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Oil-free centrifugal refrigeration compressors: from HFC134a to HFO1234ze(E) Joost J. Brasz Danfoss Turbocor Compressors Inc. Syracuse University CASE Incubation Center 2-212 Center for Science and Technology Syracuse, New York 13244-4100 USA [email protected]
ABSTRACT
The high global warming potential of HFC134a (GWP=1300) has led to the
development of a new family of man-made refrigerants with much lower global
warming potential. HFO1234ze(E) is one of those fluids. It has a GWP value of 7
and has been in commercial production as a blowing agent for a few years. Based
on cost and availability this new fluid has been selected a potential candidate to
replace HFC134a in commercial chillers.
A new family of oil-free direct-drive centrifugal compressors with HFO1234ze(E) as
working fluid has recently been introduced commercially, covering a cooling
capacity range from 200 – 300 kWthermal. These new compressors are a spin-off of
an existing platform of oil-free HFC134a products. Being oil-free eliminated oil-
refrigerant compatibility issues, which were a major stumbling block during the
transition from CFC12/HCFC22 towards HFC134a in the early 1990’s.
Due to its somewhat lower pressure and vapor density HFO1234ze(E) requires in a
slightly larger fluid module to achieve equal capacity as HFC134a. Impeller tip
speed is reduced at equal temperature lift (=difference between condenser and
evaporator saturation temperatures) as a result of the lower sonic velocity of
HFO1234ze(E). Overall compressor efficiency improves as a consequence of these
two effects.
1. INTRODUCTION
During the nineties, the air-conditioning and refrigeration industry saw a transition
from CFC’s towards HCFC’s and HFC’s. The main driver for this transition was the
discovery of stratospheric ozone layer depletion by the chlorine atoms found in
CFC’s and HCFC’s.
After the resolution of the ozone layer depletion problem by the introduction of
chlorine-free HFC refrigerants, the environmental concern started to focus on the
global warming impact of these man-made refrigerants. Legislation intended to
Tcond,sat = 35.6 0C
Tevap,sat = 5.6 0C
29.4 0C 34.7 0C
12.2 0C 6.7 0C
Refrigerant ODP GWP
CFC12 1 8500
HCFC22 0.05 1700
HFC134a 0 1300
HFO1234yf 0 4
HFO1234ze 0 6
limit the future use of refrigerants in systems with a high direct effect on global
warming has led to the development of new systems using natural refrigerants such
as CO2 as well as the development of new man-made refrigerants with much lower
global warming potential than the currently used HFC refrigerants.
Table 1 shows a number of medium pressure refrigerants used in water-cooled
chillers. The refrigerants have transitioned from fluids with high ozone depletion
potential (ODP) such as CFC12 and HCFC22 towards fluids with zero ODP but still
a large global warming potential (GWP) such as HFC134a. Today there are
candidate refrigerants with both zero ODP and very low GWP. HFO1234yf is the
result of a joint effort by DuPont and Honeywell to develop a drop-in replacement
fluid for HFC134a for mobile air conditioning (MAC) applications [1].
HFO1234ze(E) is a low GWP fluid developed by Honeywell and currently used as a
foam blowing agent [2].
Table 1. Ozone layer depletion (ODP) and global warming potential (GWP) of
various fluids.
2. CENTRIFUGAL CHILLERS AND THEIR OPERATING CONDITIONS
In the refrigeration industry centrifugal compressors are predominantly used in
water-cooled chillers although smaller capacity direct-drive centrifugal compressors
are now also applied to air-cooled chillers. Figure 1 shows a simplified equipment
diagram of a water-cooled chiller. A representative full-load operating condition for
a commercial water-cooled chiller is to reduce the temperature of the water returning
Figure 1. Simplified equipment diagram of a water-cooled chiller
from the air-handling equipment. Typical full-load entering and leaving evaporator
water temperatures as specified by the ARI rating code [3] are 2.220C (54
0F) and
6..670C (44
0F), respectively The heat absorbed by the evaporating refrigerant is
rejected in the condenser together with the heart of compression. This heat increases
the temperature of the water returning from the cooling tower and entering the
condenser at 29.440C (85
0F) to a leaving water temperature of around 34.72
0C
(94.50F) – the exact number depending on the [deal cycle efficiency of the
refrigerant as well as the compressor efficiency. The evaporator saturation
temperature will be slightly below the leaving chilled water temperature of 6.670C
(440F), say 5.56
0C (42
0F) and, similarly the saturation temperature of the refrigerant
in the condenser will be just above the leaving condenser water temperature of
34.720C (94.5
0F), say 35.56
0C (96
0F).
Overall chiller efficiency is defined using of a coefficient of performance (COP)
number which is defined as follows:
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������������ ���!��������"��#"�
For refrigeration systems this number is always larger than , which is the reason that
the word efficiency has been replaced with the term coefficient of performance. In
the US is is customary to specify chiller performance in terms of a kW/ton number
defined as:
��/�� � ������������ ���!��������"��#"�
��������� ��������%��%���������
Since a ton of refrigeration (defined as the cooling capacity obtained from the
melting of a so-called short ton (=2000 pounds) of ice in 24 hours) is equivalent to a
cooling capacity of 3.52 kW, the relationship between these two quantities is as
follows:
��� � 3.52
����
Actual state of the art COP numbers for water-cooled centrifugal chillers at ARI
full-load conditions vary 6.18 (0.57 kW/ton) to 6.77 (0.52 kW/ton). This variation is
caused by the choice of refrigerant, compressor efficiency and refrigeration cycle
details such as sub-cooling and economizing. Advertised centrifugal chiller
performance figures can be up to 7.5% percent better as a result of the measurement
tolerances allowed by the ARI-550 test code, including a 5% deviation in power and
flow and a 0.28 oC (0.5 F) deviation in temperature.
3. NEW REFRIGERANTS IN OIL-FREE COMPRESSORS
3.1 Oil-free compressors allow easy transition to new refrigerants
Oil-free centrifugal compressors can relatively easy be converted to alternative
refrigerants. The oil-free compressor operation eliminates the identification or
development and subsequent qualification of lubricants that are compatible with the
new refrigerant. The new refrigerant compatibility studies are limited to the
elastomer (O-rings) and motor insulation materials.
3.2 Chiller global warming impact of HFO1234yf and HFO1234ze(E) The global warming impact of a refrigeration system consists of at least two effects:
- a direct effect (representing the global warming due to the leakage of refrigerant
molecules into the atmosphere) and
- an indirect effect (representing the amount of global warming as a result of energy
utilization of the refrigeration system which is related to its efficiency).
The industry has accepted the TEWI concept (Total Equivalent Warming Impact) [4]
to account for both effects. For systems notorious for their high refrigerant leak
rates such as commercial refrigeration and automotive air conditioning systems the
direct effect is the predominant cause of global warming. For residential and
commercial air conditioning equipment with its much lower leak rates the indirect
effect is the main contributor of global warming, with the exact amount of global
warming being determined by the percentage of power being generated by coal, oil
and natural gas versus renewable and nuclear power generation. As a result the
cycle efficiency of a low GWP refrigerant relative to the one it replaces is an
important characteristic. It can be categorically stated that for commercial and
residential HVAC equipment the cycle efficiency of a low GWP alternative
refrigerant should not be inferior to the existing refrigerant it is intended to replace.
Table 2 compares the ideal (= assuming 100% compressor efficiency) coefficient of
performance (COP) for typical water-cooled chiller conditions, using the latest
Refprop 9.0 refrigerant properties [5]. As can be seen from that table HFO1234yf
has a 3.2% lower COP than HFC134a and should therefore to be eliminated as a low
GWP alternative for commercial HVAC applications where the energy consumption
is a major contributor to global warming.
HFO1234ze has an ideal COP that is equal to that of HFC134a. A transition from
HFC134a to HFO1234ze does not increase the indirect global warming – like a
transition towards HFO1234yf would – and dramatically reduces the direct global
warming effect.
Table 2. Ideal cycle COP comparison for HFC134a, HFO1234yf and
HFO1234ze(E) showing an almost equal COP for HFC134a and HFO1234ze(E)
and a 3.2% lower COP for HFO1234yf
Input values in blue. Black input values derived from blue ones. Output values in red.
CYCLE CONDITIONS
Evaporature saturation Tevap,sat 5.56 [0C] or 42.0 [
0F]
Evaporator superheat ∆∆∆∆Tevap,sup 0.00 [0C] or 0.0 [0F]
Condenser saturation Tcond.sat 36.11 [0C] or 97.0 [0F]
Condenser subcooling ∆∆∆∆Tcond,sub 3.89 [0C] or 7.0 [0F]
Refrigeration capacity 1232 [kW] or 350 [ton]
Compressor specific speed ns 0.76 [-]
Compressor specific diameter ds 3.40 [-]
Fluid R134a R1234ze
Speed of sound m/s 146.71 139.07
∆∆∆∆hevaporator kJ/kg 156.89 144.13
∆∆∆∆hs,compressor kJ/kg 19.55 17.98
mdot kg/s 7.85 8.55
density kg/m3 17.45 14.20
Vdot m3 /s 0.450 0.602
N (Impeller Speed) rpm 17891 14524
D (Impeller diameter) m 0.193 0.228
u2 (tip speed) m/s 181 173
Pr(Pressure ratio) - 2.57 2.60
3.3 Predicted centrifugal compressor capacity change when replacing HFC134a with HFO1234ze(E) The refrigeration capacity of a compressor is the product of the evaporator enthalpy
rise and the compressor mass flow rate. Table 3 shows the inlet volumetric flow
rates required to achieve 350 ton of cooling. As can be seen HFO1234ze(E) requires
a 33% (.601/.450) larger volumetric flow rate. Since compressors are essentially
constant flow machines, drop-in compressor behavior will result in a capacity
shortfall of 25.2% for HFC1234ze(E). However, drop-in of alternative refrigerants
in centrifugal compressors requires a speed adjustment to insure surge-free operation
without over-compressor (choke) in order to maintain compressor peak efficiency
and turn-down capability [6]. As can be seen from Table 3, the isentropic enthalpy
rise needed to compress the refrigerant from the evaporator saturation temperature of
5.56 0C to the condenser saturation temperature of 36.11
0C is quite different for
these refrigerants. The head to be delivered by a compressor designed for HFC134a
(19.55 kJ/kg) is 9% larger than that required for HFO1234ze(E) (17.98 kJ/kg).
Since head is proportional to the square of speed, compressor drop-in applications
require a speed reduction of 4% for HFO1234ze(E). Since for pressure ratios around
2.5 the compressor flow varies with compressor speed to the power 1.6 [7], and
additional capacity reduction of 7% is anticipated as a result of the speed reduction
needed for optimum compressor performance.
As a result the capacity of an existing R134a compressor will be reduced by 32.2 %
(25.2% + 7%) for HFO1234ze(E). Since the platform of the four existing HFC134a
centrifugal compressors come in frame sizes that have upward capacity jumps of
about 50% (= downward capacity jumps of about 33%) most existing R134a
compressor applications could be fulfilled with HFO1234ze(E) by just selecting the
next available compressor frame size.
Table 3. Calculation of thermodynamic properties used to predict the
compressor performance change in terms of capacity and speed when
switching from HFC134a to HFO1234ze(E)
3.4 Maintaining refrigeration capacity by moving up a frame size when replacing HFC134a with HFO1234ze(E) Moving up a compressor frame size has an efficiency advantage. Compressor
efficiency increases with impeller size - both due to Reynolds number effect and
reduced relative surface roughness. For the relatively small, high-speed centrifugal
compressors in the 200 to 700 kWth capacity range, peak efficiency increases about
2% for each jump in frame size. Figure 2 shows the increase in efficiency when
going from one frame size (TT300) to the next larger frame size (TT350).
Figure 2. Comparison of aerodynamic efficiency for two adjacent compressor
frame sizes. Running the larger TT350 compressor with HFO1234ze(E) and comparing its
performance against the performance of the TT300 compressor with HFC134a
results in equal refrigeration capacity with a 2% boost in compressor efficiency for
HFO1234ze(E) as shown in Figure 2.
Figure 3. Comparison of aerodynamic efficiency for a TT300 DTC compressor
with HFC134a against a next frame size TT350 compressor with HFO1234ze(E)
assuming identical aero efficiency for HFO1234ze(E) and HFC134a.
A remaining question is whether compressor efficiency is affected by the change
from HFC134a to HFO1234ze(E). Compressor performance is controlled by many
factors. At identical impeller tip Mach number (u2/a0) we should expect identical
performance only to be corrected for differences in frictional losses. The 5.8% drop
in actual impeller speed required for head and flow factor similarity means that all
fluid velocities will be 5.8% lower when the compressor is running with
HFO1234ze(E) compared to HFC134a. Given the identical vapor kinematic
viscosities of HFC134a and HFO1234ze(E), it is to be expected that the frictional
and mixing losses which are proportional to the square of the fluid velocities will be
reduced by 12.0%. Assuming an 80% fluid efficiency this would mean a reduction
of the 20% fluid loss by 12.0% resulting in a 2.4 point increase in aero efficiency.
That anticipated compressor aero efficiency improvement was confirmed during
back-to-back testing of TT300 and TT350 compressors with HFC134a and
HFO1234ze(E). The test results for the TT300 compressor are summarized in
Figures 4 and 5. The 2 – 2.5 point higher aero efficiencies shown in Figure 4 will
result in a correspondingly higher head factor for equal tip Mach number lines on the
compressor map, as shown in Figure 5.
u2/a0=0.87 u2/a0=1.02 u2/a0=1.16
Figure 4. Back-to-back test results of a TT300 compressor with HFC134a and
HFO1234ze(E) at three identical tip Mach numbers showing a 2-2.5 point aero
efficiency improvement for HFO1234ze(E).
Figure 4. Back-to-back test results of the DTC TT300D compressor with
HFC134a and HFO1234ze(E) at three identical tip Mach numbers showing a
head factor increase corresponding to the measured aero efficiency improvement
for HFO1234ze(E).
CONCLUSIONS
• DTC compressor capacity can be maintained when replacing HFC134a with
HFO1234ze(E) by switching the existing fluid module to the next frame size
fluid module.
• Testing at DTC has indicated the potential of a 4.0 to 4.5 % efficiency
improvement for commercial chillers when switching from HFC134a to
HFO1234ze(E). This improvement consists of a 2% benefit obtained by
selecting the next frame size compressor combined with a 2-2.5% benefit as a
result of the apparent lower viscous losses of HFO1234ze(E) versus HFC134a
thanks to its lower impeller speed.
• This improved compressor efficiency has a bigger impact on reducing the carbon
footprint of the chiller than the extremely low GWP value of HFO1234ze(E) of 6
versus 1300 for HFC134a.
• For air-cooled chiller systems using HFO1234ze(E) a condenser redesign might
be is required to prevent excessive pressure drop as a result of the 50% higher
condenser volumetric flow rate at equal capacity that could negate the potential
performance benefit of HFO1234ze(E).
REFERENCES
1. Minor, B, Spatz, M, HFO-1234yf low GWP refrigerant update, Paper 2349
of the International Refrigeration and Air Conditioning Conference at
Purdue, July 14-17, 2008.
2. Yana Motta, S.F., Vera Becerra, E.D., Spatz, M.W, Analysis of LGWP
Alternatives for Small Refrigeration (Plugin) Applications, Paper 2499 of
the International Refrigeration and Air Conditioning Conference at Purdue,
July 12-15, 2010.
3. ARI 550/590 (I-P)-2011, Performance Rating of Water Chilling Packages
Using the Vapor Compression Cycle, Air-Conditioning, Heating, and
Refrigeration Institute (formerly ARI), 2011.
4. Sand, J.R., Fisher, S.K., Baxter, V.D., TEWI Analysis: Its Utility, Its
Shortcomings, and Its Results, International Conference on Atmospheric
Protection, Taipei, Taiwan, September 13-14, 1999.
5. NIST Reference Fluid Thermodynamic and Transport Properties Database
(REFPROP): Version 9.0 http://www.nist.gov/srd/nist23.cfm
6. Brasz, J.J., Centrifugal Compressor Behavior with Alternate Refrigerants
paper 96-WA/PID-2 presented at the 1996 ASME International Mechanical
Engineering Congress and Exhibition, Atlanta, Ga. November 17-22, 1996
7. Brasz, J.J., Variable-Speed Centrifugal Compressor Behavior with Low
GWP Refrigerants, 2009 IMechE conference on Compressors and their
Systems, September 7-9, 2009.