Subject: Re: Help me with this load calc, please ?

From: [email-masked] (paul milligan)
Date: 29 Jul 1995 19:40:53 GMT

Oops...

	I just realized a basic error I made by adding the reheat to
total load.  My actual proposed total load should be 4.17 ton, not 4.74.

	Please continue to educate me from there.  :~)

Paul

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From: [email-masked] (paul milligan)
Date: 29 Jul 1995 18:59:44 GMT

Here's one I just got through working on....

	This is a WEIRD system. 100 % roof-top air is sucked into a 
water-based air scrubber ( we haven't got to the fan yet, so the pressure 
here is negative ).  There is a pre-filter ( 90% pleat ) and steam pre-heat 
coil ( for winter only, not in effect right now ), both of which I'll 
ignore.  :~)

	After the water-scrubber, the air is sucked into the fan, and then 
expelled into  ~ 10 running linear feet of solid charcoal, HEPA, and other 
bio-active mixed-media filtration, in an 18" x 18 " plenum.  Gross static 
pressure reaches 18" water column due to restriction of water-scrubber and 
filtration, but I'll ignore that, as final WC is ~ 1", therefore I take all 
heat of compression / expansion to be a net wash for my purposes when I 
then get to the heat/cool segments at a still-strong but normal static.

	After the above, there is a 2 ton R22 AC unit to 'pre-cool' the 
roof air.  There is then a 5 ton Carrier R22 residential scroll AC, and 
then a 4KW electric element.

	Final target of the system is 70 F air at 50 % - 60 % RH, year 
round, at 500 CFM.  Yes, I said 500 CFM.  

	I just got through tuning and repairing the system such that it 
holds 70 F +/- 0.5, and 55 % RH +/- 5 %.  I hope to get that RH a little 
tighter next week, tweaking the Faircloth multiplier down just a bit.

	Controls are as follows, in equipment sequence : Steam pre-heat is 
on a Penn OA stat for winter pre-heat only.   Scrubber sump temp is 
whatever evaporation and OA and makeup temp drive it to, no control.

	 2 ton unit runs continuous, no control.

	5 ton unit has a variable hot gas bypass, which is controlled by 
pneumatics, which sense RH only.  RH sensor ( standard 3 - 15 lb) goes to 
a Barber-Coleman prop-band controller, then to a Faircloth variable 
multiplier, with final pneumatic output on the HG bypass being 0-80 lbs 
pneumatic control pressure, and minimum HG bypass pressure set at ~ 60 lbs 
suction to the 5 ton unit, to prevent ice.  TXV is set to provide minimum 
compressor cooling when in maximum bypass ( Andy, you're just gonna have
to let me slide on this one  :~) ).

	Electric reheat is under DDC PID control, for sensible heat only 
( thermistor input ).  There is <NO> interconection between the controls of 
any one piece of equipment to any other piece.

	I corrected some initial problems with this 'rig', such as : The 
low-level pneumatics were being run at 35 lbs ( instead of 20 ).  The high 
level pneumatics were being run at 120 lb ( instead of 80 ).  The RH sensor 
was <BETWEEN> the 5 ton coil and the reheat element, 6 inches from the evap 
coil, and routinely had liquid water dripping off it ( once the water falls 
<down>, does it become 'low' humidity ? :~)  ). The sensible thermistor was 
~6 inches downstream from the reheat element ( I measured >20 F laminar
stratification delta in this region ).  After I regained consciousness, 
I moved both sensors 30' downstream.  The 5 ton unit was running 300 lb 
head on a cool day, 400+ on a hot day.  Unit would cut out on head and 
thermal regularly.  Unit has had 3 compressors in 2 years ( whole system 
is two years old ).  Sight glass NEVER came even close to clear with these 
pressures, and unit has NEVER had a clear sight glass in 2 years, from what 
I've been told.  There is a solenoid in the liquid line, between the guage 
ports and the condensor.  After I evacuated the system with this solenoid 
<OPEN> ( which I don't think had ever been done ), the system now runs a 
clear glass with a 260 head on a hot summer day.  Ah, yes, the glories of 
air ( or nitrogen ) in a system !  :~(

	So anyway, on to my question...

	My intuition told me that 5 ton was hugely oversized for <ANY> 500 
CFM load.  Then I ran the numbers, and the numbers tell me that my 
intuition needs a long vacation.  Would some kind ( engineering type ) 
soul care to comment on the numbers below, please ?  This is a bit out of 
my usual area.  Please remember that some of the below are approximations 
>from  my psychrometric chart, I just want to know if my understanding is 
somewhere in the ball park.

Clean air supply system - sample load

If setpoint  =  70 F ,  50 % RH  
If input air  =  85 F,  85 % RH  
Volume is ~ 500 CFM at all times

Input = 85 F, 85 % RH = 50 btu / lb, @ 14.25 cu.ft. / lb = 105,623 BTU/HR

	because :

1 / 14.25 cu. ft. / lb = .0701... lb / cu. ft.
0701 lb / cu. ft. x 500 CFM = 35 lb / minute
35 lb / minute x 50 BTU / lb = 1754 BTU / minute
1754 BTU / minute x 60 = 105,623 BTU / HR enthalpy of input air

	MINUS

Ouput = 70 F, 50 % RH = 25 btu / lb, @ 13 .5 cu.ft. / lb  =  55,555 BTU/HR

	because :

1 / 13.5 cu. ft. / lb = .0740... lb / cu. ft.
0740 lb / cu. ft. x 500 CFM = 37 lb / minute
37 lb / minute x 25 BTU / lb = 925 BTU / minute
925 BTU / minute x 60 = 55,555 BTU / HR enthalpy of output air

PLUS :

Assume : 4 KW re-heat element runs at 50 % duty cycle
	      to maintain sensible at 70 F, under PID control

Thus 4,000 W x 3.41 BTU / W = 13640 BTUH, x 50 % cycle = 6,820 BTU / HR 

105,623  -  55,555  +  6,820  =  56,888 BTU / HR total enthalpic load

56,888 / 12,000 = 4.74 TON total combined sensible and latent load.

Therefor, 5 ton ( plus capacity from the prior 2 ton ) is rock and roll city.

I eagerly await the input of those who do this stuff for a living.

You'll have to excuse me now, I think I'm getting a headache.  :~)

Paul

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From: Dave Munro <[email-masked]>
Date: Wed, 2 Aug 1995 14:20:56 GMT

I don't have my psych chart with me, and don't know what the altitude/ barometric 
pressure is for this discussion, so I will not address the math in this post as much 
as some of the basic questions and statements.

1. Allowances for fan heat need to be included. If the fan motor is in the air 
stream the motor heat must also be added. (The purpose of the 2 ton cooling coil may 
have been to counteract the fan heat???). It may be useful to verify the airflow and 
take entering and leaving air temperatures.

2. Unless there is active control of the air flow rate, the assumption that the 
system is flowing 500 cfm probably is not true. The actual air flow rate will be 
determined by the fan speed, the flow rate will vary depending on the inlet 
conditions (barometric pressure and air density).

3. Are we controlling the room or the supply air to the 70 F and 50-60% RH design 
criteria. What are the estimated minimum and maximum sensible and latent room heat 
gains?

Based on the results of tuning and repairs, it looks as though you have addressed 
most of the issues relative to cooling. There may be issues in the future related to 
the heating and/or intermittent weather conditions:

1. What are the steam coil control parameters?

2. Does the air scrubber run continuously (24 hours per day 365 days per year)? Do 
you know its saturation efficiency? Can you measure the entering dry bulb and 
wetbulb conditions as well as the leaving dry bulb?

3. Does the reheat coil have SCR control, or are their multiple stages?

Please excuse my ignorance - What is a Faircloth variable multiplier?

Dave Munro

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From: Kirk Wilson <[email-masked]>
Date: 2 Aug 1995 14:24:13 GMT

Uh!...a picture of this one would be helpful. Here is a quick looksy.

When I do load calcs (it's been a while), I use the ASHRAE Load Calculation Manual 
or computer progams based on it. Your system has the following loads:

1. Outside air from 85/85% [45 BTU/Lb] to room condition 70/50% [25.5 BTU/Lb]
    Use BTU/hr = 4.5 (CFM) (Delta h) = 44,000 BTU/hr  (3.7 tons)
(Actually your washer saturates the air to 81/99%,constant wet bulb, but the 
enthalpy is about the same)

2. Fan Heat from Table 4, Page 26.10 of 1993 ASHRAE Fundamentals
   2 HP motor (if this is size) = 6,440 BTU/Hr (motor in/ fan in)

3. Reheat Load, if it is a constant. [BTU/hr = 1.1 (CFM) (delta temp)]

3. Room Load - You've got to define these; lights, people, equipment, skin, etc.

Add all of these up for your cooling load. With your supply airflow fixed at 500 
CFM, your sensible load will define your supply air temperature. 

Your outside air load uses 3.7 tons of your 7 tons on a design day.

Hope this helps.

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From: [email-masked] (paul milligan)
Date: 2 Aug 1995 21:37:02 GMT

In article <[email-masked]>,
   Dave Munro <[email-masked]> wrote:

~~>1. Allowances for fan heat need to be included. If the fan motor is in the air 
~~>stream the motor heat must also be added. 

	Very true, but the fan motor is not in the airstream.  I think I
can ignore gain from bearing friction.

~~>2. Unless there is active control of the air flow rate, the assumption that the 
~~>system is flowing 500 cfm probably is not true. 

	A good point I had missed.  Very true.  Also, the scientist types
who use the test chambers fed by this system have balancing dampers under
their control.  500 is the design spec for max load.  I really have no idea
how close to that they run, so I take that as my max load.

~~>3. Are we controlling the room or the supply air to the 70 F and 50-60% RH design 
~~>criteria. What are the estimated minimum and maximum sensible and latent room heat 
~~>gains?

	Supply air at those numbers in the supply duct.  There are no
gains ( at least that I need to deal with ) in the test chambers, because
this is a single pass, 100 % roof air system.  BTW, it has a special evap
for extra latent bias.

~~>Based on the results of tuning and repairs, it looks as though you have addressed 
~~>most of the issues relative to cooling. There may be issues in the future related to 
~~>the heating and/or intermittent weather conditions:

	Yeh, she barcharts for a 5 day period now at +/- 0.5 F, +/- 5%, each
centered on set point.  Acceptable to the scientists using the chambers, so
I can live with it. 

~~>1. What are the steam coil control parameters?

	Don't know, it's for winter anti-freeze pre-heat only, so I haven't
looked at it yet.  Roughly, I think it maintains a minimum 45F intake air.

~~>2. Does the air scrubber run continuously (24 hours per day 365 days per year)? Do 

	Yes

~~>you know its saturation efficiency?

	NO, and a major question in my mind.  I can only SWAG it at 85 -95 %

~~>Can you measure the entering dry bulb and 
~~>wetbulb conditions as well as the leaving dry bulb?

	Around the system, yes, but not around the scrubber by itself.

~~>3. Does the reheat coil have SCR control, or are their multiple stages?

	Yes, SCR on a PID controller.  That seems to work very well.  The
problem is no interaction between it and the RH control.  Each one just
tries to do it's own thing, regardless of the other.

~~>Please excuse my ignorance - What is a Faircloth variable multiplier?

	Your ignorance is vastly exceeded by my own.  :~)  The Faircloth
is a pneumatic device that multiplies it's control input by a set amount.
I think ( not sure ) this is also called a ratio relay by some other brands.

	The difficult thing in the RH control is that the sensor is 3-15,
then the Barber-Coleman is 0-20 output into the Faircloth, wich takes
that 0 -20 and multiplies it proportionaly ( as a throttling range kind of
thing ) to become 0 - 80 lbs output to the control head of the HG bypass.

	This whole rig series is REAL sensitive, and oscillates a bit 
no matter what I do.

	Given this system being a 'gross enthalpy' kind of thing, with
no 'room' loads and no recirculation of any air, does that help look at
the loads ?

	Thanks, Dave, for looking at this one !  :~)

Paul