PTHP Baseline outperforming VAV

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Hi all,

Having some difficulty in an App. G model for a 6-floor 14,000 m2 GFA Hotel
in zone 1B (Abu Dhabi). Assuming standard design for each, should a proposed
VAV system out-perform window AC units (the baseline)?

VAV with 2 Air-cooled centrifugal chillers, chiller COP = 3.1 (2.8 if
condenser fan energy included)
Envelope - Better windows decrease heating demand by 8% rest of envelope is
minimum required for compliance

System Packaged terminal heat pump, COP = 3.52

Using Energyplus v6.0 the results are;

Site Energy in End Uses BL Proposed COP5
Space Cooling 747.5 1052.0 715.6
Heat Rejection 0.0 0.0 0.0
Space Heating 1.3 0.1 0.1
Pumps 0.0 60.0 60.0
Fans - Interior 42.2 137.7 137.7
Fans - Car park 0.0 0.0 0.0
Interior Lighting 530.5 530.5 530.5
Exterior Lighting 0.0 0.0 0.0
Service Water Heating 0.0 0.0 0.0
Receptacle/Process Equipment 146.3 146.3 146.3
Data Centre Equipment 0.0 0.0 0.0
Elevators and Escalators 0.0 0.0 0.0
Total Site Energy 1467.9 1926.6 1590.2

I have the energy for the baseline, the proposed with COP = 3.1, and then
trying COP = 5. It seems that even if we drastically increase the COP of the
chiller plant in the proposed VAV system, we still can't get any energy
savings. Is this realistic, an artifact of the simulation, or just an error
on my part?

Any guidance would be appreciated,

Marcus Jones, LEED AP

Marcus's picture
Joined: 2011-10-02
Reputation: 0

Dear Marcus,

. It seems that your Baseline system, at >5 floors, should be VAV
with a chiller instead of a heat pump (using LEED 2009).

. As modeled, your baseline DOES have a notably higher COP, so it
may well use less energy. Part load performance will have a significant
effect, so you may want to check and compare those curves.

. Also check pump delta T and total pressure drop. If you continue
with a non-chiller system, pump energy is a significant factor. If your
Baseline becomes a chiller system, pump energy will probably become similar.

. Check fan total pressure drop for each system since your fan power
is significantly larger in the Proposed system.

. In the US, it is common to apply occupancy sensors on a LEED
project and ASHRAE 90.1 allows a 10% power reduction when these are used.
You are not taking advantage of that and may want to do so.

James V. Dirkes II, P.E., BEMP , LEED AP

James V Dirkes II, PE's picture
Joined: 2011-10-02
Reputation: 203

Hi Marcus,

Also pay attention to better glass in the proposed design. In a
cooling-dominated building, better window means the heat can not easily
escape from the building. You will not easily get credit from improving your
window here.



chen yu's picture
Joined: 2011-10-02
Reputation: 200

Hi Marcus,

I agree with your selection of PTHP for the baseline, since your project
is a hotel, which is specifically listed in the notes to Table G3.1.1A
as being a residential building type.

I am not surprised that you are showing an energy penalty versus an
ASHRAE 90.1 PTHP - the cooling and fan energy is relatively low for
those units.

As James said, make sure you are accurately modeling your fan and
pumping energy in the proposed system. Unfortunately, there are no
pressure drop adjustment allowances for Systems 1 & 2.

The maximum allowable SHGC (0.25) is already pretty low for Abu Dhabi,
so your "better windows" might not have a lower SHGC than the value
required for the baseline, and you will not get a solar load reduction
in the proposed model.

Since your baseline cooling efficiency is dependent on the system
capacities, make sure you are not using too large of a baseline cooling
COP. (It should actually be calculated separately for each baseline
system.) Using the efficiency equation from Table 6.8.1D, assuming 400
cfm/ton air flow rate and removing the fan power energy from the
efficiency, I calculate COP = 2.94 for a 15,000 Btu/hr unit (for
example). There is a separate heating COP but your heating load is
negligible per your report below.



Bishop, Bill2's picture
Joined: 2011-09-30
Reputation: 0

Compared to packaged DX equipment, an air cooled chiller has two fundamental efficiency disadvantages: extra heat transfer approach and the pumping energy of the water.

Load reductions always help, but to hone in a comparison of the efficiencies of these mechanical systems, I also suggest you look closely at:

- the COP values and their corresponding rated ambient temperatures in the referenced standards and on your equipment literature

- compressor types, sizing, and curves

- maybe, thermal storage

Good luck,

Paul Riemer, PE, LEED AP

Paul Riemer's picture
Joined: 2011-09-30
Reputation: 0

HI all, thanks for the excellent feedback, just wanted to report back for
the archive;

Yes, since hotel is considered residential, I am applying PTHP.

In ASHRAE 90.1-2007 SI, table 6.8.1D, the COP equation is 3.60 - (.213 x
Cap/1000) where Capacity is in units of [kW]. This must be an error? Cap
units should be [W] or equation should be 3.60 - (.213 x Cap). Using this
equation, this limits the COP of a PTHP to between 3.15 for 2.1 kW units and
below and 2.66 for 4.4 kW and above (with COP inclusive of fan energy).
Thanks for drawing my attention back to this equation, this significantly
increases the Baseline cooling.

I agree, my pump and fan energy also needs to be correctly sized according
to pressure drop, I just used the defaults as a place-holder until detailed

Generally the point of chilled water vs. Dx having no clear advantage is a
good one, chilled water introduces greater heat transfer approach as pointed
out and requires pump energy. However, we have learned that the client is
actually only using VAV in certain zones, the guest rooms will be served by
zone terminal variable flow chilled water coils (fancoils) and AHU's with
enthalpy recovery. Unfortunately modeling this system is proving quite
challenging in EnergyPlus so I don't have any updated results yet.

Sincere thanks again for the comments!


Marcus Jones, LEED AP, M.Sc.

Marcus's picture
Joined: 2011-10-02
Reputation: 0