the hoods operating face velocity.

If the terminal velocity is allowed to be higher

than these values, then even though the hood may be tested at

a face velocity of 100 fpm (for example), it will not provide the exhaust power due to the turbulence at the face of the hood. Even air currents from people walking by the hood could disturb the hood's overall exhaust efficiency. The fume hoods and the room must be laid out so that the hoods will not be 103

affected by furniture

layouts,

high

supply air terminal

velocities and people traffic. In-place tracer gas testing should be performed after installation under several different load modes so as to determine the actual, as installed, exhausting efficiency of the entire room (or building) not just one particular hood.

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