Investigation of Heat Recovery in Different
Refrigeration System Solutions in
Supermarkets
JIGME NIDUP
Master of Science Thesis
Stockholm, Sweden 2009
Investigation of Heat Recovery in Different
Refrigeration System Solutions in
Supermarkets
Jigme Nidup
Master of Science Thesis Energy Technology 2009:499
KTH School of Industrial Engineering and Management
Division of Applied Thermodynamics and Refrigeration
SE-100 44 STOCKHOLM
Master of Science Thesis EGI 2009/ETT:499
Investigation of Heat Recovery in Different
Refrigeration System Solutions in Supermarkets
Jigme Nidup
Approved
Examiner
Supervisor
Björn Palm
Samer Sawalha
Commissioner
Contact person
Samer Sawalha
Abstract
This is a study of heat recovery in different refrigeration system in supermarkets. It
presents analysis of field measurement in two supermarkets; CO2 trans-critical system
with heat pump and CO2 trans-critical system with heat recovery only in desuperheater. The results from the field measurements showed that during the heat
recovery the minimum condensing temperature was 20oC.
Second part present computer models for heat recovery in three refrigeration systems
solutions; the CO2 trans-critical system with heat pump, CO2 trans-critical with heat
recovery in de-superheater only and a CO2/R404A cascade system.
As the preliminary study in the project the methods and procedures have been
developed for investigation of heat recoveries in three different refrigeration systems.
The parameters that may be possibly used to compare different system have been
proposed which need further refinement before drawing conclusion.
Acknowledgements
I would like thank the following members who formed the team working in the
related project: Per-Olof NIlson at IUC, Katrineholm, Miceal Antonsson at Green
and Cool, Magnus Rehnby at Bengt Dahlgren AB and Per-Erik Jansson at ICA
Sverige AB for rendering all the support. I heartily thank Sarah Johansson and
Jörgen Rogstam for allowing me to use the methods developed in their ongoing
project in “Evaluation of CO2 supermarket refrigeration systems”; David
Freléchox for all the support and suggestions.
This project would not have seen the light without patient support of my
supervisor Samer Sawalha.
Thank you!
Jigme Nidup
Stockholm, 2009
I
Table of contents
ACKNOWLEDGEMENTS........................................................................................................................ I
TABLE OF CONTENTS ..........................................................................................................................II
NOMENCLATURE................................................................................................................................. III
INDEX OF FIGURES ............................................................................................................................. IV
INDEX OF TABLES ............................................................................................................................... VI
1
INTRODUCTION ........................................................................................................................... 1
1.1
ENERGY USE IN SUPERMARKETS AND TYPICAL HEATING REQUIREMENTS .................................. 1
1.2
OBJECTIVES............................................................................................................................................ 3
2
METHODOLOGY ........................................................................................................................... 3
3
HEAT RECOVERY SYSTEM SOLUTIONS ................................................................................ 4
3.1
HEAT RECOVERY SYSTEM DESIGNS ...................................................................................................... 4
3.1.1 Heat Recovery System Design 1....................................................................................................5
3.1.2 Heat Recovery System Design 2....................................................................................................5
3.1.3 Heat Recovery System Design 3....................................................................................................6
3.1.4 Heat Recovery and Floating Condensing .................................................................................7
3.1.5 Heat Pump and Floating Condensing ........................................................................................7
4 FIELD MEASUREMENT............................................................................................................... 9
4.1
CO2 TRANS‐CRITICAL SYSTEM 1 ........................................................................................................ 9
4.1.1 System design and operation.........................................................................................................9
4.1.2 Calculations of heat recovery .....................................................................................................10
4.1.3 Existing mode of operation and selection of operating conditions for heat recovery with heat pump ..............................................................................................................................11
4.1.4 Recoverable Heat with heat pump solution ........................................................................14
4.2
CO2 TRANS‐CRITICAL SYSTEM 2 ......................................................................................................16
4.2.1 System design and operation......................................................................................................16
4.2.2 Calculation of Heat recovery ......................................................................................................17
4.2.3 Total heat recovery .........................................................................................................................19
4.2.4 Performance of Medium temperature Unit (KA3)............................................................21
4.2.5 Performance of booster unit (KAFA1)....................................................................................23
5
SIMULATION MODELS .............................................................................................................26
5.1
ASSUMPTIONS AND CORRELATIONS IN THE MODELS ...................................................................27
5.1.1 Cooling capacity as a function of outdoor temperature................................................27
5.1.2 Heating period...................................................................................................................................28
5.1.3 Other refrigeration and heat exchanger parameters .....................................................29
5.2
HEAT RECOVERY WITH HEAT PUMP ................................................................................................30
5.2.3 Evaluation of different modes of operation .........................................................................31
5.3
HEAT RECOVERY IN DE‐SUPERHEATER ONLY (TC2)....................................................................40
5.4
HEAT RECOVERY IN R404A/CO2 CASCADE SYSTEM ......................................................................46
5.5
COMPARISON OF DIFFERENT HEAT RECOVERY SYSTEM SOLUTIONS ..........................................50
6
DISCUSSION.................................................................................................................................52
7
CONCLUSION ...............................................................................................................................53
8
REFERENCES ...............................................................................................................................55
II
Nomenclature
CC
Cascade system
CFC
Chlorofluorocarbon
COP
Coefficient of Performance (refrigeration)
E
Compressor power consumption (electrical)
EES
Engineering Equation Solver
FA
Low temperature unit
HCFC
Hydrochlorofluorcarbons
HFC
Hydrofluorocarbons
HC
Hydrocarbons
HVAC
Heating, ventilation and air conditioning system
IHE
Internal heat exchanger
KA
medium temperature unit
PR
Pressure ratio
TC1
Trans-critical system 1
TC2
Trans-critical system 2
III
Index of figures
Figure 1.1: Typical energy use in a supermarket in Sweden (Arias J, 2005)...... 2
Figure 3.1: Heat recovery system design 1(Arias J, 2005) ................................. 5
Figure 3.2: Heat recovery system design 2 (Arias J, 2005) ................................ 6
Figure 3.3: Heat recovery system design 3 (Arias J, 2005) ................................ 6
Figure 3.4: Heat recovery system design 3 (Arias J, 2005) ................................ 7
Figure 3.5: Heat pump with floating condensing (Minea V, 2007) ...................... 8
Figure 4.1: Schematic diagram of trans-critical system 1 ................................... 9
Figure 4.2: Monthly average of cooling capacity, compressor power, and heat
rejected based on the measurements of trans-critical system 1(TC1).............. 12
Figure 4.3: Existing mode of operation of medium temperature unit (KA1) and
selection of operating condition for heat recovery in TC1................................. 13
Figure 4.4: Existing operating conditions of low temperature unit (FA1) and
selection of operating conditions for heat recovery in TC1............................... 14
Figure 4.5: Estimated heat recovery potential with heat pump system............. 15
Figure 4.6: Schematic diagram of trans-critical system 2 (TC2) ....................... 16
Figure 4.7: schematic diagram of de-superheater showing the assumed
parameter in the heat recovery calculation in TC2. .......................................... 19
Figure 4.8: Total cooling capacity, heat recovered and borehole capacity in TC2
......................................................................................................................... 21
Figure 4.9: Average monthly performance of medium temperature unit (KA3) in
TC2 .................................................................................................................. 22
Figure 4.10: Operation of medium temperature unit over a period of 24 hours in
TC2 .................................................................................................................. 23
Figure 4.11: Average monthly performance of (booster system) (KAFA1) in TC2
......................................................................................................................... 24
Figure 4.12: Operation of booster unit (KAFA1) over a period of 24 hours. ..... 25
Figure 5.1: The profile of cooling capacity as function of outdoor temperature 28
Figure 5.2: Schematic of heat recovery with heat pump system solution (TC1).
......................................................................................................................... 30
Figure 5.3: COP in floating condensing and heat pump modes of operation for
TC1. ................................................................................................................. 31
Figure 5.4: Operation in the heat pump mode with heat pump supply
temperature of 45 o C compared to the floating condensing mode for TC1 ...... 32
Figure 5.5: Operation of refrigeration system in air cooled condenser and
brine (secondary fluid) cooled condenser for TC1............................................ 33
Figure 5.7: System solution with heat pump supply temperature of 35oC in TC1.
......................................................................................................................... 34
Figure 5.8: Operating at conditions to supply temperature of coolant from
condenser at 45oC in TC1. ............................................................................... 35
Figure 5.9: Operating at conditions to supply temperature of coolant from
condenser at 35oC in TC1. ............................................................................... 36
Figure 5.10: Comparison of air cooled mode and heat recovery mode with
supply temperature of 35oC in TC1 .................................................................. 37
Figure 5.11: Ratio of heat recovery/rejection to cooling capacity in different
modes of operation for TC1.............................................................................. 38
Figure 5.12: Total power consumption in different mode of operation for TC1. 39
IV
Figure 5.13: Schematic of refrigeration system with heat recovery in desuperheater region (TC2). ................................................................................ 40
Figure 5.14: Coefficient of performance of system in different modes of
operation in TC2............................................................................................... 41
Figure 5.15: heat recovery at supply temperature to heat recovery system of
35oC in TC2...................................................................................................... 42
Figure 5.16: Heat recovery at supply temperature to heat recovery system of
45oC in TC2...................................................................................................... 43
Figure 5.17: Compressor power for different capacity and temperature of heat
recovery in TC2 ................................................................................................ 44
Figure 5.18: Performance of TC2 with heat recovery of 70% of the total heat
rejected at de-superheater region at supply temperature of 35oC .................... 45
Figure 5.19: Schematic of heat recovery system solution in R404A/CO2
cascade system (CC) ....................................................................................... 46
Figure 5.20: Heat recovery with supply temperature of 35oC in cascade system
(CC).................................................................................................................. 47
Figure 5.21: Heat recovery with supply temperature of 45oC in cascade system
(CC).................................................................................................................. 48
Figure 5.22: Comparative performance of heat recovery with 35oC and aircooled condenser in floating condensing mode in cascade system (CC)......... 49
Figure 5.23: heat recovery ratio for different heat recovery temperature levels
and floating condensing mode in CC................................................................ 49
Figure 5.24: Power consumption in different modes of operation in cascade
system (CC) ..................................................................................................... 50
Figure 5.25: COP of TC1, TC2 and CC in floating condensing mode. ............. 51
Figure 5.26:
Heat recovery ratio for different heat recovery system solutions
for at supply temperature of 35oC..................................................................... 52
V
Index of tables
Table 1.1: Temperature requirements in heating applications (BBR, 2008;
Wulfinghoff D R, 1999) ....................................................................................... 2
Table 4.1: Manufacturer’s data of compressors in TC1.................................... 10
Table 4.2: Monthly average values of heat recovery with the supply temperature
of brine at 35oC in TC2 ..................................................................................... 20
Table 5.1: Average and maximum compressor power (Adapted from Jaime
Arias, 2005) ...................................................................................................... 27
Table 5.2: Heating season and average monthly temperature. (Jonsson. H, et
al, 2005, SMHI, 2009 and weather base, 2009) ............................................... 29
Table 5.3: Parameters assumed in the models ................................................ 29
VI
1
Introduction
Mankind started with the use of natural ice to cool the food and then learned to
produce the ice mechanically thus making cooling independent of nature. Later
advances in the field of cooling made it possible to directly cool the food with
out the use of ice as intermediate medium. Since these mechanical refrigerators
involved the use of working fluid, development was governed by the choice of
refrigerants.
Until 1922 ammonia, carbon dioxide and sulfur dioxide were used as
refrigerants. Though these refrigerants were naturally occurring safety concerns
limited the widespread application. The introduction of new refrigerants in 1930
was the beginning of use of Chlorofluorocarbon (CFC) and hydro
chlorofluorocarbon (HCFC) (Granryd E et al., 2003). Though these refrigerants
were safe, CFCs were found to deplete the ozone layer and have high global
warming potential. Then the concern shifted to the long term effect of the
refrigeration to the environment. In Sweden, the use of CFCs was banned in
2000 and refilling of HCFCs was banned in 2002. Though HFC refrigerants
accepted as the replacement of CFCs and HCFCs, owing to its very high global
warming potential, it is accepted only as temporary solution. Regulations on
synthetic refrigerants brought in many changes in refrigeration system. Systems
became tighter, contained less volume of charge. In the process of redesigning
the refrigeration system, different solutions have evolved in the supermarket
refrigeration systems. With the focus on reduction of charge, the conventional
direct expansion systems have been replaced with indirect systems.
Supermarket refrigeration systems today may be partially indirect to completely
indirect systems.
Side by side the energy utilization in general and supermarkets in particular
have been closely scrutinized. Many of the supermarket systems in Sweden
use heat recovery system to recover heat from the refrigeration system to
reduce the net energy purchased. Re-emerging use of natural refrigerants like
propane, ammonia and carbon dioxide makes it interesting possibility to recover
heat at higher temperature. Different system solutions have been implemented
in different supermarkets in Sweden. There is a need for a comparative study of
these system solutions.
1.1
Energy use in supermarkets and typical heating requirements
In Sweden, supermarkets accounts for about 3% of the total electricity
consumed in the country. It also revealed that reduction in energy consumption
by 50% could increase the profit by 15%. (Arias J, 2005). Therefore detailed
investigation of energy utilization in a supermarket is worthwhile. Figure 1-1
shows the energy use in a typical supermarket in Sweden, as can be seen in
the figure, considerable portion of energy is consumed in refrigeration, lighting
and HVAC. While refrigeration system rejects considerable amount of heat,
separate heating is required in the HVAC system. The recovery of heat from the
refrigeration system presents potential to reduce or replace heating need for
HVAC and service water.
Energy usage in a supermarket
Others, 5%
Outdoors, 5%
Kitchen, 3%
HVAC, 13%
Refrigeration,
47%
Lighting, 27%
Figure 1.1: Typical energy use in a supermarket in Sweden (Arias J, 2005)
Common heating applications in a supermarket consist of floor heating, heater
for HVAC system, service water heater and in modern application, defrosting of
evaporator coils. Some heating applications and temperature requirements are
listed in table 1-1.
Table 1.1: Temperature requirements in heating applications (BBR, 2008;
Wulfinghoff D R, 1999)
Applications
Temperature level
(oC)
Floor heating
16-27
HVAC dimensioning temp
33
Service hot water
54
Hydronic system
71-49
Alternately with recovery system hydronic system can 54-32
use temperature level
2
1.2
Objectives
Main objectives of the thesis are as follows:
•
•
•
2
Modelling of different system solutions for heat recovery in supermarket
refrigeration system.
Collect and analyze field measurement data.
Evaluation of the performance of heat recovery systems in supermarkets
using the field measurements and the computer simulation models.
Methodology
The literature review was carried out to study the different refrigeration and heat
recovery system solutions in the literature.
Refrigeration systems in two supermarkets with CO2 trans-critical systems have
been analysed for heat recovery. The data from these supermarkets have been
collected through the online interface IWMAC (IWMAC 2009). The electric power
consumption of the compressor, pressure and temperature at key points are
recorded for every 5 minutes interval. The important performance indicators are
thereby calculated. Microsoft excel have been used for data analysis and NIST
reference properties have been used to calculate the properties of refrigerants
through the REFPROP 7.0 (NIST, 2002).
Analyses of data from the field measurement have been used as one of the
basis to make assumptions and develop the different computer simulation
models. Some of the parameters have been either taken from literature or
derived using the figures in the literature.
3
3
Heat Recovery System Solutions
Theoretically, in refrigeration system heat can be recovered from compressor
oil, de-superheater and condenser. Potential of heat recovery depends on the
quality and quantity of heat required and experience varies on the percentage
of high temperature that can be recovered. In an industrial ammonia vapour
compression system only 11.5% of the total system heat rejection is available in
superheat region and the remaining is rejected at lower quality (Reindl T D, et
al, 2007). A study of supermarket in the United States showed that heat
rejection in superheat could range from 14-20% of the total heat available
(Walker D H, 2001). Visile Minea suggested that up to 30% heat can be
recovered at the superheat region (Minea V, 2007). Therefore the percentage of
heat can be recovered depends on the refrigerants use the system solution..
Heat recovery system solutions in supermarket are used mainly to heat the
space air. The practical experience indicated that though seemingly high
quantity of heat is rejected by supermarket refrigeration systems, only 40-70%
of the necessary heat can be recovered (Arias J, 2005). Arias suggested that
refrigeration system not operating continuously to be the likely reason for less
recovery of heat. In a typical Swedish supermarket HVAC system and
refrigeration system are installed and operated by different companies.
Therefore HVAC and refrigeration systems are isolated from each other with an
heat exchanger has been cited as another reason.
To get the required temperature of the water at the using end of the real estate
of HVAC system, ideal temperature of coolant leaving condenser of the
refrigeration system is 38oC (Arias J, 2005). This increases the condensing
temperature, which increases the power consumption of the compressor. Other
concerns are the effect on the cooling capacity, matching the timing of cooling
and the heat recovery, matching temperature of heat recovery and that of
application, loss in heat exchangers, and risk of leakage and contamination in
the heat exchangers.
The fall in the outdoor temperature decreases the indoor relative humidity. This
reduces the cooling load and therefore the heat rejected by the refrigeration
system. The effect of outdoor temperature could reduce the refrigeration load
by 50-70% of refrigeration load at standard indoor conditions of 22oC and 65%
relative humidity (Arias J, 2005).
3.1
Heat recovery system designs
Different heat recovery systems have been implemented in the supermarkets in
Sweden. The following designs deal with the heat recovery at the condenser
rather than de-superheater and reject the recovered heat to the HVAC directly
or through a heat exchanger. The system solutions discussed are adapted from
the doctoral thesis by Jaime Arias (Arias J, 2005).
4
3.1.1 Heat Recovery System Design 1
Fig 3-1 shows the layout of heat recovery design 1. The coolant extracting heat
from condenser of the refrigeration system rejects heat to the HVAC system
through a heat exchanger before entering the dry cooler. The additional heat
exchanger between the air system and refrigeration coolant loop reduces the
temperature efficiency. The approach temperature of 32OC and return of 28 OC
is obtained on the HVAC side of the heat exchanger. To maintain these
temperature level, requirements of water temperature on the refrigeration (heat
rejecting side) is around 36 OC approach to the dry cooler and 32 OC return.
To ensure operating at the required temperature of the HVAC heating system,
an auxiliary heating system is used to heat the water. This requires
complicated control system, when the heat exchanger in the air system has
reached the required temperature; heat from the auxiliary system is bypassed
and rejected through the dry cooler.
The introduction of additional heat exchanger between the HVAC system and
chillers is to enable independent operation of chillers in events of break down of
the HVAC system. In Sweden, the installation and operation of space heating
system and refrigeration are often executed by different parties.
Figure 3.1: Heat recovery system design 1(Arias J, 2005)
3.1.2 Heat Recovery System Design 2
Figure 3.2 shows the arrangement of heat recovery design 2. In this design the
auxiliary heating in air system is separated from the heat recovery heat
exchanger. The drawback of heat loss from auxiliary heating to ambient is
5
eliminated in this design, the reduction in approach temperature introduced by
the additional heat exchanger is still present.
Figure 3.2: Heat recovery system design 2 (Arias J, 2005)
3.1.3 Heat Recovery System Design 3
Figure 3-3 shows the third design. This design eliminates the inadequacies in
previous designs related to the heat recovery exchanger. Since the coolant
from condenser of the refrigeration system directly rejects heat to the HVAC
heat exchanger, the system is operated with approach temperature of 36oC and
return temperature of 32oC. Auxiliary heating is placed after the heat exchanger
in the HVAC system.
Figure 3.3: Heat recovery system design 3 (Arias J, 2005)
6
3.1.4 Heat Recovery and Floating Condensing
Inherent disadvantage of heat recovery system is the need to refrigeration
system at the high condensing pressure even in winter to maintain the required
temperature of the heat utilizing system. Had the refrigeration system not been
connected to heat recovery system, in cold days, condensing pressure could be
lowered taking advantage of low ambient temperature. This could reduce the
power consumption of the compressors in the refrigeration system. But since
the temperature level of the rejected heat in this case is not of the level that can
be used by the heat utilizing systems such as HVAC, the heat from the
condenser will have to be rejected to the ambient.
System in Figure 3.4 incorporates both the advantages. Such a system enables
some chillers to run at high condensing temperatures to meet the heating
demand and rest of the chillers can operate at lower condensing temperature to
take advantage of low ambient temperature. In this system design one of the
chillers is operated at high condensing temperature so as to meet the heating
demand and the second chiller is operated at lower condensing pressure
corresponding to the outdoor temperature.
Figure 3.4: Heat recovery system design 3 (Arias J, 2005)
3.1.5 Heat Pump and Floating Condensing
Another concept of heat recovery from refrigeration system is the use of heat
pump to extract heat from the condenser coolant at low temperature and
transfer it to the HVAC system at high temperature. This system enables the
use of rejected heat at the same time allows the chillers to operate at floating
condensing mode, therefore minimize the penalty on the compressor power.
Figure 3.5 shows the schematic diagram of heat pump and floating condensing
design. Experience with HCFC-22 at a supermarket in Canada showed that
7
even with the floating condensing, temperature of brine entering the heat pump
was about 25oC and COP of heat pump was as high as 4.6 (Minea V, 2006).
One drawback of this design cited in the study was impossibility of operating the
heat pump in cooling mode during the summer. On hot summer days the brines
temperatures from the condenser of the chillers could rise up to 40oC which did
not permit the operation of the heat pumps in reverse mode.
Figure 3.5: Heat pump with floating condensing (Minea V, 2007)
Another system design for supermarkets is to recover heat only in the desuperheater region. A field measurement of such a system is discussed in
section 4.2.
8
4
Field Measurement
4.1
CO2 trans-critical system 1
The measurement is carried out on a supermarket located in the far north of
Sweden. The refrigeration system is CO2 trans-critical system and heat recovery
is with a heat pump. The system solution is designated as trans-critical system
1 (TC1)
-
+
4.1.1 System design and operation
Q_max =300kW
T_hp in= 13o C
Heat pump COP= 3.6
Oil cooler
Inter-Cooler
Oil cooler
- Pressure
KA1/KA2
-Temperature
FA1/FA2
Figure 4.1: Schematic diagram of trans-critical system 1
The refrigeration system is a parallel system with two circuits for low
temperature level and two circuits for medium temperature level. The Low
temperature units designated as FA1 and FA2 have two two-stage compressors
each. Medium temperature units KA1 and KA2 have four single-stage
compressors each. The manufacturer data of compressors are given in table
4.1.
9
Table 4.1: Manufacturer’s data of compressors in TC1
Units
model
Swept Volume
Low temperature
TCDH 372B-D
12.6m3/h @ 2900rpm
Medium temperature
TCS373-D
All four circuits reject heat to the common coolant loop which is cooled by dry
cooler. The compressor oil cooler also reject heat to the same coolant loop. A
heat exchanger is connected to the coolant loop before the dry cooler to
recover heat. This heat exchanger is connected to 300kW heat pump, which is
meant to deliver heat to the HVAC system of the supermarket. At present the
heat pump is not in operation, so all the heat is reject through the dry cooler.
4.1.2 Calculations of heat recovery
The supermarket is installed with measurements of temperature and pressure
on the refrigerant circuits. Temperatures are also measured on the coolant loop.
The measurement points are shown in figure 4.1. Power consumption of the
compressors in the four units is measured separately. Power consumption,
temperatures and pressures are measured for five minutes intervals and
recorded by online monitoring system which can be accessed through web
interface www.iwmac.se. The estimation of heat recovery is based on the
parameters on the refrigeration side. The method developed in an ongoing
project on evaluation of supermarket refrigeration systems have been adopted
for calculation of various parameters (Johansson S, 2009).
Using equation 4.1, mass flow of the refrigerant is calculated
.
m& CO2 =
η s * Vs
v2 k
[kg / s]
(4.1)
.
Where Vs is swept volume and is taken from the manufacture’s data; v2 k is the
specific volume and is obtained as the function of pressure and temperature of
refrigerant at the inlet of the compressor. Volumetric efficiency ( η s ) is calculated
as a function of suction and discharge pressures from the manufacture’s data.
The function is different for the single and the two-stage compressor. The
volumetric efficiency for the single stage compressor is obtained using equation
4.2.
2
⎞
⎛P
⎛ Pdisch arg e ⎞
⎟⎟ − 6.5843⎜⎜ disch arg e ⎟⎟ + 102.42[%]
(4.2)
η s = −0.4079⎜⎜
P
P
⎝ suction ⎠
⎝ suction ⎠
And the volumetric efficiency of two stage compressor is obtained using 4.3.
⎛P
η s = −0.0251⎜⎜ disch arg e
⎝ Psuction
2
⎛P
⎞
⎟⎟ − 1.1706⎜⎜ disch arg e
⎝ Psuction
⎠
10
⎞
⎟⎟ + 93.424[%]
⎠
(4.3)
The heat transfer across the evaporator, the condenser and the intercooler is
calculated as the product of mass flow rate of refrigerant and the enthalpy
difference across the heat exchanger using equation 4.4.
Q& = m& CO2 * ∆h[kW ]
(4.4)
Where, Q& is the heat in (kW) and ∆h is the difference in enthalpy across the
heat exchanger in (kJ/kg).
Heat loss in the oil cooler is calculated as the difference between the measured
electrical power of the compressor and the calculated shaft power using
equation 4.5. Heat loss from the compressor body to the ambient is assumed
as 7% of the measured electrical power input.
Q& oilcooler = ( E& comp − E& compshaft )[kW ]
(4.5)
Q& oilcooler is the heat loss in the oil cooler, and E& comp is the measured electrical
power of the compressor in kW. E& compshaft is the net power transmitted to the
compressor shaft and is calculated using the equation 4.6.
E& compshaft = m& CO2 * ∆h[kW ]
(4.6)
∆h is difference of enthalpy between compressor outlet and inlet.
Total heat rejected from the refrigeration system is the sum of the condensers’,
the inter coolers’ and the oil coolers’ capacities.
4.1.3 Existing mode of operation and selection of operating conditions for
heat recovery with heat pump
At present the heat recovery system is not in operation. Figure 4.2 shows the
plot of total cooling capacity, average condensing temperature (T_condensing),
common supply temperature of coolant brine (T_com,b,o) and total heat
rejected in the system. The top line is the curve of total heat rejected by the
system. This includes heat rejected from oil cooler and the inter cooler of the
two-stage compressor of the low temperature units. Oil cooler capacity was
found to be rather constant over the year at about 8kW and intercooler capacity
was found to be about 13kW.
11
25
250
20
200
15
150
T (C)
Q,E(KW)
10
5
100
0
50
-5
M
ar
ch
_0
9
Fe
b_
09
Ja
n_
09
D
ec
_0
8
N
ov
_0
8
O
ct
_0
8
S
ep
t_
08
Au
g_
08
Ju
ly
_0
8
Ju
ne
_0
8
M
ay
_0
8
Ap
ril
_0
8
M
ar
ch
_0
8
-10
Fe
b_
08
Ja
n_
08
0
Month
Cooling capacity
Compressor power
T_ambient
T_com,b,o
T_condensing
Figure 4.2: Monthly average of cooling capacity, compressor power, and
heat rejected based on the measurements of trans-critical system 1(TC1)
The common brine temperature of the heat rejection loop for the months of
January and February in 2008 was not recorded. The brine temperatures are
average for one month. In the existing mode of operation the brine temperature
is above 18oC only in July and August. The design inlet temperature of brine to
the heat pump is 13oC. This requires that the minimum supply temperature of
brine from the refrigeration system to be 18oC, this is assuming 5K temperature
difference across the heat exchanger. In order to operate the heat pump during
the heating season, the refrigeration system would have to operate at discharge
pressure corresponding to that of august (57.7 bar). This point of selection is
highlight in figure 4.2.
The selection criteria of operating conditions for individual units can be
understood from figures 4.3 and 4.4.
12
20
1.0
10
0.5
0
0.0
N
O
Ap
ril
_
COP
1.5
M
ar
ch
_0
9
30
Fe
b_
09
2.0
Ja
n_
09
40
D
ec
_0
8
2.5
ov
_0
8
50
ct
_0
8
3.0
Se
pt
_0
8
60
Au
g_
08
3.5
Ju
ly
_0
8
70
Ju
ne
_0
8
4.0
08
80
M
ay
_0
8
4.5
M
ar
ch
_0
8
90
Fe
b_
08
5.0
Ja
n_
08
Q(kW),T(C)
100
Month
Cooling capacity
Total heat rejected
T_cond
T_com,b,o
COP
Figure 4.3: Existing mode of operation of medium temperature unit (KA1)
and selection of operating condition for heat recovery in TC1
From figure 4.3, the operating conditions for medium temperature unit to run the
system in heat recovery mode is selected as pointed out by the ellipse in the
figure. Corresponding to common brine outlet temperature of 18oC, condensing
pressure is selected at 58 bars and COP of 3.2 in the heat recovery mode.
13
2.0
40
1.8
35
1.6
30
1.4
1.2
1.0
20
COP
Q(kW),T(C)
25
0.8
15
0.6
10
0.4
5
0.2
M
ar
ch
_0
9
Fe
b_
09
Ja
n_
09
D
ec
_0
8
ov
_0
8
N
ct
_0
8
O
Se
pt
_0
8
Au
g_
08
Ju
ly
_0
8
Ju
ne
_0
8
M
ay
_0
8
08
Ap
ril
_
M
ar
ch
_0
8
0.0
Fe
b_
08
Ja
n_
08
0
Month
Cooling capacity
Total heat rejected
T_com,b,o
T_cond
COP
Figure 4.4: Existing operating conditions of low temperature unit (FA1)
and selection of operating conditions for heat recovery in TC1
From the above figure, operating conditions for heat recovery mode is selected
corresponding to common brine outlet temperature of 18oC. Therefore low
temperature units need to be operated at condensing temperature of at least
20oC at COP of 1.7. These operating conditions have been assumed to
calculate the heat that can be recovered if operated in heat recovery mode.
4.1.4 Recoverable Heat with heat pump solution
From the above selections, the refrigeration system is set to operate at
condensing temperature of 20oC on the heat recovery mode. The COP of
medium and low temperature units are set to 3.2 and 1.7 respectively during the
heat recovery mode. Using these values, compressor power, oil cooler
capacity, and the total heat rejected is calculated. Total heat rejected is used
as the heat source for the heat pump. Using the design COP of the heat pump,
which is obtained from the heat pump manufacturing data (CIAT, 2009) total
heat that is provided by the heat pump, is calculated. This potential is compared
with the maximum capacity of the heat pump, thus heat potential is limited to
maximum capacity of the heat pump. Power consumption of the heat pump is
estimated by dividing the total heat supplied by the heat pump by its COP.
Total heat recovery as calculated is presented in figure 4.5. To compare the
heating potential and the maximum capacity of the heat pump, heat recovery
was considered even for the warm months of June, July and August.
Presented in the plot from bottom to top of the curve are power consumption of
the compressors of the refrigeration system while operating without any heat
14
recovery (E_ref_ref only mode), the power consumption of the compressors of
refrigeration system when operating on heat recovery mode (E_ref_HR mode),
total power consumption in heat recovery mode (E_tot_HR mode), Cooling
capacity of the refrigeration system, COP of heating and heating capacity of the
heat pump. The total power on heat pump mode (E_tot_HR mode) is sum of the
power consumed by the compressors of the heat pump and the refrigeration
system.
5
300
Maximum capacity of heat pump= 301kW
4
250
3
COP
Q,E(kW)
200
150
2
100
1
50
M
ar
ch
_0
9
Fe
b_
09
Ja
n_
09
D
ec
_0
8
N
ov
_0
8
ct
_0
8
O
Se
pt
_0
8
Au
g_
08
Ju
ly
_0
8
Ju
ne
_0
8
M
ay
_0
8
Ap
ril
_0
8
M
ar
ch
_0
8
0
Fe
b_
08
Ja
n_
08
0
Month
E_ref_ref only mode
E_ref_HR mode
E_tot_HR mode
cooling capacity
heat pump capacity
COP heating
Figure 4.5: Estimated heat recovery potential with heat pump system.
The flat curve for heat pump capacity from April to October is because of
reaching the maximum capacity of the heat pump. The upper limit of COP of
heating is limited by COP of the heat pump. It can also be observed that COP
of heating is lower during the heating season than on the warmer months. This
can be attributed to the increase in power consumption of compressors in the
refrigeration systems during the heating season. During the heating season, the
condensing pressure is maintained at higher pressure to maintain the brine
temperature leaving the condenser at a level necessary for operating for the
heat pump.
15
4.2
CO2 trans-critical system 2
This supermarket is located near the city of Goteborg, which is at the western
coast of Sweden.
4.2.1 System design and operation
+
+
Floor heating
Desuperheater
Desuperheater
Oil cooler
KA3
-Pressure
-Temperature
KAFA1/KAFA2
Ground heat
source
Figure 4.6: Schematic diagram of trans-critical system 2 (TC2)
This is a CO2 trans-critical system with two booster systems for the low and
medium temperature levels (KAFA1&KAFA2) and a medium temperature circuit
(KA3). Compressor in the medium temperature circuit, and high stage of
Booster system, is TCS373-D model from Dorin with swept volume of 12.6m3/h
at 2900rpm. The booster compressor in the booster system is SCS362-D model
from Dorin with swept volume of 10.7m3/h at 2900rpm.
On the heat rejection side, all three units are connected with a de-superheater
before the air cooled condenser. The de-superheaters are connected to a
common brine loop which is used to recover the heat. The recovered heat is
used for floor heating and HVAC system of the building. The compressors’ oil in
this system is cooled by a separate air cooler.
The heat recovery system operates so as to maintain the supply temperature of
the brine to the heating system at about 40oC. In the field measurement the
average temperature were about 35oC. So the system fails to maintain the
temperature at 40oC, instead we get 35oC in average. The capacity of individual
unit is maintained by opening or closing the electronic valve connected to the
16
refrigerant line after the condenser. This raises or decreases the condensing
pressure thereby controlling the capacity at fixed temperature. The supply
temperature of the brine is also controlled by controlling the flow rate of brine to
the de-superheater. Supply of brine to each de-superheater is controlled with
the variable speed pump. The refrigerant line is also externally sub-cooled from
the borehole before the supplying to the cabinets.
4.2.2 Calculation of Heat recovery
The heat recovery calculations are based on the refrigerant parameters and
calculated using the same method developed in the project on evaluation of
refrigeration systems in supermarkets (Johansson S, 2009).
Using equation 4.1, the mass flow of the refrigerant in medium temperature
level circuit is calculated. Volumetric efficiency (η s ) is calculated as a function of
suction and discharge temperature from the manufacture’s data. The volumetric
efficiency for the medium temperature compressor and high stage compressor
in the booster system is obtained using equation 4.2. The heat transfer across
the evaporator and de-superheater is calculated as the product of mass flow
rate of refrigerant and the enthalpy difference across the heat exchanger using
equation 4.4. Heat loss in the oil cooler is calculated as difference of measured
electrical power of compressor and the calculated shaft power of the
compressor using equation 4.5.
For the booster systems, the mass flow of refrigerant in the medium and low
temperature level need to be evaluated separately to calculate the respective
cooling capacities.
m& CO2total = m& CO 2 medium + m& CO2low [kg / s ]
(4. 7)
m& CO2total is the total mass flow of refrigerant; m& CO2medium is the mass flow of
refrigerant in the medium temperature level, and m& CO2low is the mass flow in the
low temperature level.
The volumetric efficiency of the booster compressor is calculated using
equation 4.8.
⎛ Pdisch arg e
η s = −0.1139⎜⎜
⎝ Psuction
2
⎞
⎛P
⎟⎟ − 4.1854⎜⎜ disch arg e
⎠
⎝ Plow
⎞
⎟⎟ + 95.12[%]
⎠
(4.8)
Since both the low and medium temperature levels depend on the high stage
compressor for cooling, it is necessary to divide up the compressor power in
order to evaluate performance of the booster system. Total cooling capacity is
divided as the following:
Q& high = Q& medium + Q& low + E& booster,shaft [kW ]
(4.9)
17
Q& medium is the cooling capacity of medium temperature; Q& low is the cooling load of
the low temperature level, and E& booster ,shaft is the net power transferred to the
refrigerant at the low stage and is calculated as
E& boostershaft = m& CO2low * ∆h[kW ]
(4.10)
∆h is the enthalpy difference across the inlet and outlet of the booster
compressor.
Then performances are calculated using following equations:
⎛ Q& high ⎞
⎟
(4.11)
COPhigh = ⎜
⎜ E& high ⎟
⎝
⎠
COPhigh is the COP calculated across the high stage compressor . Q& high
includes the cooling load from both low temperature and the medium
temperature sides. Therefore compressor power ( E& high ) has to be divided into
low temperature side and that of the medium temperature side.
⎛ Q&
⎞
E& high ,medium = ⎜ medium ⎟[kW ]
⎜ COP ⎟
high ⎠
⎝
(4.12)
E& high,medium is the portion of power consumed in order to provide the cooling load
of the medium temperature level.
COPtot _ booster =
Q& medium + Q& low
E& high + E& booster
(4.13)
COPtot _ booster , is total COP of the booster circuit.
COPbooster =
Q& low
E& high ,low + E& booster
(4.14)
COPbooster is calculated using the power consumed to provide cooling capacity at
the low temperature level of the booster system. E& booster is the power of booster
is the portion of power of the high stage compressor which
compressor. E&
high ,low
is consumed to provide cooling for the low temperature level. It is calculated
using equation 4.15
E& high,low = E& high − E& high,medium[kW ]
(4.15)
18
To calculate the heat recovery capacity on the refrigerant side the
measurement of temperature and pressure before and after the de-superheater
is necessary. While measurements of pressure and temperature for key points
on the refrigerant line were available from September 2008, the measurement
of temperature at the exit of de-superheater was available only from march,
2009. The temperature measurement of brine in and out of the de-superheater
was available from October 2008. The average difference between the
temperature of hot gas out of de-superheater and that of the temperature of
brine at the inlet of the de-superheater for month of March is used to back
calculate the hot gas temperature for the past months. The average
temperature difference between hot gas and the brine inlet for the heat
exchanger of the medium temperature unit KA3 was 4K, but for low
temperature unit KAFA1 was 1K and for KAFA2 was 0K. Since the approach
temperature of 1 and 0 is not reasonable, which could be accounted to
measurement error, both the values have been discarded and approach
temperature of 4K was used for all the de-superheaters. Therefore the
temperature difference of 4K is added to the common brine inlet temperature to
get the temperature of hot gas out at the de-superheater for the missing
measurements.
Heat recovery brine loop
T2 (measured)
Approach temperature
(assumed) = T1-T2 =4K
De-superheater
T1 (Calculated)
Refrigerant loop
Figure 4.7: schematic diagram of de-superheater showing the assumed
parameter in the heat recovery calculation in TC2.
4.2.3 Total heat recovery
Table 4.2 presents monthly average values of power consumption, cooling
capacity, COP of refrigeration, total heat rejected, heat recovered and subcooling with borehole. The heat recovered in de-superheater is the heat
recovered and the percentage of heat recovery is calculated as the ratio of heat
recovered in de-superheater to that of the total heat rejected by the refrigeration
system. The percentage of heat recovered through de-superheating varies
from 24 % to 35 %.
19
The sub-cooling with the borehole is expressed as the percentage of cooling
capacity to appreciate improvement due to the sub cooling with borehole. The
total heat rejected in the system is equal to sum of cooling capacity and
compressor power minus oil cooler capacity, borehole capacity and heat loss
from the compressor body. The heat loss from compressor body is assumed as
7% of the compressor power.
The heat balance for the calculations gives an error of about 4%. The error is
quite acceptable and could be accounted to errors in measurement.
Table 4.2: Monthly average values of heat recovery with the supply
temperature of brine at 35oC in TC2
Description
Oct-08
Nov-08
Dec-08
Jan-09
Feb-09
Mar-09
Apr-09
Average
outdoor
temperature (o C)
10
6
4
3
2
5
11
Compressor power (kW)
57
59
57
54
55
56
55
Cooling capacity(kW)
183
177
174
163
168
176
Total COP
3.34
3.10
3.04
3.00
3.00
3.02
3.27
Total heat rejected (kW)
Q_desuperheater (heat
recovered) (kW)
Q_borehole
(sub
cooler)(kW)
211
201
179
166
164
170
187
50
55
56
55
57
56
46
9
12
29
30
33
32
24
Q_oilcooler (kW)
% of Heat recovery =
(Q_desuperheater/total
heat rejected)
Q_borehole
as
percentage of cooling
capacity
11
11
12
11
12
11
10
24%
27%
31%
33%
35%
33%
25%
5%
7%
17%
19%
20%
19%
13%
165
The relationship between cooling capacity, heat recovery and condensing
temperature can be observed in figure 4.8. Since heat recovery is only in the
de-superheater, heat recovery is the de-superheater heat referred as
Q_desuperheater.
20
60
3.5
50
3.0
40
COP
Q, (kW), T(C)
2.5
30
2.0
20
1.5
10
0
1.0
Oct-08
Nov-08
Dec-08
Jan-09
Feb-09
Mar-09
Apr-09
Month
T_ambient
Q_desuperheater
Q_borehole
T_cond
COP
Figure 4.8: Total cooling capacity, heat recovered and borehole capacity
in TC2
The oil cooler capacity doesn’t change much over time. The evaporation
temperatures are found to be rather constant over time for low temperature unit
at -35oC and medium temperature at -10oC. It is seen that the heat recovery
capacity is increased during colder months of December to March compared to
warmer months of October, November and April. Therefore condensing
temperature during the colder months is higher by 3oC and COP of the system
drops to 3 compared to 3.3 during the warmer months.
The capacity is increased by raising the condensing pressure, which is
achieved by closing the electronic valve in the refrigerant line after the
condenser. The temperature of the brine is maintained by regulating the flow
rate of brine into the de-superheater. It is more convenient to understand these
controls on individual units over shorter period and has been discussed in the
following section.
4.2.4 Performance of Medium temperature Unit (KA3)
Figure 4.9 presents the monthly average of cooling capacity, COP, condensing
temperature, heat recovered and the borehole capacity
21
80
5.0
70
4.5
60
4.0
40
3.5
COP
Q,E(kW),T(C)
50
30
3.0
20
2.5
10
0
2.0
Oct-08
Nov-08
Dec-08
Jan-09
Feb-09
Mar-09
Apr-09
Month
T_cond
Cooling capacity
Q_desuperheater
T_ambient
Q_borehole(subcooling)
COP
Figure 4.9: Average monthly performance of medium temperature unit
(KA3) in TC2
The heat recovery capacity increases, while the cooling capacity decreases.
The increase in heat recovery capacity and increase in condensing temperature
are directly proportional. This reduces the COP of the refrigeration system. The
borehole sub cooling capacity is low during the period from October to January,
while condensing temperature has been increasing during these months. Sub
cooling from the ground source reduces the compressor power consumed and
therefore improves COP. The improvement in COP due to sub cooling can be
observed in the months of February and March.
Heat recovery and control of medium temperature unit for 24 hours is presented
in figure 4.10. The top curve is the percentage opening of the electronic valve in
the refrigerant line after the condenser. The middle curve is the discharge
pressure. The bottom curve is the heat recovered in the de-superheater. It is
seen that heat recovery starts about 7 AM and stops at around 10 PM. The
discharge pressure is increased after 7 AM to meet the heating demand. The
pressure is increased by partially closing the electronic valve. The valve is
almost 100% during the night time when there is no heating demand.
22
100
90
80
Q(kW), P(bar), Capacity(%)
70
60
50
40
30
20
10
0
00:00
02:25
04:30
06:55
08:40
10:20
12:10
13:55
16:00
17:50
20:35
23:00
Tiime (march 24,2009)
Electronic valve opening
2 per. Mov. Avg. (Discharge Presure)
2 per. Mov. Avg. (Q_desuperheater)
Figure 4.10: Operation of medium temperature unit over a period of 24
hours in TC2
4.2.5 Performance of booster unit (KAFA1)
The monthly average of cooling capacity, COP and heat recovery capacity are
presented in figure 4.11
23
3.5
60
50
3.0
40
COP
Q,E(kW),T(C)
2.5
30
2.0
20
1.5
10
0
1.0
Oct-08
Nov-08
Dec-08
Jan-09
Feb-09
Mar-09
Apr-09
Month
T_cond
Q_desuperheater
Q_borehole(subcooling)
cooling capcacity
COP
Figure 4.11: Average monthly performance of (booster system) (KAFA1) in
TC2
It is seen that that the heat recovery capacity is limited in booster system. It is
also necessary to maintain higher condensing temperature of about 23oC in the
colder months to recover the same or less amount of heat to that of October
and April. The cooling capacity drops quite sharply in colder month. The
decrease in cooling capacity is due to the decrease in cooling demand in the
medium temperature unit. The cooling capacity of low temperature unit is found
to be rather constant.
The control of the booster system over a period of day and heat recovery
capacity is presented in figure 4.12. It can be seen that the discharge pressure
is controlled within smaller range of variation than in the case of KA3. This
could be accounted for the fact that the heat recovery capacity in the booster
system is higher than in case of the medium temperature level system.
24
100
90
80
Q(kW),P(bar), Capacity (%)
70
60
50
40
30
20
10
0
00:00
01:40
03:20
05:05
06:45
08:25
10:05
11:45
13:25
15:05
16:45
18:25
20:05
21:45
23:25
Time (March 24, 2009)
Electronic valve opening
2 per. Mov. Avg. (Discharge pressure)
2 per. Mov. Avg. (Q_desuperheater)
Figure 4.12: Operation of booster unit (KAFA1) over a period of 24 hours.
25
5
Simulation models
Comparative study of three different heat recovery system solutions have been
done using simulation model in engineering equation solver (EES). The system
solutions are:
•
CO2 trans-critical system with heat recovery with heat pump. The model is
similar to the one in section 4.1 and is designated as TC1.
•
CO2 trans-critical system with heat recovery only in de-superheater. It is
similar to the experimental supermarket discussed in section 4.2, except that
external sub-cooling is omitted in the model. It is designated as TC2
•
CO2/R404A cascade system. R404A is used as refrigerant on the high
temperature side and CO2 is used as refrigerant in the low temperature side.
The intermediate fluid (brine) is circulated between the evaporator of R404A
and CO2 condenser. The medium temperature cabinets are also cooled with
the brine. It is designated as CC.
The parameters of the heat exchangers used in the models have been
calculated from the experimental values. However, in some cases averages has
to be taken in order to incorporate fair comparison. An important term that
needs to be explained is “floating condensing”. Though, the lowest condensing
temperature is limited to 10oC, systems normally operating within this condition
are referred to as floating condensing. COP of heating has been introduced to
compare heat recovery in different systems and is expressed as;
COPHeat =
Q& heating
− E&
E&tot ,recovery
(5.1)
tot , floating
Where, Q& heating is net heat supplied for heating, E& tot ,re cov ery is the total power
consumption when operating in heat recovery mode and E& tot , floating is the total
power consumption as if the system is running in floating condensing and
without heat recovery; thus, the system is operating at the lowest condensing
pressure possible and consumes the lowest energy to provide the required
cooling.
A term that has been introduced is the heat recovery ratio (HRR), which is the
ratio of the heat recovered to the cooling capacity. The COP of the refrigeration
system is the total system COP and not of individual units.
26
5.1
Assumptions and Correlations in the models
5.1.1 Cooling capacity as a function of outdoor temperature
The cooling capacity is mainly dependent on outdoor temperature and relative
humidity. Since indoor relative humidity is attached to the outdoor temperature,
the cooling capacity increases with increasing outdoor temperature. Jaime Arias
has compiled considerable plots for compressor power in respect to outdoor
temperature (Arias, 2005). It was observed from the plots in above literature
that during the cold season between October and May the compressor power is
more or less constant. During this period outdoor temperature is less than 10oC.
Thus average power has been taken for temperature below 10oC. The
maximum compressor power consumption is reached in the month of July;
therefore maximum power and temperature have been read from the plots
presented by Jaime Arias (Arias, 2005) and tabulated in table 5.1.
Table 5.1: Average and maximum compressor power (Adapted from Jaime
Arias, 2005)
Maximum
compressor
power
Average compressor Maximum
(kW) on the days
Power (kW) for out outdoor
maximum
door
temperature temperature with
o
o
outdoor temperature
below 10 C
( C)
Supermarket
Sala
46
21
56
Hjo
25
24
32
Hedeamora
15
28
27
Täby
60
28
100
Kista
35
23
65
Farsta
40
23
56
Average
36
25
56
A correlation has been developed for the compressor power as a function of
outdoor temperature in the range between 10oC and 25oC.
E& compressor = 1.32 * Toutdoor + 23.6[kW ]
(5.2)
Using the correlation 5.2, the compressor power at outdoor temperature of 35oC
has been extrapolated and the value of 70kW was obtained. Average power
below 10oC is approximately 50 % of the maximum power at maximum
temperature. Therefore compressor power above 10oC is obtained in terms of
percentage of the maximum power and outdoor temperature as follows;
E& compressor = 0.02 * Toutdoor + 30[kW ]
(5.3)
Applying the same relation to the cooling capacity, the following correlation is
used to calculate cooling capacity at different outdoor temperature:
Q& 2 = 0.02 * Toutdoor + 30[kW ] : Toutdoor > 10o C
27
(5.4)
Q& 2 = 0.5 * Q& max [ kW ] ; Toutdoor ≤ 10o C
(5.5)
Where, Q& 2 is the cooling capacity in kW, Toutdoor is the outdoor temperature in oC.
and Q& max maximum/design cooling capacity in kW.
The profile of cooling capacity obtained from the above relation is presented in
figure 5.1.
% cooling capacity
100
75
50
25
0
-10
0
10
35
Out door temperature (C)
Figure 5.1: The profile of cooling capacity as function of outdoor
temperature
In the above figure the cooling capacity is plotted as a function of outdoor
temperature. The cooling capacity is expressed in percentage of the maximum
cooling capacity/design capacity. The full capacity, 100% at the vertical axis in
the plot, is reached at outdoor temperature of 35oC. Below this temperature the
cooling capacity is decreasing due to the decreasing humidity ratio.
The equations 5.4 and 5.5 describes the correlation between outdoor
temperature and the cooling capacity and are used in the simulations models to
obtain the cooling capacity as function of outdoor temperature.
5.1.2 Heating period
To have reasonably fair comparison of different systems, it was necessary to
define the heating period in terms of outdoor temperature. Monthly average
temperatures on the start and end of heating season for some cities in Sweden
have been compared in table 5.2. The average monthly temperatures for cities
of interest have been taken from an online web accessible at
www.weatherebase.com. The monthly data obtained from the weather base
have been compared to past record from SMHI for four cities (SMHI, 2009).
From the information of heating season and average temperature of these
cities, it is seen that heating season begins when the average outdoor
temperature is below 10oC.
28
Table 5.2: Heating season and average monthly temperature. (Jonsson.
H, et al, 2005, SMHI, 2009 and weather base, 2009)
Start of
Monthly average
End of
Monthly average
Location
heating
heating
temperature (oC)
temperature (oC)
weather
weather
SMHI
SMHI
base
base
Kiruna
15-Jun
8
15-Aug
11
Umeå
01-Jun
12
01-Sep
8
Östersund
01-Sep
8
06-Jun
10
Karlstad
15-Sep
11
10
16-May
10
10
Stockholm
18-Sep
11
11
14-May
10
10
Goteborg
24-Sep
12
14
07-May
11
13
Average
10
11
10
11
In a supermarket, heating is required throughout the year. However heating
demand reduced drastically when daily average outdoor temperature reaches
10oC. Therefore 10oC has been fixed as outdoor temperature below which heat
recovery is required in the simulation models.
5.1.3 Other refrigeration and heat exchanger parameters
To enable fair comparison of the three systems, field measurements from three
supermarkets have been used as reference to fix the parameter of refrigeration
and heat exchangers. The assumed values are presented in table 5.3.
Table 5.3: Parameters assumed in the models
Description of parameter
Value
Evaporation temperature of low temperature cabinet
-35oC
Evaporation temperature of medium temperature cabinet
-10oC
Approach temperature difference of dry cooler
5K
Approach temperature difference of condensers
5K
LMTD for condenser of Cascade system
6K
Internal heat exchanger efficiency R404A unit in CC system
50%
Internal heat exchanger efficiency CO2 unit in CC system
20%
Capacity of pump in percentage of cooling capacity of high 4%
stage(R404A) unit in cascade system
Heat losses in the tubing of intermediate brine loop
10kW
Oil cooler capacity of CO2 compressors
15%
Heat losses from CO2 compressors
7%
The performances of the compressors have been modeled by curve fitting the
manufacturers’ data
29
Heat recovery with heat pump
+
Q_max =300kW
-
5.2
45oC
T_hp in= 13o C
Heat pump COP= 3.6
T_con,b,o =18oC
dTapp=5K
dTsc=2K
dTapp=5K
Oil cooler
Inter-Cooler
dTapp=5K
Oil cooler
dTsc=2K
dTsc=2K
Qoilcooler=15%
Qoilcooler=15%
dTsh,ex=10K
dTsh=10K
dTsh=10K
-10oC
-35oC
dTsh,ex=15K
KA1/KA2
- Pressure
-Temperature
FA1/FA2
Figure 5.2: Schematic of heat recovery with heat pump system solution
(TC1).
This is a parallel refrigeration system with CO2 trans-critical system in both low
and medium temperature levels. It has a single-stage compression system on
the medium temperature unit and two-stage in the low temperature unit. The
condensers in both units reject heat to the secondary coolant which is
connected to the heat pump system on the heat recovery system. The return
from the heat pump is connected to the roof top dry cooler. The heat pump on
the heat recovery side is designed to operate with COP of 3.6 within the
temperature limit at the heat pump evaporator of 13oC/7oC. Operating under
this condition the supply temperature from the heat pump is 45oC. Alternatively
the heat pump could be operated to supply temperature of 35oC to the heated
space with COP of 4.7.
In order to obtain the temperature of brine at inlet of the heat pump evaporator
(T_hp, b,in in figure 5.2) of 13oC, the supply temperature of coolant from
condenser (T_con,b,o) has to be 18oC. This is taking account of temperature
difference of 5K between the fluids in the heat exchanger. The system has been
simulated in the floating condensing mode. COP and pressure at which the
corresponding supply temperature of brine from condenser is 18 is selected and
the system is controlled to run at this operating condition for ambient
temperature up to 10oC. Beyond this the refrigeration system is again set back
to floating condition. Figure 5.3, shows the plot of COP in floating condensing
30
mode (COP_FC) and after the condition have been changed to operate the
heat pump during the heating season (COP_HP_S45). In the heat pump mode
the supply temperature of the heat pump is 45oC. The lowest condensing
temperature was set to 10oC; therefore, in floating condensing mode the curve
of COP is constant for ambient temperatures below 0oC. In the heat pump
mode the COP is constant at about 2.6 for ambient temperature below 10oC
and for ambient temperature above 10oC; the refrigeration system operates in
floating condensing mode.
50
4
45
3.5
40
3
35
2.5
25
2
COP
T(C)
30
20
1.5
15
1
10
0.5
5
0
0
-10 -8
-7
-5
-3
-1
0
2
T_hp,b,in
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
COP_HP
COP_FC
T_con,b,o
Figure 5.3: COP in floating condensing and heat pump modes of operation
for TC1.
At about 18oC ambient, the transition from sub- to trans-critical region takes
place.
5.2.3 Evaluation of different modes of operation
The system has been simulated to operate as heat pump with supply
temperature of 45oC. Presented in the figure 5.4 are system’s COP operating in
floating condition with coolant loop (COP_FC) and heat recovery mode
(COP_HP). Also presented is the total compressor power of the refrigeration
system in floating condensing (E_ref_FC) and recovery mode (E_ref_HP). Total
heat delivered by the heat pump and the heating COP are plotted for the
heating period. Power consumption of the compressors are only considered,
therefore does not account the auxiliary power such as fan power or circulation
pumps.
31
4
300
3.5
250
3
200
2
150
COP
E,Q (kW)
2.5
1.5
100
1
50
0.5
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
E_ref_FC
E_ref_HP
E_tot_HP
Q_HP(HR)
COP_FC
COP_HP
COP heating
Figure 5.4: Operation in the heat pump mode with heat pump supply
temperature of 45 o C compared to the floating condensing mode for TC1
As can be observed from figure 5.4, power consumption could increase by
average of 35% as penalty of operating the refrigeration system at the higher
pressure to operate the heat pump system. Heat capacity up to 187kW is
available while operating under 50% of cooling capacity of the refrigerating
units. The condensing temperature of the refrigeration system while operating in
heat pump mode is about 19oC. This condensing temperature corresponds to
operation of the refrigeration system at ambient temperature above 10oC.
Therefore beyond outdoor temperature of 10oC, there is no increase in
compressor power of the refrigeration units due to heat recovery.
Performance in heat recovery mode is compared to the cases when the system
operates in floating condensing with air cooled condenser and with condensing
with coolant loop. System’s cooling COP with air cooled condenser
(COP_FC_AC) and coolant (secondary fluid) cooled condenser (COP_FCB) is
presented in figure 5.5. There is a shift in the curves of COP in two cases which
is due to the difference in condensing temperatures. Due to additional
temperature difference introduced by the coolant, for same ambient conditions,
the system with secondary fluid runs at condensing temperature higher than air
cooled condenser.
32
50
4
45
3.5
40
3
35
25
2
COP
Temperature (C)
2.5
30
20
1.5
15
1
10
0.5
5
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
T_con_FCB
T_con_FC_AC
COP_FCB
COP_FC_AC
Figure 5.5: Operation of refrigeration system in air cooled condenser and
brine (secondary fluid) cooled condenser for TC1.
To study the performance of heat pump with supply temperature of 35oC, the
performance in the heat pump mode with supply temperature of 35oC is shown
in figure 5.7. Heat pump operates at COP of 4.7, which is higher than the COP
of het pump while operating with supply temperature of 45oC. This reduces the
total power consumption. The average increase in the power consumption of
refrigeration system as a result of operating in this mode is about 34%
compared to that consumed when operating in floating condensing mode.
Average COP of heating is 3.9.
33
5
300
4.5
250
4
3.5
200
2.5
150
COP
Q,E(kW)
3
2
100
1.5
1
50
0.5
0
0
-10 -8 -7 -5 -3 -1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
E_ref_FC
E_ref_HP
Q_HP(HR)
E_tot_HP
COP_FC
COP_HP
COP_heating
Figure 5.7: System solution with heat pump supply temperature of 35oC in
TC1.
At outdoor temperature higher than 10o C the refrigeration system is already
operating at higher pressure so system can provide the required coolant
temperature of 18oC, therefore there is no need for increasing the condensing
pressure. Increase in total power at this point is only from heat pump. Another
alternative mode of operation that is considered is to avoid the use of heat
pump in the heat recovery system where the refrigeration system will operate
as a heat pump. It is simulated by increasing the condensing pressure in the
refrigeration system to insure a supply temperature of coolant (inlet to the
heating system) out of the condenser of 45oC. The performance parameters of
such system are plotted in the following figure 5.8.
34
350
4
3.5
300
3
250
2.5
2
COP
Q,E(kW)
200
150
1.5
100
1
50
0.5
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
Q_con_FC
E_ref_FC
Q_con_FH(HR)
E_tot_FH
COP_FC
COP_FH
COP heating
Figure 5.8: Operating at conditions to supply temperature of coolant from
condenser at 45oC in TC1.
It can be seen that average COP of heating is about 2.5 and compressor power
increases by more than 300% in extreme case and 270% in average. Also at
this operating condition COP of refrigeration system is less than 1 and
discharge pressure under this operating condition is about 114 bars. This would
mean that compressors of medium temperature units operating at pressure
ratio more about 5.
Figure 5.9, shows the performance of system operating at fixed head pressure
to supply a supply temperature from the condenser at 35oC.
35
400
4.5
4
350
3.5
300
3
2.5
COP
Q,E(kW)
250
200
2
150
1.5
100
1
50
0.5
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12
14 16
18 19
21 23 24
26 28
30 31
33 35 37
38 40
T_ambient (C)
E_ref_FC
E_ref_FH
Q_con_FH (HR)
Q_con_FC
COP_FC
COP_FH
COP_heating
Figure 5.9: Operating at conditions to supply temperature of coolant from
condenser at 35oC in TC1.
With average heating COP about 3.3 and COP of the refrigeration system about
1.2 during the heating season, it provides heat recovery capacity of 168 kW.
Though capacity of heating is quite large and COP of heating is quite
satisfactory, COP of the refrigeration system is reduced considerably.
Since the basic case of reference is the floating condensing with air cooled
condenser instead of the coolant loop and dry cooler case, it is interesting to
compare the performance of the refrigeration system with heat pump to the
refrigeration system with air cooled condenser as presented in figure 5.10. The
compressor power consumption of the refrigeration system in air cooled floating
condensing mode (E_ref_FC_AC) is compared to the power consumption when
operating at fixed head pressure to recovery heat at 35oC (E_ref_FH). The
compressor power consumption of the refrigeration system is higher in heat
recovery mode than in air-cooled condenser mode even beyond the heating
season. It is because of the temperature difference introduced by the secondary
fluid used in the heat recovery mode.
36
500
4
450
3.5
400
3
350
Q,E (kW)
250
2
200
COP
2.5
300
1.5
150
1
100
0.5
50
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
E_ref_FC_AC
E_ref_FH
Q_con_FH
COP_FC_AC
COP_FH
COP heating
Figure 5.10: Comparison of air cooled mode and heat recovery mode with
supply temperature of 35oC in TC1
Since cooling demand tend to decrease when the heating is required in cold
season, a ratio of heat that is available to the cooling capacity could be used as
a parameter to compare different systems. The ratio of capacity of heat
recovered to the cooling capacity for different modes of operation are presented
in figure 5.11. The heat recovery ratio in heat pump mode with supply
temperature of 45oC is represented as HP_S45 and heat pump mode with
supply temperature of 35oC as HP_S35. The heat recovery ratio for fixed head
of refrigeration system to supply heat at 45oC and 35oC are represent as
FH_T_b,o 45 and FH_T_b,o 35 respectively. Beyond the heating period the
ratio represent the ratio of heat rejected to cooling capacity.
37
2.2
Ratio of Heat recovery to cooling capacity
2.0
1.8
1.6
1.4
1.2
1.0
-10
-8
-7
-5
-3
-1
0
2
4
6
7
9
11
12
14
16
18
19
21
23
24
26
28
30
31
33
35
37
38
40
T_ambient (C)
HP_S45
FH_T_b,o 45
FH_T_b,o 35
HP_S35
Figure 5.11: Ratio of heat recovery/rejection to cooling capacity in
different modes of operation for TC1
Considering the ratio of heat recovered to the cooling capacity, heat recovery
with the fixed head pressure to supply temperature of brine out of condenser at
45oC is seen to be highest at almost double the cooling capacity. This is mainly
due to the fact that this mode of operation has the highest power consumption.
However, this ratios and power consumption in different systems must be
compared for the different modes.
Figure 5.12 shows the total compressor power in different mode of operation. In
the same plot the compressor power in floating condensing mode without heat
recovery and for the case if air cooled condenser would be used is included in
the plot.
38
180
160
140
Power (kW)
120
100
80
60
40
20
0
-10
-8
-7
-5
-3
-1
0
2
4
6
7
9
11
12
14
16
18
19
21
23
24
26
28
T_ambient
E_FC
E_tot_HP_S45
E_FH_T_b,o 45
E_FH_T,b,o 35
E_tot_HP_S35
E_FC_AC
Figure 5.12: Total power consumption in different mode of operation for
TC1.
Comparing figure 5.11 and 5.12, Heat pump mode with supply temperature of
45oC has heating ratio of 1.8 and at the same time power consumption is just
over 80% of the total power consumed by system operating with fixed head
pressure with supply temperature of brine at 45oC. Comparatively for both the
temperature levels of 35oC and 45oC, heat pump system has better
performance than the fixed head pressure system.
39
5.3
Heat recovery in de-superheater only (TC2)
+
30
+
35 Floor heating
dTapp=5K
dTsc=0.5K
dTsc=0.5K
dTapp=5K
dTapp=5K
dTapp=5K
Qoilcooler=15%
Qoilcooler=15%
dTsh=10K
Tevap
-10oC
dTsh=10K
Oil cooler
dTsh,ex=10K
Tevap -35
KA3
Qoilcooler=15%
-Pressure
-Temperature
Tevap -35oC
dTsh,ex=15K
dTsh=10K
KAFA1/KAFA2
Figure 5.13: Schematic of refrigeration system with heat recovery in desuperheater region (TC2).
Such a system solution is already in operation in a supermarket in Sweden and
analysis of field data is discussed in section 4.2. However, in the simulation
models the sub-cooling from borehole is not considered in the simulation model
discussed in this section. It is interesting to study this system solution for many
aspects. The system solution is different from the earlier design on both
refrigeration and heat recovery side. The refrigeration side consists of two
parallel circuits. One of the parallel circuits consists of a medium temperature
level. The second parallel circuit is a booster system consisting of medium
temperature level circuit and a low temperature level circuit.
In this design, the heat recovery takes place only in the de-superheater region
and condensation takes place in the air cooled condenser. The system
simulated in two modes of operation, to supply the temperature of recovered
heat at 35oC and 45oC. In both mode of operation, the heat recovery capacity is
then fixed at 30% of the total heat rejected at the condenser. This value is fixed
based on the field measurements of a similar heat recovery system, discussed
in section 4.2. Concerning the operation during heat recovery with supply
temperature of 35oC, the system is simulated to operate with variable capacity
of heat recovery; so the system was not controlled according to requirements of
the heat recovery. Then the refrigeration system is set to operate at elevated
operating pressure to recover 30% of the rejected heat at a temperature of
35oC. Similarly operating conditions are selected for heat recovery at 45oC. The
COPs of the system operating in floating condensing, and fixed percentage of
40
heat recovery at supply temperature of 35oC and 45oC are shown in Figure
5.14.
4
3.5
3
COP
2.5
2
1.5
1
0.5
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
COP_FC
COP_FH_35
COP_FH_45
Figure 5.14: Coefficient of performance of system in different modes of
operation in TC2
In the following figures (5.15 and 5.16) some values beyond the heat recovery
period are not shown in the plot, these values have not been included because
they are not much of interest for the current analysis. Figure 5.15 shows the
heat recovery and power consumption in floating condition and operation at
fixed head pressure during heating period.
41
80
1.5
70
1.4
1.4
60
1.3
1.3
COP
Q,E (kW)
50
40
1.2
30
1.2
20
1.1
10
1.1
0
1.0
-10
-8
-7
-5
-3
-1
0
2
4
6
7
9
11
12
14
16
18
T_ambient (C)
E_FC
E_FH
Q_HR_FH
Q_HR_FC
COP heating
Figure 5.15: heat recovery at supply temperature to heat recovery system
of 35oC in TC2
The power consumption in the refrigeration system due to the heat recovery of
30% of the total heat rejected in the de-superheat region increases by 54% and
subsequent increase amount of heat recovery is over 100 %. Maximum COP of
heating is just over 1.4.
When the refrigeration system is operated at an elevated pressure to maintain
heat recovery supply temperature of 45oC, the power consumption increases by
85 % and increase in quantity of heat recovery is 180 %. This can be observed
in figure 5.16.
42
1.4
60
1.3
50
1.3
40
1.2
30
1.2
20
1.1
10
1.1
COP
Q,E(kW)
70
0
1.0
-10
-8
-7
-5
-3
-1
0
2
4
6
7
9
11
12
14
16
18
T_ambient (C)
Q_FC
E_FC
E_FH
Q_FH
COP heating
Figure 5.16: Heat recovery at supply temperature to heat recovery system
of 45oC in TC2
Since in both the cases the heat recovery capacity is fixed at 30% of the total
heat rejected, the ratio of heat recovery capacity to cooling capacity at 35oC and
45oC is same and equal to 0.4. Thus, the compressor power consumption is
an important parameter for the comparison. Compressor power consumption in
different mode of operation is presented in figure 5.17.
43
80
70
Power (kW)
60
50
40
30
20
10
-10
-8
-7
-5
-3
-1
0
2
4
6
7
9
11
12
14
16
18
T_ambient (C)
E_FC
E_FH_35
E_FH_45
Figure 5.17: Compressor power for different capacity and temperature of
heat recovery in TC2
It can be seen that the compressor power consumption is considerably higher
to operate the system in heat recovery mode for supply temperature of 45oC. It
is because the refrigeration system needs to be operated at higher condensing
pressure to recover same capacity of heat at 45oC compared to the 35oC case.
So far the percentage of heat recovery is only 30% of the total heat rejected.
The performance of the system with different amount of heat recovery was
considered and performance to recover 70% of the heat at 35oC is presented in
figure 5.18.
44
3.0
100
2.5
80
2.0
60
1.5
40
1.0
20
0.5
0
0.0
-10
-8
-7
-5
-3
-1
0
2
4
6
7
9
11
12
14
16
18
T_ambient (C)
Q_HR_FC
Q_HR_FH
E_FH
E_FC
COP heating
Figure 5.18: Performance of TC2 with heat recovery of 70% of the total
heat rejected at de-superheater region at supply temperature of 35oC
45
COP
Q,E(kW)
120
5.4
Heat recovery in R404A/CO2 cascade system
Figure 5.19: Schematic of heat recovery system solution in R404A/CO2
cascade system (CC)
This is a cascade system solution with R404A in the high stage and CO2 in the
low temperature. Brine is circulated in the medium temperature cabinets and
the CO2 condenser. Heat recovered in two temperature levels on the R404A
heat rejection side in two separate loops, de-superheater and the condenser
region.
First the system is operated in the floating condensing. To recover heat at
supply temperature of 35oC, the refrigeration system will have to operate at the
condensing temperature of about 36oC which results in COP of 1.7. Figure 5.20
shows the operation of refrigeration system in floating condensing mode and
fixed head pressure to provide brine supply temperature of 35oC. Top two lines
in the plot are COP of the refrigeration system on floating condition (COP_FCB)
and with the conditions set for heat recovery (COP_FH). The two COP lines join
above an outdoor temperature of 10oC and follow the shape of floating
condensing mode. The pair of curves below this plot shows heat rejected on
46
300
3
250
2.5
200
2
150
1.5
100
1
50
COP
Q,E(kW)
floating condensing (Q con_FCB) and on heat recovery mode (Q con_FH(HR)).
Next pair of curves is the represents the compressor power consumption in
floating condensing mode (E_FCB) and in heat recovery mode (E_FH). The last
curve is the heat from de-superheater (Q_desup_FH).
0.5
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12
14 16
18 19
21 23 24
26 28
30 31
33 35 37
38 40
T_ambient (C)
Q con_FCB
E_FCB
E_FH
Q con_FH(HR)
Q_desup_FH
COP_FCB
COP_FH
o
Figure 5.20: Heat recovery with supply temperature of 35 C in cascade
system (CC)
Power consumption of the refrigeration system increase by 63% when,
operating under heat recovery mode to deliver supply temperature of 35oC. Out
of 169 kW of total heat recovery capacity 42 kW is high temperature heat from
the de-superheater at about 76oC. It is seen from the plot that considerable
capacity of heat from de-superheater is available even while operating on
floating condensing mode. Therefore it can be used for other applications in the
supermarket throughout the year. Average COP of heating in this system is
about 7.
To recover heat at condenser supply temperature of 45oC, the refrigeration
system is made to operate at corresponding COP of 1.4 during the heat
recovery period. Figure 5.21 compares the performance with that of floating
condensing mode.
47
400
3
350
2.5
300
2
200
1.5
COP
Q,E(kW)
250
150
1
100
0.5
50
0
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
Q con_FCB
E_FCB
E_FH
Q con_FH(HR)
Q_desup_FH
COP_FH
COP_FCB
o
Figure 5.21: Heat recovery with supply temperature of 45 C in cascade
system (CC)
In case of heat recovery with supply temperature of 45oC average, the increase
in compressor power consumption is more than 94%, when the outdoor
temperature is below 1oC, the power consumption is about 200% of the power
that refrigeration system would consume under floating condition. Average
COP of heating is 5.
The system performance of heat recovery with supply temperature of 35oC has
been compared to the refrigeration system operating under air cooled floating
condensing mode in figure 5.22. Presented in the plot is capacity of heat
recovery during the heating season (Q con_FH (HR)), COP of refrigeration in air
cooled floating condensing mode (COP_FCAC) and heat recovery mode
(COP_FH). The last pair of curves is the compressor power consumption in
heat recovery mode (E_FH) and air cooled condensing mode (E_FCAC).
48
4
170
3.5
150
3
130
110
2
90
COP
Q,E(kW)
2.5
1.5
70
1
50
0.5
30
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T amb (C)
Q con_FH(HR)
E_FH
E_FCAC
COP_FCAC
COP_FH
Figure 5.22: Comparative performance of heat recovery with 35oC and aircooled condenser in floating condensing mode in cascade system (CC)
One way to study the system performance is to compare the heat recovery ratio
at different temperature levels of heat recovery. Figure 5.23 presents the heat
recovery ratio at 45oC (FH_T, b, o, 45), 35oC (FH_T, b, o, 35) and on floating
condensing (FCB) mode.
1.9
Ratio of heat recovery to cooling capacity (HRR)
1.8
1.7
1.6
1.5
1.4
1.3
1.2
1.1
1
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
FCB
FH T_b,o 35
FH T_b,o 45
Figure 5.23: heat recovery ratio for different heat recovery temperature
levels and floating condensing mode in CC.
49
It can be seen that heat recovery ratio is higher for heat recovery at 45oC than
for 35oC and floating condensing mode. However net increase in compressor
power consumption is also considerably high as can be compared in figure
5.24. The top tow curves represents the power consumption for heat recovery
at 45oC (E_FH_T,b,o 45) and 35oC (E_FH_T,b,o 35). Two curves in the bottom
are power consumption when operating the refrigeration system in floating
condensing (E_FCB) and with air cooled condenser (E_FCAC) mode.
180
160
140
Power (kW)
120
100
80
60
40
20
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
E_FCB
E_FH_T,b,o 35
E_FCAC
E_FH_T,b,o 45
Figure 5.24: Power consumption in different modes of operation in
cascade system (CC)
5.5
Comparison of different heat recovery system solutions
To compare different system solutions it is proper to begin with the discussion
on performance of different refrigeration system solution operating in floating
condensing mode. Figure 5.25 presents the COP of refrigeration of different
system solution in floating condensing mode. It can be see that CO2 transcritical system 1 and CO2 trans-critical system 2 has better performance at
lower outdoor temperatures. Trans-critical system 2 has the highest COP and
cascade system has the lowest COP for outdoor temperature below 10oC. For
outdoor temperatures higher than 23oC, cascade system has the highest COP
and trans-critical system 1 has the lowest.
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4
3.5
3
COP
2.5
2
1.5
Heat recovery
1
0.5
0
-10 -8
-7
-5
-3
-1
0
2
4
6
7
9
11 12 14 16 18 19 21 23 24 26 28 30 31 33 35 37 38 40
T_ambient (C)
TC1
TC2
CC
Figure 5.25: COP of TC1, TC2 and CC in floating condensing mode.
Some parameters that could be used to compare different heat recovery system
solutions are COP of refrigeration system, COP of heating and heat recovery
ratio. The comparison of these parameters for TC1, TC2 and CC for heat
recovery temperature of 35oC and for the heating season is presented in figure
5.26. From left in the figure are performance of TC1 in heat pump mode and
without heat pump; the TC2 system to recover 30% and 70% of heat
respectively. The last group of columns is the performance of cascade system
(CC).
It can be seen from the figure that TC1 with heat pump system has the highest
COP and TC1 without the use of heat pump has the lowest COP. The cascade
system has the highest COP of heating and HRR ratio while TC2 with heat
recovery percentage of 30% of the total heat rejected has the lowest COP of
heating and HRR.
51
7.00
COP
COPheating
HRR
6.00
COP,HRR, COP heating
5.00
4.00
3.00
2.00
1.00
0.00
TC1 with heat pump
TC1 without heat pump
TC2 with 30% heat
recovery
TC2 with 70% heat
recovery
Cascade (CC)
Heat recovery system solutions
Figure 5.26: Heat recovery ratio for different heat recovery system
solutions for at supply temperature of 35oC.
Though the comparisons of the system solutions are self explanatory, this study
is insufficient to draw general conclusion on the performance of the systems.
6
Discussion
With the drive to reduce net energy consumption, heat recovery is gaining
popularity in supermarket refrigeration systems. In Sweden most of the new
supermarkets are installed with heat recovery systems. Field measurements of
trans-critical system 1 and trans-critical system 2 did give some insight on the
performance of heat recovery. However more detailed measurement for longer
period is necessary for trans-critical system 2. In case of trans-critical system 1,
actual field measurement on the heat recovery system would have provided
more information on the performance of the system. Never the less; the
analysis of field measurement in trans-critical system 1 and trans-critical system
2 has highlighted enough details to further the discussion of the subject.
For trans-critical system 1, if the refrigeration system was operated without heat
recovery, condensing temperature was as low as 13oC, where as to operate in
heat recovery mode with heat pump the minimum condensing temperature has
to be 20oC. This puts burden on the refrigeration system by reducing the COP
of refrigeration system. But on the other hand, large capacity of heat is
available. The power consumption as calculated is for the maximum capacity
therefore the high values of power consumption by the heat pump may be
misleading. In reality the heat pump could be running in part load thus lower
52
consumption. The measurement of the heat recovery side in the future studies
would enable better observation of the heat recovery performance.
From the field measurement of trans-critical system 2, it was found that up to
30% of the total heat rejected can be recovered at a temperature of 35oC. The
rest of heat is rejected to the ambient through the air cooled condenser. It is
however seen though amount of heat recovered is less compared to transcritical system 1 the condensing temperature has to be raised above 20oC,
which increases the pressure ratio thus the COP of refrigeration drops. There is
also additional consumption of power by the condenser fans to reject the
excess heat. The performance of this system is better because there is the
external sub-cooling from the ground source which improves its COP.
The field measurement for cascade system has not been conducted in this
study.
As discussed in section 5.5, using the models the comparative different system
solutions could be compared for system performance such as COP of
refrigeration system. COP of heating and heat recovery ratio (HRR). It is
however necessary to conduct detailed field measurements and detailed
analysis prior to drawing conclusion on the system performances. Further for
completed evaluations of the systems it is necessary to study the annual energy
consumption of the systems taking consideration of cooling and heating loads
over the year in different climates.
7
Conclusion
With the energy price continuing to rise, reducing the net energy purchase is a
goal of any enterprise. Many new supermarkets refrigeration systems in
sweden are equiped with the heat recovery system. While it has presented the
opportunities for energy conservation there remains the challenge of matching
the temperature and the amount of heat recovered to the demand in the
supermarket.
In this thesis the field measurement have been performed in two super markets
with CO2 trans-critical system. The heat recovery systems were different in the
two cases. In CO2 trancritical system 1, where heat recovery system with a
heat pump it is found that the minimum condensing temperature of the the
refrigeration has to be 20oC compared to possible lower codensing temperature
of 10oC. The maximum COP of heating is limited to design COP of the the heat
pump. In CO2 trans-critical system 2, the condensing temperature during the
heat recovery mode is about 23oC. The increased condensing temperature
doesn’t effect the COP because of the enhancement in the cooling capacity due
to external sub cooling from the ground source. Heat rejection from oil cooler in
both the cases were found to be more or less constant over the year. The heat
recovery capacity from the inter cooler was found to be constant too.
One of the problems faced in this project was to get the reading at the point of
interest in both the refrigeration side and on the heat recovery side. The
comprehensive field measurement that records reading at both available heat
53
on refrigeration side and heat using end such as HVAC and service water
heater is necessary to give better indication of system performance.
Future works should incorporate the field measurement of cascade system,
long term measurement of trans-critical system 1 and trans-critical system 2.
The study must also incorporate the measurement of annual energy
consumption taking consideration of cooling and heating needs in different
climates. Theoritical analysis was carried out using the computer simulation
models for three different system solutions: CO2 trans-critical system 1, CO2
trans-critical system 2 and R404A/ CO2 cascade system.
In reality the systems solutions vary in some form or other. This makes it difficult
to compare different systems based on a certain performance criteria such as
COP refrigeration system, COP of heating, power consumption heating and
cooling capacity etc. Therefore computer simulation modeling is important to
develop fair comparison of different systems.
54
8
References
Arias J, 2005 “Energy Usage in Supermarkets-Modelling and Field
measurements”, Doctoral thesis, Department of Energy Technology, KTH,
Stockholm, Sweden
Arias J, Lundqvist P, 2006 “Heat recovery and floating condensing in
supermarkets” Journal of Energy and Buildings, Vol.38, pp. 73-81
Brownell K. A, 1998 “Fixed and Floating Head Comparison for Madison Ice
Arena” Report, Energy Centre of Wisconsin, WI, USA
Dorin (2009), “The widest CO2 compressor range, carbon dioxide for all your
needs”,
http://www.dorin.com/documents/Download/19/CO2_-0809a.pdf
accessed on march 2009
Granryd E, Ekroth I, Lundqvist P, Melinder Å, Palm B, Rohlin P, 2003
“Refrigeration Engineering” Department of Energy Technology, KTH,
Stockholm, Sweden.
Iwmac (2009), “Centralised operation and surveillance, by use of WEB
technology”, available at: http://www.iwmac.no/english/
Minea V, 2006 “Improved Supermarket Refrigeration and Heat Recovery
System” ASHRAE Transaction: Vol.112, Part 2, pp. 592-596
Minea V, 2007 “Energy Efficiency of a Supermarket Refrigeration/Heat
Recovery System with Secondary Fluids” International Congress of
Refrigeration, 2007, Beijing, China
Johansson S, 2009 “Evaluation CO2 supermarket refrigeration systems”, MSc
Thesis, Department of Energy Technology, KTH, Stockholm, Sweden
Reindl D. T, Jekel T.B, 2007 “Heat Recovery in Industrial Refrigeration”
ASHRAE Journal August 2007
Wulfinghoff D. R, 1999 “Energy Efficiency Manual” Energy Institute Press,
Maryland, USA
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