Mitsubishi Heavy Industries Technical Review Vol. 56 No. 4 (December 2019)
1

Development of Large Capacity Compound TypeCentrifugal Chiller and Centrifugal Heat Pump Using
R1234ze(E)JYUN MIYAMOTO*1

HIROTAKA TANIMURA*1

KAZUKI NISHII*1

HIDEHIRO TANAKA*1

TAKAHIRO KONO*1

TAKAAKI MIURA*2

Conventionally, heat source equipment for low-temperature processes in the industrial field
has used propane (R290) and ammonia (R717, NH3), but they place a heavy burden on customers
for safe operation due to their flammability and toxicity. Therefore, alternative products with lowenvironmental impact are being demanded. Currently, the move toward terminating the use ofHFCs is being accelerated with the agreement of countries around the world, such as the adoptionof the Kigali Amendment to the Montreal Protocol. In the air conditioning field, R1234ze(E), which
has low toxicity and a GWP (Global Warming Potential) of 1 or less, has been used for centrifugal
chillers used as heat source equipment for large-scale facilities. To expand the application rangeof low-GWP refrigerants in the industrial field, we have established a technology for themulti-stage compression refrigeration cycle using R1234ze(E) that enables the supply of -20°C or
lower-temperature brine and 60°C hot water. This technology can also be used for heating tower
applications.

|1. Introduction
Heat source equipment in the air conditioning field has mainly used HFCs, while naturalrefrigerants have mainly been adopted in the industrial field where environmental protection isemphasized. Heat source equipment for low-temperature processes has used R290, R717, etc. R290
has flammability and R717 is toxic and slightly flammable, so they place a heavy burden on
customers in terms of safety in operation.
On the other hand, at the 28th Meeting of the Parties to the Montreal Protocol (MOP28) heldin October 2016, the Kigali Amendment to the Protocol was adopted to prescribe the obligatory
gradual reduction of HFC (hydrofluorocarbon) production and consumption, etc., thus clearlysetting the course toward terminating the use of HFCs. To prevent global warming, refrigerants forheat source equipment in the air conditioning field are rapidly being switched to low-GWP
refrigerants. Centrifugal chillers are cooling and heating sources that are applied to district heatingand cooling, air conditioning for buildings and factories and process cooling for chemical and foodfactories. Centrifugal chillers produce chilled water using a centrifugal compressor and have acooling capacity of 150 USRt (US refrigeration tons) to 5000 USRt. Currently, centrifugal chillers
in the air conditioning field use R1234ze(E), which has a low toxicity and a GWP (GlobalWarming Potential) of 1 or less. To expand the application range of low-GWP refrigerants in theindustrial field, we have established a technology for the multistage compression refrigeration cycle
using R1234ze(E) that enables the supply of -20°C or lower-temperature brine and 60°C hot water.
This technology can also be used for heating tower applications.
This paper introduces the equipment that has started operation as a heat source for low
temperature processes.

*1

Chiller & Heat Pump Engineering Department, Mitsubishi Heavy Industries Thermal Systems, Ltd.

*2

Manager, Chiller & Heat Pump Engineering Department, Mitsubishi Heavy Industries Thermal Systems, Ltd.

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 4 (December 2019)2|2. Specifications of developed equipment
2.1Refrigerant

As a refrigerant for the developed equipment, we selected one that had a theoretical COP
equivalent to that of the conventional refrigerant and did not significantly increase the equipment
design pressure in the supply range from brine of -20°C or lower to hot water of 60°C. In addition,
the selection criteria included that the refrigerant to be used was not a flammable gas or toxic gas in
the refrigeration safety regulations of the High Pressure Gas Safety Act in terms of manageability.
The criteria also included that the refrigerant to be used had applications other than in centrifugal
chillers and a reasonable amount of production was expected in terms of availability. As a result ofcomparing the physical properties of the candidate refrigerants shown in Table 1, we selected
R1234ze(E), which had already been adopted for heat source equipment in the air conditioning
field and as a magnesium masking gas.
Table 1 Comparison of refrigerants
Refrigerant

R290
propane3

R717
ammonia0
Flammable
Flammablegas
gas
Toxic gas

R134a

R1234ze(E)<1

R1234yf

R1233zd
(E)
1GWP
1300
<1
Handling in refrigeration
Specified
Specified
Inert
Inert gassafety regulations of High
inert gas
inert gas
gas*1
Pressure Gas Safety Act
Classification in ASHRAE
A3
B2L
A1A2L
A2L
A1
Standards 34
Saturation pressure (-25°C)
0.203
0.151
0.106
0.077
0.123
0.014
[MPa.abs]
Saturated gas density (-25°C)
4.63
1.30
5.51
4.43
7.17
0.88
[kg/m3]
Saturated liquid density
560.6
671.5
1373.41309.91250.51376.6(-25°C) [kg/m3]
Saturation density (60°C)
2.117
2.616
1.682
1.277
1.642
0.391
[MPa.abs]*2Theoretical COP [-]
2.68
2.73
2.69
2.67
2.60
2.77
*1: When R1233zd (E) is used at 0.2MPaG or less, the High Pressure Gas Safety Act does not apply.
*2: Two-stage expansion cycle, ET: -25°C, CT: 37°C, Eff.: 80%

2.2

Refrigeration cycle

In the case of equipment in the industrial field, in comparison with that in the air
conditioning field, there is an issue that the difference in fluid temperature between the evaporatorside and the condenser side is significant and the pressure ratio (the ratio of the condenser pressure
to the evaporator pressure) is high, thereby reducing the efficiency of the compressor and limiting
the operation range. Figure 1 shows typical fluid temperatures on the evaporator side and thecondenser side in the air conditioning field and the industrial field, as well as the pressure ratio
thereof. In the figure, the upper end of the bar indicates the fluid temperature on the condenser sideand the lower end the evaporator side.

Figure 1 Fluid temperature range and pressure ratio in air conditioning and industrial fields* The pressure ratio was converted using the R1234ze(E) saturation pressure at the fluid temperature on thecondenser and evaporator sides.

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 4 (December 2019)
3To deal with a high pressure ratio, we developed a multistage compression cycle in which
two compressors are connected in series, based on a two-stage compression compressor structurethat had been proven in centrifugal chillers in the air conditioning field. In the case of a multistagecompression cycle, the air volume of the low-pressure side compressor is higher due to its larger
suction gas specific volume compared with the high-pressure side compressor, so we adopted alarge-diameter impeller, considering the operable range. We examined the compressor’s operability
in a high-efficiency area within the actual operating temperature and capacity ranges and its
controllability to avoid rotating stalls and choke areas. In addition, we adopted a 4-stagecompression 1-stage expansion economizer cycle to downsize the unit and the refrigerant charge
amount. Figure 2 is a schematic diagram of the refrigeration cycle.

Figure 2 Referigeration cycle

2.3

External view and specifications of developed chiller

The low-pressure side compressor was installed on the evaporator and the high-pressure sideon the condenser so that large-diameter gas piping takes the shortest paths with a smaller pressure
loss. The economizer and other containers were placed on the condenser side as much as possible
to insulate them from the low-temperature evaporator and minimize the cold insulation area. The
adopted R1234ze(E) falls under a specified inert gas under the refrigeration safety regulations of
the High Pressure Gas Safety Act, so we installed a designated refrigerant detector in the chillerbody.

Figure 3External view (taken at time of factory trial operation)

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 4 (December 2019)
4The developed machine was based on centrifugal compressors and heat exchangers ofcentrifugal chillers that had been proven in the air conditioning field. In addition, to deal withissues in the temperature range below -20°C including the low-temperature brittleness of metalmaterials and the hardening of resin materials, ensuring compatibility between the lubricating oil
and the refrigerant and the increase in the specific volume of the refrigerant gas, we carefully
carried out material selection, structural design and control design through elastomer tests andevaluation tests.
Figure 3 and Table 2 show the external view of the developed machine (taken at the time offactory trial operation) and the main specifications, respectively.
Table 2Item
Refrigerant
Refrigeration oilUnit-

Refrigeration cycle-

External dimensions(L x W x H)
Rated capacity

USRtm

Main specifications

Specification
R1234ze(E)POE
4-stage compression 1-stage expansion economizer cycle(Figure 2)
7.7×4.2×3.4

768.5-16 (inlet)/-21 (outlet), common for specification 1 andBrine temperature°C
specification 2
Cooling waterSpecification 1: 32 (inlet)/37 (outlet)°C
temperature*1
Specification 2: 18 (inlet)/23 (outlet)
Compressor
Centrifugal type, fixed speed drive, 2 units
Heat exchanger
Shell and tube type (evaporator and condenser)
Economizer
Plate heat exchanger, 3 units*1: Cooling water temperature varies depending on the outside air temperature.

2.4

Chiller control

The developed machine uses a multistage compression cycle in which two compressors areconnected in series, and it has more control mechanisms than the centrifugal chillers proven in theair conditioning field. To secure the control stability during transients, we optimized the control
parameters using dynamic characteristic analysis.
Figure 4 and Figure 5 give the dynamic characteristic analysis results (typical example)
from startup with the brine inlet temperature of -16°C and the cooling water inlet temperature of32°C to the point where the brine outlet temperature decreased to -21°C. We confirmed that thelow-pressure side compressor started at the timing of 90 seconds shown in Figure 4, the first-stage
inlet vane (IGV) openings of the low-pressure side compressor and the high-pressure sidecompressor increased and the brine outlet temperature decreased to the target value of -21°C. Theair volume and the pressure ratio of both compressors increased after the start of the compressors
and did not enter unstable areas such as rotating stalls.

Figure 4 Dynamic characteristic analysis result 1 (control valve opening, temperature)Mitsubishi Heavy Industries Technical Review Vol. 56 No. 4 (December 2019)
5Figure 5 Dynamic characteristic analysis result 2 (flow rate coefficient, pressure coefficient)

The compressor air volume (flow rate coefficient) in Figure 5 was calculated using equation(1), and the pressure ratio (pressure coefficient) equation (2). Qst denotes the compressor airvolume, D the impeller diameter, u the impeller peripheral speed and Had the heat insulation head.
φ=π

Qst

4

D 2u

gH
μad = 2adu

Equation (1)Equation (2)|3. Evaluation of operation dataFigure 6 presents the operation data of the developed machine that started operation as heatsource equipment for low-temperature processes, from startup (10:10) to shut down (18:30) on acertain day. Figure 7 plots the data of COP with respect to the load factor for several cooling water
inlet temperature ranges in winter.

Figure 6

Operation data from startup to shut down

The developed machine was operated as heat source equipment for processes according to
the facility load. The load factor (cooling capacity) was stable at around 80% of the rated capacity
and the brine outlet temperature was constant at the target value (-21°C). We confirmed that thefirst stage inlet vane (IGV) and the main expansion valve followed the cooling capacity, brine
temperature and cooling water temperature, and that the developed machine could work withoutcarry-over to the compressor and excessive fluctuations in the evaporator pressure.
The heat exchanger performance of the developed machine and the pressure loss at each partsatisfied the planned values and we achieved the planned COP (Table 2, Specifications 2). We
confirmed that when the cooling water inlet temperature decreased, the COP increased, and thateven when the cooling water inlet temperature decreased to 12°C, the developed machine could
maintain the brine outlet temperature stably at the target value (-21°C).
The COP is calculated by equation (3). Where Qbrine denotes the cooling capacity (heatexchange amount of the evaporator) and Wchiller elec-in the chiller power consumption (sum ofhigh-pressure side compressor power consumption Whigh pressure side comp elec-in and low-pressure sidecompressor power consumption Wlow pressure side comp elec-in).

Mitsubishi Heavy Industries Technical Review Vol. 56 No. 4 (December 2019)6

COP = Qbrine / Wchiller elec-in

Equation (3)Wchiller elec-in = Whigh pressure side comp elec-in + Wlow pressure side comp elec-in

Equation (4)Figure 7

Load factor and COP in winter*1: The brine outlet temperature is constant at -21°C.
*2: The vertical axis represents the ratio between the measured COP and the rated COP under specification
condition 2.

|4. Conclusion
We adopted R1234ze(E) with a GWP of 1 or less and the multi-stage compression
refrigeration cycle and established the technology for heat source equipment for processes in theindustrial field. The developed machine, which started operation as heat source equipment for
low-temperature processes, maintained the brine outlet temperature stably at the target value(-21°C) according to the equipment load and achieved the planned COP. In addition, wedemonstrated the COP increased when the cooling water inlet temperature decreased, and that thechiller control followed the cooling water inlet temperature stably down to 12°C.
This development enables the expansion of the application range of low-GWP refrigerants in
the industrial field. Going forward, we will continue to make efforts to promote low-GWP
refrigerants and develop chillers continuously to satisfy our customers.

References
(1)
(2)
(3)

Intergovernmental Panel on Climate Change (IPCC), Working Group I Contribution To The Ipcc FifthAssessment Report Climate Change 2013, 2015.
ASHRAE, 2015 Addenda Supplement to ANSI/ASHRAE Standard 34-2013, Designation and Safety
Classification of Refrigerants, 2015.
NIST REFPROP Ver.9.1.1