1. technology and computing

ENERGY EFFICIENCY STUDY WALK IN COOLER Evaporator
Reihl Efficient, llc Cool-Vap technology
FINAL REPORT “WHITE PAPER”
REIHL COOL VAP THERMAL STORAGE CONTROL
EFFICIENCY TESTING
Prepared by:
Total Energy Solutions, LLC in concert with managing partners of Eco-Refrigeration Supply
4/29/2013
1
DISCLAIMER
TES advises end-user and the utility/sub-contract entity of no incurred liability for or through discovery of the findings
rendered subject in this sumbmittal to further energy measurement valuation for copressor equipment, by others. We,
the consultant agrees, to the fullest extent permitted by law to indemnify and hold harmless the client, it’s officers,
directors and employees against all damages, liabilities or costs, including reasonable attorneys’ fees, to the extent
caused by the consultant’s negligent performance of professional services under this study and that to it’s sub-consultants
or anyone for whom the consultant is legally liable. The client agrees, to the fullest extent permitted by law to indemnify
and hold harmless the consultant, it’s officers, directors and employees against all damages, liabilities or costs, including
reasonable attorneys’ fees and defense costs, to the extent caused by the clients’ negligent learning or lax performance
thereof, in receipt of professional education by the consultant under this agreement and that to it’s sub-clients or anyone
for whom the client is legally liable.
Neither the consultant nor the client or responsible party shall be obligated to indemnify the other party in any manner
whatsoever for the included indemnified party’s or constituent’s own negligence.
ACKNOWLEDGMENT
We would like to thank Mr. Mike Reihl, Eco Refrigeration and CMU Business Development Groups and
partnerships involved for the available funding to test this product and provide the “White Paper” research
contents for Reihl Efficient CoolVap Technology. At no time does TES assume responsibility for product claims
as independent third party evaluation auditor for energy savings application of Reihl/Eco-Ref. Products.
2
TABLE OF CONTENTS
Executive Summary .........................................................................................................................4
Technology Description....................................................................................................................5
Introduction .....................................................................................................................................6
Section 1 Test Design and Setup .....................................................................................................7-9
Section 2 Data Acquisition …………………. .........................................................................................10
Section 3 Refrigeration Cycle Analysis ..........................................................................................11-16
Section 4 Cooler Refrigeration Components..................................................................................17-20
Section 5 Measured Test Data……………………..................................................................................21-30
Section 6 Energy Consumption Improvement...............................................................................31
Conclusions ....................................................................................................................................32
3
EXECUTIVE SUMMARY
This project tests the performance of a thermal storage concept that is providing new technologies that are
commercially available that can reduce the power and energy requirements of :
Scenario 1:
Baseline case – a conventional set up, with ECM Fan Modules on continuously…temperature
Scenario 2:
Additional Energy Savings monitored beyond the baseline during conventional setup, with Emerson Reset Controller added
to the installation and refrigeration system optimized for CoolVap installation.
The evaluation attempts to establish a correlation between the tested pre-install and post-install situations for energy power
consumption that best represents the energy savings of the technologies, with compressor power representing major power
consumption of the walk-in cooler system.
The energy testing for pre- and post- monthly installation is established to prove beyond reasonable doubt, the energy savings
application of associated Cool-Vap Technology. The product’s advantage in energy savings of the Thermal Storage concept,
is applicable to walk-in cooler by minimizing the cooling “Demand” that the CoolVap Evaporator Coil requires from compressor
energy. Compressor energy savings results from less compressor starts per hour, minimizing the energy consumption
associated with hours the compressor must run on average. (A reset controller has been established in the Reihl Efficienct
product, that provides ECM Fan Motor reset while maintaining Cooler Temperature, with sensor positioned inside thermal
storage media….resulting in additional compressor holdoff time. The system control operation was not optimized to fully
capture this efficiency effect at this time.)
Energy Savings achieved by CoolVap Evaporator box exchange alone.
The energy consumption savings has been established only for modification of the Evaporator Box being installed and
system recharged with refrigerant, at this time. The energy savings achieved for Turk Lake party store by only upgrading to
the CoolVap Evaporator box as measured over relatively long term was much greater than expected considering further
system control optimization was not conducted. The energy savings achieved were 37% for simple
evaporator upgrade at Turk Lake.
(Further performance gains and cost savings are expected with full system optimization to fully capture CoolVap device
system operating efficiency potential at Turk Lake, with further tuning activities planned for May 8, 2013…Further data
recording will commence at that time.)
Energy Savings achieved for the second location were 75% for simple evaporator upgrade at
the second location, a 132% improvement.
4
Technology description
This report and testing of Reihl Efficienc CoolVap Thermal Flywheel evaporator coil effect on compressor energy
consumption readings, for a utility measure application provides sample pre- and post- energy consumption, and appropriate
calculated differences in each similar time-period of client Walk-In Cooler usage for the Retail Food Industry.
Reihl Patent Details:
Michigan Inventor Develops Patent for Method and Apparatus for Controlling Temperature of a Temperature Maintenance
Storage Unit ALEXANDRIA, Va., Sept. 4 -- Michael Reihl, Rosebush, Mich., has developed a patent (8,250,881) for a
"method and apparatus for controlling temperature of a temperature maintenance storage unit." The abstract of the patent
published by the U.S. Patent and Trademark Office states: "A temperature maintenance storage unit includes a fan assembly,
an evaporator coil assembly, a compressor unit, and an interior within which goods may be stored, with the fan assembly
configured to selectively generate air currents over the evaporator coil assembly and within the interior. The temperature
maintenance storage unit further includes an insulation pack assembly comprising a sealed generally flexible insulation pack
containing a temperature control fluid, an insulation thermostat and an interior thermostat, with the insulation thermostat
monitoring the temperature of the insulation pack assembly and the interior thermostat monitoring the temperature of the
interior. The insulation pack assembly being mounted directly to the evaporator coil assembly with the insulation thermostat
selectively activating the fan assembly in response to monitored temperatures of the insulation pack assembly, and with the
interior thermostat selectively activating the compressor unit in response to monitored temperatures of the interior."
5
INTRODUCTION
Refrigeration accounts for roughly 47% of total electric energy usage in convenience stores , according to an independent
consultancy. Energy Star reports that refrigeration energy represents 38% of total electric energy usage, compared to
combined energy usage of 23% consumption by building HVAC operational requirements. Reihl Efficient, LLC is targeting
the strongest energy consumption element in the operation of Retail Foods Industry, linked with the Petroleum distribution
C-stores in the MidWest.
Beverage cooler system electrical energy usage by components includes lighting, evaporator box fans, display door
anti-sweat heater controls, and refrigeration compressor.
The greatest energy consuming component is the refrigeration
compressor. The goal is proof by measurement of refrigeration compressor energy consumption, with the Reihl Efficient
CoolVap product acheives reduction of energy consumption through reduction of compessor run time achieved by patented
Thermal Flywheel mass storage strategy. Energy testing performed focuses on enhanced evaporator boxes with Thermal
Flywheel technology to reduce electrical consumption, compared to standard evaporator boxes without Thermal Flywheel
technology.
Two energy efficient operational implementations for reducing the electric energy consumption were intended in this test.
-
Baseline scenario measures the energy consumption differential accumulated for the effect of Thermal Flywheel on
the compressor using existing head pressure and controls. The pre- and post- power measurement data, included
for each model scenario will indicate energy savings in the accumulated power over time required of the
compressor within the refrigeration cycle.
-
Additional Energy Savings monitored beyond the baseline during conventional setup, with Emerson Reset
Controller added to the installation and refrigeration system optimized for CoolVap installation.
(Further testing will be required to conclusively establish further savings of additional energy controls operation under
optimized system conditions.)
Measured compressor energy consumption data represents testing of Compressor Energy before and after CoolVap
evaporator component installation in the cooler system. System modifications were performed with minimal control setting
changes possible, to maintain as consistent results possible within a real world C-store. Temperature and humidity sensors
were installed at strategic points of the refrigeration sytem to monitor proper cooling performance before and after the Riehl
Efficient CoolVap evaporator product.
6
SECTION 1
TEST DESIGN AND SETUP
1.1 Turk Lake Location
Test Design
The most important aspect of this project was to study the effect of energy efficient solutions available
for evaporator coil products on the power usage, cooling load, and merchandizing capability of a retail food
walk-in display case.
Cooler system test platform
Real world “party store” walk-in cooler chosen by Eco-Refrigeration was Turk Lake Party Store located in Montcalm, MI.
Initial observation revealed 2 separate compressors: 3hp 220v single phase compressor, and 1.5hp 220V single phase
compressor.
The 3hp compressor was chosen for electrical measurement due to extreme short cycle operation of the
1.5hp compressor..indicating an issue. Variations caused by product loading pull down and variable door openings, causing
both sensible and latent heat load variation were averaged by running compressor energy data logging over extended
period of multiple days.
Each installed compressor was separately connected to its own evaporator box.
Upon initial visit, it was noted there was a total of 38,000BTU/hr of evaporator cooling capacity installed in the
cooler. After CoolVap installation, the total cooling capacity was reduced to 28,000BTU/hr.
Cooler Temperature/Humidty Parameter sensing
Multiple temperature/humidity sensors were installed to confirm cooler performance characteristics before and after
installation of CoolVap Evaporator boxes. Temperature sensors were “named” by programming numerical identifiers into the
sensor internal memory (and inscribed on exterior) Sensor 1 through Sensor 5.
Exterior sensors: placed outside the cooler loading door to sense the internal building ambient temperature which would
conduct through insulated cooler walls, and through air infiltration when cooler doors were opened and closed.
Interior sensors: Placed to read near/mid/far zones of cooler relative
Sensor placement at mid height was chosen in event of any thermal stratification, although evaporator box placement and
airflow seemed to give relatively uniform temperatures when measured at various vertical points with a fast reading
thermocouple. Mid height would represent average product temperature under steady state thermal conditions if any
stratification did exist.
Temperature/humidity sensors located throughout the cooler system as follows
Cooler Exterior
-
Sensor 1
In compressor room area
Sensor 2
Outside surface of cooler wall (store ambient)
Cooler interior
-
Sensor 3
Inlet air stream of evaporator connected to compressor not electrically measured, high height (close to
loading door)
-
Sensor 4
Approximately middle length of cooler, mid height
Sensor 5
End length of cooler mid height (furthest from loading door)
7
1.1 cont.
Cooler/Building Temperature Operation
Cooler temperature maintained at a fixed dry bulb temperature of 40○F while relative humidity variation was minimized by
performing tests over a short enough weather interval, that humidity is very close for before and after CoolVap
installation test scenarios. All other operational parameters including suction and discharge pressures remained
unchanged as practical for real world test scenario.
Store Ambient Temperatures were controlled by a setback thermostat which maintained the same temperature settings
during the test intervals.
68F daytime/62F night time set points for thermostat.
8
1.2 Second Location (Undisclosed for owner privacy)
Test Design
The most important aspect of this project was to study the effect of energy efficient solutions available
for evaporator coil products on the power usage, cooling load, and merchandizing capability of a retail food
walk-in display case.
Second location cooler type is a standard walkin cooler with one loading door and no glass reach in
doors. This is type cooler is more representative of
that found in a restaurant or fast food store.
Cooler system test platform
Near Mt Pleasant, MI.
Initial observation revealed compressor: 1 hp 220v single phase compressor,
Cooler Temperature/Humidty Parameter sensing
Multiple temperature/humidity sensors were installed to confirm cooler performance characteristics before and after
installation of CoolVap Evaporator boxes. Temperature sensors were “named” by programming numerical identifiers into the
sensor internal memory (and inscribed on exterior) Sensor 1 through Sensor 5.
Temperature/humidity sensors located throughout the cooler system as follows:
Cooler Exterior
- Sensor 5 In compressor vicinity outside cooler
Cooler interior
- Sensor 2 Evap box outlet air stream
- Sensor 3 upper right corner of cooler looking in loading door, ceiling height
- Sensor 4 lower left corner of cooler looking in loading door, floor height
- Sensor 1 Inlet air stream of evaporator
Cooler/Building Temperature Operation
Cooler temperature maintained at a fixed dry bulb temperature of 40○F while relative humidity variation was minimized by
performing tests over a short enough weather interval, that humidity is very close for before and after CoolVap
installation test scenarios. All other operational parameters including suction and discharge pressures remained
unchanged as practical.
Building Ambient Temperatures were controlled by a thermostat which maintained the same temperature settings during the
test intervals.
Special efforts were taken to control humidity and recover from opening of large overhead door when
product delivered to the building.
Cooler lower door seal was really not functional, so cooler performance was
Susceptible to humidity in building air.
9
SECTION 2 - DATA ACQUISITION
2.1
Data Collection Procedure
The Hioki Model 3197 Power Quality Analyzer also functions as a data acquisition system capable of computing energy
draw, through post processing of the measured voltage and current measurements taken during the data
Logging.The data were then downloaded and detailed calculations were performed. The collected data point
from the five-minute intervals were averaged into hourly values where necessary and used for further
screening of the results. The advantage of using hourly averages is that the data trends can still be
displayed with an acceptable resolution while enabling the engineering model to generate relevant
calculated hourly results, such as cooling load setpoin variance of cooler and total quantity of power use.
2.2 Electrical Data Acquisition
A Hioki Instruments Model 3197 Power Quality Analyzer was used to log the test data.
The voltage and current conditions of both legs of 220V power feed to the 3hp compressor were monitored and recorded
during the multiple day tests. Every 5 minutes, the data acquisition system measured the voltage and current data, then
recorded into test instrument data buffer.
2.3 Temperature/Humidity Data Acquisition
Temperature/humidity data was collected on 2 minute basis as programmed into the remote programmable data
loggers, as temperature and humidity are both important factors to monitor for system performance state….and
affects energy consumed during system operation. Upon completion of testing phase, data was downloaded from 5
Temp/humidity loggers
Loggers used during testing sequence : LASCAR Electronics Model EL-USB-2
.
10
SECTION 3
REFRIGERATION CYCLE ANALYSIS
Using US Cooler Energy Data from 2010 and the most recent compressor information from Copeland
refrigeration data, a series of calculations were performed to obtain the key parameters including
the refrigeration load of the display case. Once the data were downloaded from the data logger and the
data of interest were extracted, preliminary reductions and calculations were performed. These
calculations included averaging of temperature, pressure, refrigerant mass flow, and discharge air
velocity.
3.1
General cooler consumption (US COOLER DATA)
Power data for US Cooler is listed below based upon 2010 average readings in kw/sqft/mo rolling averages.
Cooler Average Cost per month
Size
Cost/mo
Kwh/mo
6×6
$67.50
661.76
6×8
$67.50
661.76
8×8
$120.70
1183.33
8×10 $113.84
1116.00
8×12 $113.84
1116.00
10×10 $144.15
1413.23
10×12 $144.15
1413.23
Note: The above figures are estimates in a controlled environment; your exact numbers will vary.
• *These numbers were figured using the 12-month rolling average of $0.1022 kilowatt hour cost. According to the Energy
Information Administration this is the average cost in the United States for commercial electricity.
• This chart was created using several assumptions that can affect your actual operating cost.
1. The type of insulation in the walk-in.
2. Efficiency of the refrigeration system.
3. Inside and outside temperature of walk-in.
4. Where the walk-in is located.
5. The temperature and the weight of the product entering the walk-in.
6. How often the door is opened.
11
3.2
Refrigeration Cycle Analysis
The following analysis represents an refrigeration analysis conducted by Southern California Edison(So Cal Edison),
analyzing removal of heat from the cool space of a cooler similar to that at Turk Lake. The refrigerant at Turk Lk was
R409a….which will have slightly different thermodynamic coefficients, but the same basic procedure would be followed
for R409a refrigerant as R404a refrigerant. This is included to illustrate the complex system interactions in refrigeration
systems and give a general understanding of how compressor energy consumption is affected and energy is removed
from the cool space.
So Cal Edison analysis begins
~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~
Using refrigeration data, a series of calculations were performed to obtain the key parameters including
the refrigeration load of the display case. Once the data were downloaded from the data logger and the
data of interest were extracted, preliminary reductions and calculations were performed. These
calculations included averaging of temperature, pressure, refrigerant mass flow, and discharge air
velocity.
The total cooling load of the display case can be determined based on the refrigeration effect and mass
flow rate of refrigerant. Determination of refrigeration effect and other quantities, such as heat rejected at
the condenser, heat of compression, and sub-cooling quantities, depend on the refrigerant enthalpies at
specific locations within the refrigerant lines. Enthalpies can be either obtained from the refrigerant
manufacturer’s data at various temperatures and pressures, or calculated with respect to specific heat
capacities and temperatures. In this analysis, some of the enthalpies were obtained from the NIST’s
Refrigeration Property (RefProp) Program, version 6.0, and some by calculation.
Once the temperatures and pressures were obtained, the enthalpies were determined. DuPont’s Suva
Refrigerant Expert Program, version 2.0 was used to determine the saturated refrigerant temperatures.
The NIST RefProp program was utilized to determine superheated refrigerant enthalpies. The data logger
provided all pressures in pounds per square inch of gage units, and after conversion to absolute units, the
NIST RefProp program was used to obtain the enthalpies.
The enthalpies in the saturated phase were calculated using temperature-dependent expressions provided
by DuPont, as well as using basic thermodynamic relationships. Equation 3-1, provided by DuPont, was
used to determine the saturated enthalpies (in kJ/kg) of refrigerant 404A for a temperature range of -20°C
to 40°C. The temperatures of the saturated liquid were first converted to Celsius, and then inserted into
Equation 3-1 to obtain the corresponding saturated enthalpies. The enthalpy of saturated liquid in the
condenser was found using Equations 3-1 and 3-2.
H = A + BT + CT2
(3-1)
H
= Enthalpy, (kJ/kg)
A
= 200
B
= 1.438333
C
= 0.003916667
T
= Temperature, (°C)
where A, B, and C were constants, determined by DuPont, from the relationship between saturated
enthalpy and temperature. Next, Equation 3-2 was used to convert the enthalpy in kJ/kg to Btu/lb.
Because of a change in reference states from SI (Metric) to English units, a reference conversion factor, H
(ref), was included in Equation 3-2.
12
H (Btu/lb) = [H (kJ/kg) - H (ref)] • 0.43021(Btu/lb / kJ/kg)
H (ref)
(3-2)
= 145.6 kJ/kg for R404A
The Refrigerant Expert program and the NIST RefProp do not provide sub-cooled enthalpies; therefore,
in order to determine the enthalpies for sub-cooled state (i.e., at the expansion valve inlet and at the outlet
of the condenser), the thermodynamic relationship between enthalpy and temperature was incorporated.
3.3
Enthalpy at the Expansion Valve Inlet and Condenser Outlet
In order to incorporate the thermodynamic relationship between enthalpy and temperature, the correct
liquid specific heat capacity was needed. The equation for determining the heat capacity was provided by
DuPont, and it establishes the specific heat capacity of liquid refrigerant 404A for a temperature range of
-40°F to 140°F. The net effect of system sub-cooling can be captured by comparing saturated liquid
enthalpy and the final liquid enthalpy. The same approach was employed to determine the liquid line
enthalpy at the inlet of the expansion valve (Equation 3-3) and at the outlet of the condenser (Equation 34).
Cp expan-in = 0.306 + 4.083E-4 T expan-in - 1.194E-6 T2 expan-in + 8.056E-8 T3 expan-in
(3-3)
Cp expan-in = Liquid Heat Capacity (based on temperature at the inlet of expansion valve), (Btu/lb °F)
T expan-in = Average temperature of the sub-cooled liquid at the inlet of expansion valve (value directly
extracted from the data acquisition system), (°F)
Cp cond-out = 0.306 + 4.083E-4 T cond-out - 1.194E-6 T2 cond-out + 8.056E-8 T3 cond-out
(3-4)
Cp cond-out = Liquid Heat Capacity (based on temperature at the outlet of the condenser), (Btu/lb °F)
T cond-out
= Average temperature of the sub-cooled liquid at the outlet of the condenser (value directly
extracted from the data acquisition system), (°F)
The temperature difference between the saturated liquid in the condenser and the sub-cooled liquid was
needed in order to find the corresponding enthalpy change of refrigerant due to sub-cooling. Equations 35 and 3-6 were used to determine the temperature difference between the saturated liquid in the condenser
and sub-cooled liquid entering the expansion valve and leaving the condenser, respectively. The
temperature of saturated liquid in the condenser (SCT in °F) was determined by using DuPont’s Suva
Refrigeration Expert Program based on discharge pressure (pressure at compressor outlet) data from the
data acquisition system.
∆T expan-in = SCT – T expan-in
(3-5)
∆T expan-in = Temperature difference between SCT and sub-cooled liquid entering the expansion valve,
(°F)
∆T cond-out = SCT – T cond-out
(3-6)
∆T cond-out = Temperature difference between SCT and sub-cooled liquid at the outlet of the condenser,
(°F)
13
Next, the enthalpy change between the sub-cooled liquid and the saturated liquid was calculated by
utilizing their thermodynamic relationship, ∆H = Cp • ∆T. Equations 3-7 and 3-8 were used to determine
this enthalpy change at the inlet of the expansion valve and outlet of the condenser, respectively.
∆H expan-in = Cp expan-in • ∆T expan-in (3-7)
∆H expan-in = The enthalpy change between sub-cooled liquid entering expansion valve and saturated liquid in condenser, (Btu/lb)
∆H cond-out = Cp cond-out • ∆T cond-out (3-8)
∆H cond-out = The enthalpy change between sub-cooled liquid leaving condenser and saturated liquid in condenser, (Btu/lb)
Finally, the enthalpy of the sub-cooled liquid was computed by subtracting the enthalpy change between the sub-cooled and
saturated liquid from the enthalpy of the saturated liquid. Equations 3-9 and 3-10 were used to determine the enthalpy of the
sub-cooled refrigerant at the inlet of the expansion valve and at the outlet of the condenser, respectively.
H expan-in = H satliq - ∆H expan-in (3-9)
H expan-in = The sub-cooled liquid refrigerant enthalpy at the expansion valve inlet, (Btu/lb)
H satliq
= Saturated liquid enthalpy (value determined using Equations 3-1 and 3-2), (Btu/lb)
H cond-out = H satliq - ∆H cond-out (3-10)
H cond-out = The sub-cooled liquid refrigerant enthalpy at the outlet of the condenser, (Btu/lb)
3.4
Refrigeration Effect
The refrigeration effect is the quantity of heat that each unit of mass of refrigerant absorbs to cool the
refrigerated space. It simply represents the capacity of the evaporator per pound of refrigerant. This
quantity was derived by subtracting the refrigerant enthalpy at the evaporator inlet (before the expansion
valve) from the slightly superheated refrigerant enthalpy at the outlet of the evaporator (Equation 3-11).
RE = Hevap-out - Hexpan-in
(3-11)
RE = Refrigeration effect of the refrigerant in the evaporator, (Btu/lb)
Hevap-out = Superheated refrigerant enthalpy at the evaporator exit (value determined by using the NISTprogram), (Btu/lb)
14
3.5
Refrigeration Load
The refrigeration load of the case is the rate of cooling or heat removal (in Btu/hr) that takes place at the
evaporator of the display case (Equation 3-12). This quantity is obtained by multiplying the refrigeration
effect by refrigerant mass flow rate, which is extracted from the data acquisition system. The total case
load was determined by using Equation 3-12.
Q case = MFref • RE • k
(3-12)
Q case
= Total refrigeration load of the case, sensible and latent, (Btu/hr)
MF ref
= Mass flow rate of refrigerant, (lb/min)
k
= Conversion factor, (60 min/hr)
It is sometimes useful to determine the refrigeration load in tons. Thus, the refrigeration load of the case
can be divided by 12,000, a conversion factor for Btu/hr to tons (Equation 3-13).
Q case (tons) = Qcase / 12000 (3-13)
Q case (tons) =
3.6
Refrigeration load, (tons)
Compressor Normalized Power
All tests were conducted using a single 6.5 HP compressor equipped with a variable speed drive to
accommodate the fluctuating capacity requirements under varying part-load operations. Under different
cooling load conditions, the compressor operated at efficiencies different than its peak design.
Additionally, the inverter losses of the variable speed drive further obscured the true power demand of the
compressor under different load conditions. Using a normalized compressor power value can provide
more clear comparisons between different test conditions. The use of average compressor efficiency for
these tests can eliminate the effects of part-load compressor and variable speed drive inefficiencies.
Hence, obtaining the heat of compression became a critical point of interest. The heat of compression
was determined based on the monitored mass flow rate of refrigerant and the difference between the
enthalpies of refrigerant at the compressor inlet and discharge (Equation 3-14).
Q comp = MFref • k • (Hcomp-out - Hcomp-in) (3-14)
Q comp
k
= Heat of compression, (Btu/hr)
= Conversion factor, 60, (min/hr)
Hcomp-out
= Superheated enthalpy at the outlet of the compressor (value determined by using the NIST program),(Btu/lb)
Hcomp-in = Superheated enthalpy at the inlet to the compressor (value determined by using the NIST Program), (Btu/lb)
15
The compressor power was, therefore, normalized based on average compressor efficiency. An estimation approach was utilized to
determine the average compressor efficiency. The actual amount of power supplied to the compressor was recorded by the data
acquisition system. These data were used to calculate the actual compressor efficiency (Equation 3-15).
η = Qcomp / (kWcomp • k)
η
(3-15)
= Actual compressor efficiency represented for each test scenario
kWcomp = Average power supplied to the compressor for each test scenario (recorded by the dataacquisition system), (kW)
k
= Conversion factor, (3,413 Btu/hr/kW)
Once representative compressor efficiency was determined for each test condition, the values were then averaged to obtain an
overall average efficiency (Equation 3-16).
Η overall = Ση / n
(3-16)
η overall = Overall average efficiency of compressor
n
= Number of test conditions
After determining the overall average efficiency of the compressor, the compressor power was normalizedusing Equation 3-17.
kWnorm-comp = Qcomp / (ηoverall • k)
(3-17)
kWnorm-comp = Normalized compressor power, (kW)
k
= Conversion factor, (3,413 Btu/hr/kW)
3.7
Total Quantity of Rejected Heat
As the name implies, the total quantity of rejected heat is the rate of heat rejection (in Btu/hr) that takes
place at the condenser. This total quantity of rejected heat comprises the heat absorbed from the
refrigerated space (cooling load) and the heat of compression in the compressor. Once the enthalpy of the
refrigerant at the inlet and outlet of the condenser was determined, the total quantity of rejected heat can
be obtained by utilizing Equation 3-18.
Q rejected-heat = MFref • (Hcond-in - Hcond-out) • k (3-18)
Q rejected-heat = Total quantity of heat rejected at the condenser, sensible and latent, (Btu/hr)
H cond-in = Superheated refrigerant enthalpy at the inlet of the condenser (value determined by using the NIST program), (Btu/lb)
k
= Conversion factor, (60 min/hr)
~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~
End of So Cal Edison Analysis
16
Section 4
Cooler refrigeration components
4.1 Monitored Compressor Characteristics (Turk Lk)
Tecumseh Product Co. AGA4534AXN AG155ET-017-J7 compressor R12 208/230V 1ph 3 HP
Type
Reciprocating
Application
HBP - High Back Pressure
Refrigerant
R12
Voltage/Frequency 208-230V ~ 60Hz 200-220V ~50Hz
Product Specifications
Performance
Condition
ASHRAE
Test Voltage
230V ~ 60HZ
Refrigeration Capacity
Btu/h
33700
kcal/h
8492
W
9874
Input Power
W
3600
Efficiency
Btu/Wh
9.3600
kcal/Wh
2.3600
W/W
2.7400
Evaporator Temperature
7.2°C (45°F)
Condensing Temperature
54°C (130°F)
Ambient Temperature
35°C (95°F)
Return Gas Temperature
35°C (95°F)
Liquid Temperature
46°C (115°F)
General
Evaporating Temperature Range -6.7°C to 12.8°C (20°F to 55°F)
Motor Torque
High Start Torque (HST)
Compressor Cooling
Fan
Mechanical
Displacement (cc)
100.6710
Oil Type
Synthetic Alkylate
Viscosity (cSt)
53
Oil Charge (cc)
1955
Sound Power dB(A)
0
17
Monitored compressor Characteristics (cont.)
Electrical
Voltage Range (50 Hz)
180-242
Voltage Range (60 Hz)
187-254
Locked Rotor Amps (LRA)
0
Rated Load Amps (RLA 50 Hz)
Rated Load Amps (RLA 60 Hz)
17.0000
Maximum Continuous Current (MCC in Amps)
28.0000
Motor Resistance (Ohm) - Main
0.6870
Motor Resistance (Ohm) - Start
2.5200
Motor Type
CSR
Overload Type
INTERNAL
Relay Type
Potential Relay
Agency Approval
CSA Listed, UL Recognized
4.2
Original Evaporator Boxes
Turk Lk
4 Fan Evaporator :
Snyder General Model LHC-270-1
Retrofitted with 4 ECM motors
2 Fan Evaporator:
Krack Corp Model SK-28-104A
Retrofitted with 2 ECM motors
4.3
Improved Replacement Evaporator Boxes
Turk Lk
RE17K ( connected to monitored compressor)
RE11K
18
4.4 Monitored Cooler Components (Second location)
Compressor Characteristics
Tecumseh Product Co. AJA7480ZXDXC compressor R12 208/230V 1ph 1.5 HP
R404A Medium Temp Condensing Unit
Length 24.3"
Height 13.4"
Suction Conn. 5/8" SAE
Max. Fuse Size 20
Liquid Conn. 3/8" SAE
BtuH Capacity @ 20°F Evap.
Pressure Control Low
Receiver Capacity Lbs. 6.0
BtuH Capacity @ 30°F Evap.
HP 1
Wt. Lbs. 106
BtuH Capacity @ -10°F Evap.
Volts
208/230-1
B/M 2B1102-9
Width 17.8"
BtuH Capacity @ 10°F Evap.
BtuH Capacity @ 0°F Evap.
Compressor # AJA7494ZXD
Min. Circuit Amps 12.6
Shipping Weight 106.00
Shipping Width
29.00
Shipping Length
29.50
Shipping Height
14.00
7980
9,520
4330
6,720
5,500
19
4.4 cont.
Replacement evap box, Standard current technology
Bohn Standard Model: ADT070AK
General Information
Category
Unit Coolers
Type
Walk-In Unit Coolers
Style
Low Profile
Voltage
115/1/60
Defrost Type
Air Defrost
FPI
6
Rating Point °F
10°F TD 25°F SST
Capacity (BTUH)
7,000
CFM
1,460
Motor Type
PSC Motor
No. of Fans
2
Motor HP
1/15
Motor Voltage
115/1/60
Motor Watts
164
Motor Amps
2.0
Height (in.)
15
Depth (in.)
14 7/8
Length (in.)
45 1/2
Approx. Net Weight (lbs.)
45
Coil Inlet OD
1/2
Coil Inlet ODF
-
Suction ID
7/8
External Equalizer OD
1/4
Drain MPT
3/4
Side Port OD
-
Hot Gas Pan Conns.OD
-
Performance Data
Electrical Ratings
Physical Data
Connections Data (in.)
20
Section 5
5.1
Measured Test Data
Test data Baseline condition (Original Evaporator Box)
Turk Lk location
Electrical – Measured Compressor Energy
Compressor disabled to
change evap box
Energy 44.373kWh
Time 64 hours
Cooler temp pull down
starts with CoolVap Evap
3hp Compressor ENERGY SLOPE = 44.373kWh/64hrs= 0.6933kWh/hr (2365.734 BTU/hr)
21
Cooler temp pull down ends
with CoolVap Evap Box
5.1 Electrical – Measured Compressor Energy (Cont.)
Turk Lk location
Electrical Demand (Original Evaporator Box)
Measured interval
22
5.1 cont.
Turk Lk location
Temperature/Humidity test data (original Evaporator box)
Measured interval
23
5.2
Test data Improved condition (CoolVap Evaporator Box)
Turk Lk location
Electrical – Measured Compressor Energy
3hp Compressor Electrical energy
3hp Compressor Electrical energy
consumption over 140hour interval
inflection point
3hp Compressor ENERGY SLOPE = 67.073kWh/140hrs= 0.4764 kWh/hr (1634.511 BTU/hr)
24
5.2 Electrical – Measured Compressor Energy (Cont.)
Turk Lk location
Electrical Demand (CoolVap Evaporator Box)
Inflection point in electrical demand
more visible than energy curve.
25
5.2 cont.
Turk Lk location
Temperature/Humidity data (CoolVap Evaporator Box)
26
5.3
Test data (both Evaporator boxes shown on same plot)
Second location
Electrical use Measured of Cooler System Energy (including fans)
Note : The cooler system at second location was wired to allow monitoring of fans and compressor combined. This
energy is different than that measured at Turk Lake party store. Adding evaporator fans consumption to compressor
consumption…is a more complete picture of cooler energy consumption.
Data recording continued during the
installation of second evaporator box, therefore both results are contained in same data trace.
Initial test install with
CoolVap evap box
Energy use = 14.359kWh
Initial test CoolVap
evap box end point
Second test install with
Heatcraft evap box
Energy use = 70.948kWh
CoolVap Evap Box
Cooler System ENERGY SLOPE = 14.359 kWh/94hrs= 0.1527kWh/hr (or 521.3752 BTU/hr)
Heatcraft Evap Box
Cooler System ENERGY SLOPE = 70.948 kWh/120hrs= 0.5912kWh/hr (or 2017.371 BTU/hr)
27
5.3 cont.
Test data Improved condition (CoolVap Evaporator Box)
Second location – Electrical Demand
Cooler System Demand
With CoolVap evaporator box
28
5.3 cont.
Cooler System Demand
With Heatcraft evaporator box
Observation – Cooler System demand significantly lower with CoolVap evaporator box. Indicating
compressor motor is drawing less power, working less hard.
lowering kW demand on compressor.
29
Refrigeration system presents less load, thus
5.3 cont.
Second location
Temperature/Humidity data
CoolVap evaporator box
Heatcraft evaporator box
30
Section 6
Energy consumption improvement
Turk Lake location
Electrical compressor energy consumption improvement
Standard value = 2365 BTU/hr
,
Improved value = 1625 BTU/hr
Simple Ratio = 1625/2365 = 0.6871
(used 68.71% of energy with CoolVap installed)
Energy consumption improvement = 2365 – 1625 = 740 BTU/hr
Percentage improvement = (initial-improved)/((initial+improved)/2)
= (2365 - 1625)/((2365 +1625)/2) =
37% energy improvement
Second Location
Electrical compressor energy consumption improvement
Initial value = 2017.371BTU/hr
,
Improved value = 521.375 BTU/hr
Simple Ratio = 521.375/2017.317 = 0.2588
(used 25.88% of energy with CoolVap installed)
Energy consumption improvement = 2017.371– 521.375 = 1496
BTU/hr
(0.43843 kWh/hr)
Percentage improvement = (initial-improved)/((initial+improved)/2)
= (2017.371- 521.375)/(( 2017.371 + 521.375)/2) =
31
132% improvement
CONCLUSION
The demonstrated performance of the Reihl Efficient Evaporator FanCoil, during these tests shows there are
new options to save significant energy of refrigeration systems operational energy use.
The use of this new energy efficient technology in these projects has reduced the electric consumption of the
refrigeration’s system by a value of 37% on one installation and over 72% on second installation.
This represents a significant savings of energy which adds to the profit margin of any commercial business
with which operates refrigerated coolers, or refrigerated rooms.
The additional clearly observed benefit from the second location temperature/humidity data, Reihl Efficient
CoolVap product better manages humidity removal from the cooler air. This is tangible benefit in normal cooler
operations, since humidity removal consumes energy which could otherwise be used for cooling food or
beverage products. When a cooler is loaded with 100+ cases of beverage, CoolVap should be able to cool the
product more quickly, since more of the refrigeration capacity is available to lower the temperature of product
load due to lower cooler humidity….instead of being spent lowering humidity.
Reihl Efficiency, LLC and it’s flagship CoolVap product offers a valuable enhancement to Walk-In Cooler
systems, both on new cooler construction and retrofit to existing coolers to significantly reduce cooler energy .
32