NATIONAL BUREAU OF STANDARDS REPORT

DEVELOPMENT OF A METHOD FOR TESTING AND RATING

THE COOLING LOAD OF REFRIGERATED TRUCK BODIES

by

C, W. Phillips and R. W. Penney

Report to

Transportation and Facilities Research Division

Agricultural Research Service

U, S. Department of Agriculture

U. S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS
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NATIONAL BUREAU OF STANDARDS REPORT

NBS PROJECT NBS REPORT

42103-40-4212260 9378

August 18, 1966

DEVELOPMENT OF A METHOD FOR TESTING AND RATING

THE COOLING LOAD OF REFRIGERATED TRUCK BODIES

by

C„ W. Phillips

Environmental Engineering Section

Building Research Division, IAT

and

R» W. Penney

Transportation and Facilities Research Division

Agricultural Research Service

U. S„ Department of Agriculture

Report to

Transportation and Facilities Research Division

Agricultural Research Service

U. S» Department of Agriculture

IMPORTANT NOTICE NATIONAL BUREAU OF STA r

for use within the Government. B and review. For this reason, the

| whole or in part, is not authorizi Bureau of Standards, Washington

the Report has been specifically pi

Approved for public release by the

Director of the National Institute of

Standards and Technology (NIST) on October 9, 2015

accounting documents intended

ibjected to additional evaluation

isting of this Report, either in Office of the Director, National

le Government agency for which

ies for its own use.

II. S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS

DEVELOPMENT OF A METHOD FOR TESTING AND RATING

THE COOLING LOAD OF REFRIGERATED TRUCK BODIES

by

C. W. Phillips and R. W. Penney

1 . Acknowledgement

The authors wish to acknowledge the assistance of John W. Grimes

and Donald M. Kupfer of the Environmental Engineering Section in

conducting the tests and preparing the material for this report.

2 . Introduction

Lack of a standard method for determining the cooling load of

refrigerated truck bodies has made it difficult, if not impossible,

to effectively rate or select these vehicles, so widely used in

the transportation and distribution of perishable foods and other

commodities

Following the development, through Government-industry cooperation,

of a standard method for measuring the cooling load of refrigerated

trailers, persons and organizations, from industry and Government,

similarly concerned with the manufacture, selection, application,

and utilization of refrigerated trucks, established a project to

develop a suitable method for rating these vehicles (see Project

Organization)

It is important to recognize that measurement of the cooling load

of a refrigerated enclosure (such as a truck, trailer, railway car,

or warehouse) does not provide a direct indication of the total

cooling capacity required of the refrigerating apparatus for the

enclosure. Additional application factors such as loading and

unloading techniques, cargo temperatures, door opening schedules,

and required temperature pull-down rates all influence service

loads and the sizing of the equipment, some perhaps to a greater

degree than the rated cooling load of the insulated enclosure.

This report presents the results of cooling load tests, with and

without simulated solar heating, and certain other tests of five

standard or prototype insulated truck bodies of various types

and sizes available for commercial refrigerated service. Based

on these test results, a recommended method is presented for

measuring the cooling load and weight gain of refrigerated trucks.

The method is applicable also to smaller trailers and containers

of size and usage comparable to the refrigerated trucks.
3 . Purpose

The principal objective of this project was the development of

a procedure or technique suitable for use as a standard method

for rating the cooling load of insulated truck bodies used for

the transportation of perishable and frozen food or other commodities

requiring refrigeration. The desired method was to take into account

the following:

(1) The cooling load of the vehicle under standard conditions.

(2) The rate of weight gain due to moisture accumulation

caused by air leakage under standard conditions.

(3) The increase in cooling load due to solar heating.

4. Background

In an earlier project, a rating method for refrigerated trailers

was developed to evaluate the cooling load, air leakage, and moisture

gain of those vehicles intended for long-haul highway transportation

of frozen foods [ 1 ]

.

Certain differences were noted between usage of highway trailers

and local delivery trucks, which indicated that different test

procedures should be used in rating the two types of vehicles.

In general, the highway trailers operate day and night, usually

at road speeds, and are opened for loading and unloading only at

terminals or major transfer points. The additional cooling load

caused by solar radiation while the vehicle is stationary is

possibly less than that caused by air leakage when the vehicle

is in motion. Doors are opened infrequently, if at all, during

a trip.

Delivery trucks are stationary for the most part, operate primarily

during daylight hours, and are subjected to a wide variety of door

opening schedules during the working period. Thus, the rating

method for trailers required means for simulating the effect of

forward motion, but not solar radiation or door opening. The

rating method for trucks, in contrast, requires means for providing the effect of solar radiation, but not necessarily the

effect of forward motion. It is recognized that there are ex- ceptions to the general case.

Of the radiant energy absorbed by a surface over insulation such

as an insulated truck wall, only a small percentage will actually

be transferred through the insulation, while a high percentage

will be radiated to surrounding lower temperatures or given up to

convective air movement at the surface. Bright metallic surfaces

2
absorb less solar energy than dark painted surfaces, but in

turn they are less able to re-radiate the energy absorbed.

Thus, differences in the temperatures attained by the surfaces

are less than those based solely upon the solar absorbtances of

the different surfaces. As each surface is heated above ambient

air temperature, it loses additional heat to convective air

currents in proportion to the difference in temperature between

the surface and the air.

For these tests it was desired to determine the effect of equal

solar heating, regardless of the finish of the test vehicle, or

the various types or methods of construction of the various truck

bodies in the test series. A low temperature wavelength radiation source was used for the simulated solar tests so that all

vehicles tested would be heated approximately as if they were painted with dark paint and exposed to solar radiation of equal

intensity. It was not an intended purpose of these tests to determine what type of surface treatment was the best for minimizing

the heat gain due to solar exposure.

Initially, it was the intent of this project to include the

effects of door openings on (a) air and moisture exchange, (b)

the increase of cooling load. Much work was done on these subjects; however, it now appears that these effects are related

more to the "service load" of the vehicle than to a rating method

for the insulated body itself. The effects of air and moisture

exchange during door openings, with the attendant increase in

cooling load, will be covered in a report to be published at a later date.

5 . Project Organization

The Environmental Engineering Section, NBS; Transportation and

Facilities Research Division, USDA; and the Truck Body and Equipment Association, beginning July 1, 1960, sponsored a cooperative

effort to develop a suitable method for rating refrigerated trucks

that would take into account the factors outlined (see Background).

The project was carried out at NBS under the direction of the

Environmental Engineering Section. A Project Steering Committee

was appointed, consisting of representatives from interested

industry and Government organizations. The initial Project

Steering Committee meeting was held on October 14, 1960. Paul R.

Achenbach, of the National Bureau of Standards, was elected and

served as chairman. Harry R. McGee, formerly of the Truck Body

and Equipment Association, served as secretary until succeeded by

Robert F. Guilfoy, of the U. S. Department of Agriculture.

3
It was a recommendation of the Steering Committee that the proposed rating technique be as consistent as possible with that

used for rating trailers [1], Thus, a test facility could,

without major difficulty, rate either trucks or trailers.

Based on this, the decision was made to maintain the 0°F interior

temperature and the 100°F, 50% RH ambient temperature and

humidity conditions used to rate trailers.

Of the five test vehicles used, some were prototype and some production models (see Description of Test Vehicles). They

were furnished by the following cooperating manufacturers;

Murphy Body Works, Inc., Wilson, N. C.

Boyer town Auto Body Works, Boyer town, Pa.

Hackney Bros. Body Company, Wilson, N. C.

The Heil Company, Milwaukee, Wisconsin

Divco Truck Division, Divco-Wayne Corporation, Detroit, Mich.

The first of the five vehicles was placed in the test chamber in

April 1961.

6 . Test Systems

To measure the cooling load of the trucks tested in this project

a metered heat sink with comparison heater was used [2j. This

apparatus was similar to that used in the rating method for trailers

and is similar to that recommended in the proposed rating method

for trucks. A measured mass flow rate method was used for the

independent simultaneous measurement.

Alternative methods considered for the proposed rating method

were: use of liquid nitrogen (or other expendable refrigerant)

expanded into the vehicle under test, transient-state cooling load

tests, and reverse heat loss tests. None of these was selected for

various reasons. Reverse heat loss tests do not account for the

effect of moisture deposit in the insulated spaces; transientstate cooling load tests are better suited to homogeneous material

construction and might not properly account for the effect of moisture deposit in the insulated cavities. The use of direct discharge

of liquid nitrogen into a truck body was tried [3]. Control by use of thermocouple-operated solenoid valves was satisfactory.

Direct weighing of the nitrogen tank gave demand and load values.

One serious drawback for rating purposes was found; the slight pressure in the truck interior needed to discharge the expanded nitrogen gas to the outside tended to offset the air leakage rate of

the vehicle. In other words, it favored a truck with high air

leakage rates. If a suitable metered heat exchange method can be

found to allow indirect cooling by the liquid and gaseous nitrogen,

it can then be evaluated for possible advantages in cost and simplicity for rating purposes. For all alternative methods, an independent simultaneous measurement would be required.

4
6 . 1 The Metered Heat Sink Principle

The basic metered heat sink apparatus with the comparison heater i

shown in Figure 1. An important feature of this method, outlined

in the following derivation, is that measurement of the mass flow

rate, M, and the specific heat, Cp, of the refrigerant (brine) is

not required to determine the heat flow rate, q t , into the truck

body under test. The heat flow rate into the truck, sensible and

latent, is represented by

q. = c M At. - h

t p t where c is the specific heat of the brine, Btu/lb deg F

P M is the mass flow rate of the brine, lb/hr

At is the brine temperature change in the air

cooling coil, deg F

h is the total electric power input to the truck

(fans, heaters, etc.), Btu/hr.

The electrical input to the brine in the comparison heater is

represented by

q = c M At

x p c

where c^ is the specific heat of the brine, Btu/lb deg F

M is the mass flow rate of the brine, lb/hr

At

c is the brine temperature change in the comparison

heater

Combining the two equations,

q + h qc

C

P

M =

At

t = At^

&t

t . q t = qc— - h

At

c Note that both M and c are taken to be constant in the two equations combined; M will be constant at any given time with a closed

liquid brine system. The change in Cp of methylene chloride brine

between the mean coil temperature and the mean comparison heater

temperature is less than 0.5 percent for the operating ranges used

and recommended.

In the basic metered heat sink method with the comparison heater,

a cooling coil refrigerated by an adjustable brine flow is placed

in the body under test, and the comparison heater is placed in

- 5

-
the brine line from this cooling coil. The refrigerating equipment for cooling the brine is outside of the test chamber. In

a variation of this apparatus, based on a design by M. Altieri r4~|, a complete water-cooled refrigerating unit is placed in the body

under test. A comparison heater could be placed in the water

line leaving the refrigerating unit to provide an independent

simultaneous measurement. The proposed rating method (see

Appendix II) recommends the basic comparison heater method using

brine and includes requirements for the necessary instrumentation

and operating procedure to provide the required accuracy for

rating purposes. An independent simultaneous check is provided

by the measurement of mass flow rate by flowmeter.

6 . 2 Air Leakage Determination Methods

In the course of this project, the air leakage rate of the trucks

was studied from two aspects. One was the air leakage, under

standard test conditions, as calculated from the observed weight

gain rate, and the other was the measured air leakage from the

cargo space by (1) a helium trace katharometer measurement under

standard test conditions [5], and by (2) a static pressure test

wherein the cargo space was pressurized under isothermal conditions.

The air leakage as determined from the weight gain is a function

of both the air movement between the truck interior and exterior and

the air movement into and out of the insulated walls of the

vehicle. This method provides a basis for calculating the cooling

load due to air leakage. It is determined from the weight gain

by assuming that the air enters the vehicle at ambient conditions,

and leaves saturated at the interior temperature (see Air Leakage

Calculations). This yields a minimum air leakage rate and corresponding latent and sensible cooling loads.

The static pressure air leakage test and the helium trace air leakage

test each measure an air flow rate which is principally influenced

by the resistance to air movement between the cargo space and the

exterior. Neither of these tests provides a basis for determining

the cooling load due to air leakage under standard test conditions.

In these tests, a truck with a well-sealed interior skin and a poorly sealed exterior skin (contrary to good design practice)

might show an air leakage rate no higher than a truck with a wellsealed exterior skin, which would have a much lower latent cooling

load

.

A degree of correlation does exist between static pressure test

results and cooling loads in that a truck with a high air leakage

rate will no doubt have a high air leakage cooling load, regardless

of the mode of entry and the inability to calculate that load.

6
FIGURE

7 . Description of Test Facilities

The ambient test chamber used was 25 ft long, 17 ft wide, and

12 ft high. Two large doors at one end permitted entry of the

trucks. The ambient test conditions were 100°F temperature and

507o RH, and provisions for conditioning the ambient air were

provided by two fan coil units discharging into a common plenum

above a perforated ceiling. Air return was direct into each

air handler. Refrigeration machinery, steam regulators, and

controls were located external to the test room. Figure 2 shows a view of the test chamber.

The interior temperature of each test vehicle was maintained

at 0°F (35°F for one test to represent medium temperature applications) by a refrigeration system consisting of a packaged-unit

brine cooler, a turbine-type brine pump, an air cooling coil,

and the associated piping and valves. The brine cooling unit was capable of two-speed operation for better control of cooling

capacity over the wide range required. The primary refrigerant

was R22 (monoch lorodif luoromethane ) . The secondary refrigerant,

or brine, was R30 (methylene chloride). At the mean coil temperature used in these tests, the specific heat of the brine was approximately 0.273 Btu/lb deg F . The turbine type pump, driven by

a 1/2-hp motor, circulated the brine through the secondary re- frigerant circuit. The pump was located in the return line,

just before the chiller. The rated capacity of the pump was

8 gpm (water) against a 40 psi head.

An air cooling assembly located in the truck body contained a coil, blower, damper, and electric heaters. The coil was selected

with only four fins per inch to permit extended operation without

defrosting. Three electric resistance heaters were mounted

between the coil and the blower. These were for control purposes.

In figure 3, the air cooler assembly is opened to show the coil

and resistance heaters. At the coil inlet, a damper was installed

as a means of controlling air flow. At the blower outlet, a baffle distributed the chilled air to the various parts of the truck

body interior. Figure 4 shows the brine chiller, brine pump,

and the housing (at bottom right) of the comparison heater used

in the metered heat sink apparatus. Figure 1 shows schematically

the cooling load calorimeter or metered heat sink system as applied

for truck body testing. The cooling coil inside the truck body

was defrosted automatically in a short time (<10 min.). This

defrost system circulated brine through the cooling coil and the

comparison heater (which provided heat for the defrosting operation),

while by-passing the chiller. The brine remaining in the chiller

was thus sub-cooled by the primary refrigerant system, which con- tinued to operate at low speed. The sub-cooled brine in the

chiller was circulated through the cooling coil at the end of

the defrost period to restore the test temperature in the truck

rapidly.

7
The brine circuit was heavily insulated and vapor sealed in

order to eliminate frost and minimize heat gain. To prevent

any vertical forces from being exerted on the truck which would

interfere with the proper weighing of the vehicle, two flexible

lines were installed horizontally near the point where the brine

pipes enter the truck body. The thermocouple wells for the

brine lines at the truck were located so as to respond to the

temperature of the brine immediately before entering and after

leaving the truck. The comparison heater consisted of a special

piping configuration enclosed in a plywood box about 40 by 20 by

20 in. (see Fig. 5). The box was insulated with loose cork

fill. Thermocouple wells were fitted into the piping near the

inlet and outlet. Between the temperature measuring wells, an electric resistance heater was mounted inside the pipe. The

thermocouple wells for the comparison heater were located so that the brine would be well mixed at the points of measurement.

Piping within the box was installed so that vapor would not

collect around points of measurement. A surface thermostat,

located on the pipe containing the heater element, protected

the heater against damage due to overheating. The insulated

box was also vapor sealed so that moisture would not accumulate

in the insulation. Energy input to the brine by the comparison

heater was measured by a watt-hour meter on the control panel.

An electronic integrating flow meter with a volume-measuring sensing

element was located in the brine line between the coil outlet and

the comparison heater. The weight rate of flow of the brine

is the product of the volumetric rate and the specific gravity

of the brine. The weight rate of flow provides one factor for

determining a comparative value for the refrigeration load at

the coil.

The temperature measuring system used for these tests employed

copper constantan thermocouples, a 16-point electronic recording

potentiometer, an electronic indicating potentiometer and a precision-grade laboratory potentiometer. The charts from the

recording potentiometer were used to determine that steady-state

conditions were maintained throughout the measurement period. The

temperatures used to calculate cooling loads were read on the

electronic indicating potentiometer and/or the laboratory potentiometer. The latter was used to obtain a greater degree of accuracy in determining the critical temperature differences between

the brine entering and leaving the comparison heater and the air

cooling coil. These temperature differences were sensed by

calibrated five-junction copper-constantan thermopiles, which

were inserted in the thermocouple wells in the brine lines.

Individual thermocouples were also placed in the wells to monitor

the thermopiles and to determine mean brine temperatures.

- 8 -
The interior temperature in the truck and the ambient temperatures

of the test room were sensed both by parallel-connected (for

average temperatures) and individual thermocouples suspended 6

in. from the interior and exterior corners. Other thermocouples were placed to measure various temperatures of interest, such

as those of the interior and exterior truck surfaces, primary

refrigerant temperature, and those of the air entering and leaving

the cooling coil assembly.

It was necessary to weigh the defrost water from the cooling

coil to determine total moisture gain and the portion of the

total gain which accumulated in the insulation space. To accomplish this, a rubber hose connected to the defrost pan of the

cooling coil emptied into a container in the test room outside

of the truck. An electric cord heater in the defrost pan and

rubber hose prevented freezing during the defrost operation.

A water trap on the end of the hose prevented outside air from

being drawn through it into the truck interior.

Three separate platform scales were used to weigh the truck to

determine the weight gain caused by moisture picked up from air

leakage. Each rear wheel rested on a scale. The front axle was supported on a third scale. These scales had a capacity of 6000

lb each, and were sensitive to differences of 1/2 lb at the loading

used. Figure 6 shows a truck installed on the scales.

7.1 Solar Simulation Apparatus

Previous observations of unmoving trucks and trailers exposed to

bright summer sunlight at NBS had indicated skin temperatures

approaching 70°F above ambient, and a value between 65 and 70°F was arbitrarily selected an an approximate upper roof skin temperature limit for the simulated solar tests. It was postulated that

the maximum amount of solar energy would enter an insulated

vehicle when it was parked with its longitudinal axis in a northsouth direction so that both longer sides and the roof would be

irradiated in the course of a cloudless day.

The idea of a rotating mechanism to duplicate the effect of the

daily sun movement over a vehicle was discarded in favor of

adjustable stationary banks of electric heating elements and

parabolic reflectors facing each of two sides and the top of a test vehicle [6]. To simulate the effect of sun movement during

the day, the voltage applied to the heaters was varied. Each

bank was 15 ft long by 9 ft wide and consisted of 45 parabolic

reflectors, 1 by 3 ft at the bank face (see Fig. 2). Each

reflector was equipped with an 81-ohm helical electric resistance

coil wound around a 7-mm O.D. glass tube mounted in the reflector

9
so that the heater axis was at the focus of the parabolic re- flector. The heating elements were designed to operate without

visible radiation at 208 volts. Maximum heat dissipation of

each heater, including both convective losses and radiant transfer, was about 600 Btu/hr f

t

2 of proj ected reflector areas „ The

total heat release of one bank of 45 heaters operating at maximum design voltage was approximately 83,000 Btu/hr. Figure 7 shows one of the test trucks in the test room for simulated

solar load tests.

7.2 Static Pressure Air Leakage Apparatus

The static pressure air leakage apparatus consisted of a centrifugal blower connected to the truck interior, with an orifice

in the connecting 2-in„ pipe line. The pressure drop across the

orifice was used to calculate the air flow rate. A manometer

was used to measure the pressure difference between the interior

and the exterior of the truck. The orifice measurements were

checked by pitot tube sweeps across the inlet air pipe in the

truck

„

7 o 3 Helium. Trace Katharometer Apparatus

The helium trace katharometer apparatus used to check the air

leakage rate was a version of the helium trace apparatus described

in reference 5 modified to permit "batch" sampling of the airhelium mixture in the truck. The modified katharometer was used

both for the steady-state tests in this report and for the subsequent tests of air exchange through door openings which will be

covered in a later report. The details of the modified katharometer

will be described fully in that report.

8. Description of Test Vehicles

8. 1 Test Vehicle 1

Truck body 1 was a walk-in ice cream delivery design with exterior

dimensions of 183 in. long by 91 in. wide by 87 in. high. A

single door at the rear was the only access, except for a pass

door on the right side forward of the wheel well. Internal

volume was 544 cu ft.

The metal exterior, including the roof, was painted a gloss white,

the interior metal walls were smooth except for anchors to which

refrigerated plates could be attached on the ceiling and forward

wall

.

10
Insulation consisted of 7 in. of expanded polystyrene plus 2 in. of glass fiber in the roof, 7 in. of expanded polystyrene

in the walls, and 6 in. of expanded polystyrene plus 1 in. of

cork in the floor.

Figure 8 shows vehicle 1.

8.2 Test Vehicle 2

Truck body 2 was an integral cab design having a door entering

from the cab, a door on the curb side, and a single rear door.

Cargo body external dimensions were 176-1/2 in. long by 80-1/2

in.- wide by 75-1/2 in. high. Internal volume was about 388 cu ft.

The exterior surface was painted a dark color of a non-glare finish. Interior surfaces had corrugations on the walls, with a sheet metal floor and ceiling.

The insulation in the walls and roof was fiber glass (6 in. and

8 in. thick, respectively) and there was 7 in. of expanded polystyrene in the floor.

Figure 9 shows vehicle 2.

8.3 Test Vehicle 3

Truck body 3 was a reach-in design with three doors on each side.

The rear opened to a storage compartment used for supplies, which

was not refrigerated and which did not communicate with the larger

space. Exterior dimensions were 182 in. long by 90-1/2 in. wide

by 88 in. high. The metal exterior surface was painted a gloss

white. The internal refrigerated volume was 432 cu ft.

Insulation consisted of 6 in. of urethane in the roof, 4 in. of

urethane in the walls, and 3 in. of urethane plus 1 in. of cork

in the floor. The urethane was foamed in place. Spacer strips were provided on the metal interior walls.

Figure 10 shows vehicle 3.

8.4 Test Vehicle 4

Truck body 4 was constructed of urethane slabs resin bonded to

multi-layer glass laminate for both interior and exterior surfaces.

The bonding agent penetrated the glass laminate, attaching it to

11
the insulation and forming a hard smooth surface after curing

in the mold 0 Sides, ends, and roof were formed in a mold as one subassembly and the floor as another. These parts were

then assembled and resin-bonded with glass laminate overlapping

at all joints. Insulation thickness was 2 in. in walls and

3 in. in floor and roof. The color inside and out was white.

Overall exterior dimensions were 140 in. long by 83 in. wide by

63 in. high. Internal volume was 333 cu ft.

A 22-1/2 in. wide by 44-1/2 in. high door was located at the rear

of each side.

Figure 11 shows vehicle 4.

8.5 Test Vehicle 5

Truck body 5 was an integral cab design. Entry was by double door

at the rear or by a sliding door from the driver’s compartment.

The metal exterior surface was painted a flat grey. The metal

interior lining had corrugated sheets on the roof, floor, and side

walls

„

A combination of glass fiber and preformed urethane board insulations was used as follows: rear wall, 3 in. urethane and 1 in.

glass fiber; front wall, 2 in. urethane and 1 in. glass fiber;

side walls and floor, 3 in. urethane; and roof, 3 in. urethane

and 1 in. glass fiber. The glass fiber was used to fill between

stiffening and structural members.

The internal volume of the body was 183 cu ft and its exterior

dimensions were 92 in. long (not including integral cab) by 79

in. wide by 68 in. high.

Figure 12 shows vehicle 5.

9 . Laboratory Test Procedures

9 . 1 Steady-State Cooling Load and Weight Gain

After each truck was placed on the scales in the test chamber,

the brine lines, electric cables, and thermocouples were connected. The truck interior was then refrigerated to 0°F temperature (35 °F for one medium temperature test), and the ambient

temperature and humidity were controlled at 1QQ*T and 50% RH«

When controlled conditions were obtained, the scales were balanced

By this time, the change in air density in the truck had taken

place, and no longer affected the readings.

12
Air leakage rates during steady-state cooling load tests were measured by means of the modified helium-trace katharometer

For a steady-state heat balance, the instruments were read

at half-hour intervals until a period of test was obtained

having not less than 6 hours of uninterrupted test conditions

with essentially constant heat gain. Each test was repeated

one or more times, with the vehicle maintained under test con- ditions for several days.

Weight gain readings were made during the entire test period

for each truck with adjustment made for water removed from the

cooling coil, if defrosting was required. Defrosting, if re- quired, was done at the end of a period of data observation,

allowing nearly 16 hours for recovery of steady-state conditions

prior to the next period of data observation.

9 . 2 Solar Load Test Procedure

The test of a vehicle under simulated solar heat load was similar to that of steady-state, with the interior temperature

set at 0°F, and the ambient air temperature and humidity held

at 100°F and 50% RH. Simulated solar heating was imposed by

the banks of electric heaters. The measured temperature of one of the heater rods operating at the maximum voltage necessary for test was 700°F. At this temperature the wavelength of

maximum emission is about 4.5 microns. As the voltage is lowered,

the temperature of the heater approaches the ambient temperature

of 100°F, at which temperature the wavelength of maximum emission

is about 9.3 microns. The significance of selecting a radiation

source at temperatures of 700°F or lower is related to the

capacity of various surfaces to absorb radiant energy. Surfaces

covered with nonmetallic paint, such as most truck bodies, re- gardless of color, will absorb about the same fraction of radiation at the wavelengths above 4 microns as dark-painted surfaces

will absorb at the shorter wavelengths of solar radiation. For

this reason it was considered unnecessary to repaint any of the

painted vehicles in order to establish similar conditions for

the simulated solar tests. Metallic surfaces such as aluminum

or stainless steel, or metallic-painted surfaces would have

required special consideration. All the vehicles received for

the test series were painted with nonmetallic paints.

To establish a pattern for cyclic variation of the simulated

solar energy values of hourly insolation on a vertical east

and west surface, and on a horizontal (roof) surface were taken

13
from U. S. Weather Bureau curves for June 21 at latitude 40°N.

These are shown in Figure 13. Such a solar day extends from

4:40 a.m. to 7:20 p.m., with maximum insolation values of 235

Btu/hr ft2 on a vertical east wall at 7:30 a.m., and on a vertical west wall at 4:30 p.m.; and 320 Btu/hr ft3 on a horizontal

surface at noon.

The voltage on the bank of heaters over the roof was adjusted to

produce a temperature rise approaching 70 degrees above the ambient

temperature of 10Q°F, thus simulating solar irradiance at noon. With this value as a maximum, the voltage on the roof heater at

other times of the day was adjusted to produce a power dissipation

proportional to the height of the curve in Figure 13 for the

roof. The voltage on the other two banks was adjusted independently

to provide power dissipations proportional to the heights of the

other two curves in Figure 13. The same cycle of power input was used with all vehicles.

Seven tests were run on one vehicle (body E) to determine the

effect of solar heat gain on the 4-hour maximum cooling load (average

cooling load rate over the 4-hour period of maximum load) when:

(1) The ambient temperature was held constant.

(2) The ambient temperature was varied sinusoidally to follow

simulated daily cycle with diurnal range of 20 deg F.

The seven tests run were:

Ambient

Solar

simulation

Conditions Mean temp.

°F

Temp, range

°F

1. Steady 100 No

2. Steady 100 -- Yes

3. Variable 100 90 to 110 No

4. Variable 100 90 to 110 Yes

5. Variable 90 80 to 100 Yes

6. Variable 90 80 to 100 No

7. Steady 90 -- No

Tests 2 and 1 were run to compare the observed cooling load rates

under steady 100°F ambient temperature with and without a solar

cycle. Test 7 was made with the ambient temperature held steady

at 90°F without a solar cycle for comparison with test 1 at 1Q0“F

ambient. Tests 4 and 5 were conducted with a solar cycle, and

with the ambient temperature varied 20 deg F sinusoidally over

a 24-hour period to yield daily mean temperatures of 100 “F and

90®F, respectively. In Tests 3 and 6, the same ambient temperature

variations were repeated, but the solar cycle was omitted. For

all seven tests, the interior temperature was held at 0

ttF and

the ambient humidity at 50% RH.

14 -
HOURLY

INSOLATION

ON

CLOUDLESS

DAYS

JUNE

21,

LATITUDE

40°

NORTH

00

(3inNllAI) 13 0S/ma‘N0l±V10SNI

SOLAR

TIME

FIGURE

13

9 .3 Static Pressure Air Leakage Test

In the static pressure air leakage tests, the interior of the

trucks were pressurized to approximately 0.10 in. of water above

atmosphere. The tests were made under isothermal conditions with

both interior and exterior of the truck at room temperature.

When the air flow necessary to maintain the desired pressure

across the truck body walls was established, the pressure drop across the orifice in the 2-in. pipe was observed. This flow

was verified by use of a pitot tube, as described under test apparatus. A test was made with the doors unsealed, and with the doors

sealed with tape.

10. Test Results and Discussion

10.1 Steady-State Tests

The results of steady-state cooling load tests on the five vehicles

are shown in Table 1.

Table 1. Cooling load due to air leakage and transmission at 100°F

ambient temperature & 50% R.H. and 0°F interior temperature 3

Cooling load. Btu/hr

Truck Total b Transmission Air leakage

A 3700 2530 1170

B(0°) 3200 2180 1020

B(35°) 2150 1510 640

C 2550 > 2500 < 50

D 1800 1730 70

E 1850 1800 50

a Note exception

b Values rounded

of 35°F interior temperature

to nearest 50 Btu/hr.

run for truck B.

Cooling loads ranged from 1800 to 3700 Btu/hr . The measured steady

state total cooling load includes the transmission cooling load and

the air leakage cooling load. No simulated solar heating was used

in the steady state tests.

15
Comparison of the cooling loads determined by the metered heat

sink method (primary method) and flow meter method (secondary,

or check, method) for one of the test vehicles during a series

of special test runs is shown in Figure 14.

The gain in weight of the vehicles during the tests was attributed to moisture deposition and was used to calculate

the air leakage cooling load* Air leakage cooling loads ranged

from 32 percent to less than 2 percent of the total cooling

load of the truck bodies tested,

10*2 Weight Gain and Air Leakage

The rate of weight gain in pounds per hour, and the rate of

air leakage in cubic feet per minute, for the five vehicles are given in Table 2*

Table 2. Air leakage rate and weight gain rate at standard conditions

Test Weight gain Air leakage 3 Air leakage^1 vehicle lb/hr cubic ft /min cubic ft/min

A 0.47 5.8 4.7

B(0°F) 0.41 5.0 3.8

B(35®F) 0.29 4.2 - C 0.014 0.2 - D 0.028 0.3 - E 0.022 0.3 0.39

a Calculated from weight gain*

b By helium-trace katharometer

»

Air leakage computed from measured weight gain of each truck, on

the basis of ambient air entering the body, depositing moisture,

and leaving saturated at 0®F temperature is shown for all five

test vehicles. Measured air leakage rates for three of the test

vehicles using the modified helium-trace katharometer are also

shown in Table 2. The deviation of these values from the calculated air leakage rates based on weight gain was approximately

207o and for trucks A and B, was consistent with the concept that the weight gain air leakage accounts for all air

exchange, not just that which occurs to and from the cargo space.

For truck E, the magnitude of the air leakage rates is believed

to be too small to permit effective comparison of the two measurements *

- 16 -
4000 ro cvj ~ anoH aad rua ‘avoi oniiooo xonai

DAYS

OF

OPERATION

An air leakage test for selected static pressure difference

between the cargo space and the test room was run on four of the

five vehicles. Table 3 shows the air leakage rates for the

test vehicles with door sealed and unsealed.

Table 3. Air leakage under static pressure

Truck

Pressure across walls

Leakage

cubic feet

rate

per min

in. W. G. Doors

unsealed

Doors

sealed

A 0.07 105 102

B 0.11 92 41

C 0.10 8 < 8

D 0.10 2 -- E Not tested — —

Comparing the observed leakage rates for Trucks A and B with

doors sealed and unsealed shows that the excessive leakage of

Truck A was through the body, whereas for Truck B, the door

seals accounted for more than half of the leakage. The

contrast between the observed air leakage rates for trucks A

and B and those for Trucks C and D in Table 3 indicates a wide

range of effectiveness of various techniques for reducing

air leakage.

Smoke tests of some of the bodies revealed leakage at the doors

and door frame members, body seams, and door gaskets.

10.3 Simulated Solar Heating Tests

Figure 15 shows the exterior surface temperatures achieved in

a typical solar simulation test. Note the shape similarity to curves in Figure 13, the insolation curve. The maximum 4-hour

average observed cooling load was taken as the refrigeration

requirement under solar loading. The percentage increase of the

maximum 4-hour solar loads over the steady-state cooling loads are shown in Table 4.

17
Table 4. Percent increase in cooling load with

simulated solar heat over steady state

Truck Percent

A 24

B Not tested

C 20

D 19

E 25

For the four trucks tested, the average percent increase was 22.4 with individual values ranging from 19.4 to 25.4 percent.

10.4 Comparative Tests at Variable and Steady Conditions

Seven tests were run on truck body E to explore and determine

truck cooling loads with and without simulated solar heating

when 1) the ambient temperature was held constant at either

100°F or 90°F and 2) the ambient temperature was varied to

follow an approximately sinusoidal daily cycle having a range

of 20 deg F with its maximum usually occurring 3.5 hours

later than the solar noon and yielding mean ambient temperatures

of either 100°F or 90°F o The results are given in Table 5,

in which the cooling loads when not steady are the average loads over the 4-hour period of maximum load. Also shown in Table 5 are the differences or changes in cooling load made apparent

when particular tests are compared.

Test pairs a, b and c in Table 5 show that the simulated solar

cycle increased the cooling load about 450 Btu/hr, and that the

increase was independent of the ambient temperature level and of

the choice of a steady ambient temperature or the selected daily

sinusoidal cycle of ambient temperature. Test pair d shows that

reducing the steady ambient temperature from 100°F to 90°F (a

reduction of 10 percent in the temperature difference) reduced

the cooling load by 9 percent. Test pair e shows that the same reduction in average ambient temperature, but with a daily cyclic

variation of 20 degrees in each case, reduced the cooling load

about 11 percent. Test pair f shows that when the solar cycle

was superimposed on the variable ambient temperature cycle, reducing the average ambient temperature from 100°F to 90 6F decreased

the cooling load 9 percent. Test pairs g and h show that the

effect of the sinusoidally-varied ambient temperature was to

increase the cooling load to a 4-hour maximum average value

180 Btu/hr greater than the corresponding steady ambient value

at lOO^F, with or without a simulated solar cycle. A smaller in- crease (120 Btu/hr) was indicated by test pair i, at 90®F mean

ambient temperature.

18
SURFACE

TEMPERATURES

DURING

SIMULATED

SOLAR

TEST

OIOO

0300

0500

0700

0900

1100 N

1300

1500

1700

1900

2100

2300

SOLAR

TIME

Table 5. Cooling loads for truck body E with s teady and variable

ambient conditions, 0 ®F temperature: inside, and with

and without simulated solar heating

Test

comparisons

Test

no

.

Ambient

conditions

Mean

temperature

of ambient

Solar

cycle

Cooling load

Measured Difference

Btu Btu

Degrees F. per hour per hour

a 1 Steady 100 No 1,870

^450

2 Do 100 Yes 2,320

b 3 Variable 100 No 2,050 ^450

4 Do 100 Yes 2,500

c 5 Variable 90 Yes 2,280 •^460

6 Do 90 No 1,820

d 1 Steady 100 No 1,870

^170

7 Do 90 No 1,700

e 3 Variable 100 No 2,050

^230

6 Do 90 No 1,820

f 4 Variable 100 Yes 2,500 ^220

5 Do 90 Yes 2,280

g

1 Steady 100 No 1,870 hso

3 Variable 100 No 2,050 )

h 2 Steady 100 Yes 2,320 }l80

4 Var iable 100 Yes 2,500 )

i 7 Steady 90 No 1,700 ^120

6 Variable 90 No 1,820 )

j 1 Steady 100 No 1,870 ?410

5 Variable 90 Yes 2,280 )

1/ Steady = ambient temperature same throughout the day. Variable « ambient

temperature followed daily temperature cycle.

19
The cooling load in Test No. 1, with a steady ambient temperature

of 100°F, was 102.7 percent of that observed in Test No. 6 with

the maximum ambient temperature of 100®F during the daily cycle

but with an average ambient temperature during the maximum four

hours of 99.5°F, The temperature difference between the ambient

and the truck interior for Test No. 1 was 100 degrees, 100.5 percent of the temperature difference of 99.5 degrees for the maximum

four hours of Test No. 6. Comparing the ratios of cooling load

and temperature differences for these two tests indicates that

the heat capacity of the test truck body reduced the cooling load

about 2 percent during the maximum four hours of the variable

ambient test. This suggests that the steady-state test at 10Q aF

ambient temperature is an acceptable substitute for the more

complex variable ambient test procedure.

Test pair j indicates that Test No. 1, conducted with a steady

IQO^F ambient temperature and no solar cycle had a cooling load

410 Btu/hr less than the 4-hour maximum load of Test No, 5,

conducted with a sinusoidally-varied ambient with a 9Q eF mean

temperature and with a solar cycle. The latter condition is

thought to be reasonably realistic for the solar and ambient

exposure of an operating truck on a typical hot day, and the

resulting 4-hour maximum cooling load observed is probably a good first approximation to the maximum cooling load of truck

body E due to climatic factors alone. The actual in-use cooling

load would, of course, be increased by product loads and by the

additional load due to door openings which would depend on the

truck service.

Since there was moderately good agreement as to the increase

of cooling load due to a simulated solar cycle (approximately

22 percent) for all four trucks tested, as shown in Table 4,

it is suggested that a fair approximation to the maximum cooling

load of a truck due to climatic exposures alone can be estimated

by multiplying the cooling load obtained in a test at a steady

100°? ambient, with no solar cycle, by a factor of 1.22 (i.e.,

1 + 410/1870 for truck E). On this basis, the determination of

comparable 4-hour maximum cooling loads for trucks can be effected

by tests at steady 100°F ambient temperature, with no solar cycle,

conditions the same as those called for in the procedure

previously developed for the rating of refrigerated trailers.

20 -
Table 6 compares the solar test results with estimated values

obtained by multiplying the steady-state cooling load by 1.22

£ Table 6. Maximum cooling load with solar heating;

comparison of estimated vs. test results, Btu/hr

Truck Est imated^ By test Estimated

test

A 4510 4600 0.98

B 3900 Not tested __ C 3110 3070 1.01

D 2200 2150 1.02

E 2260 2320 0.97

a Four-hour average

b 1.22 x steady state cooling load in Table 1

The deviation from unity of the ratio of estimated maximum cooling

load to the maximum cooling load measured in test is no greater

than 0.03. While no solar simulation test was run on a vehicle

whose interior temperature was at 35°F, the data obtained in the

0°F interior temperature tests and a theoretical analysis of the

effect of a 35° rise in interior temperature (see Appendix I)

yields multiplication factors that can be applied and which

include probable solar effects. These factors are: For a truck

tested at 0°F interior temperature, and rated for 35°F interior

temperature, a multiplication factor of 0.87 times the measured

cooling load at 0°F. For a truck tested at and rated for 35°F

interior temperature, a multiplying factor of 1.34 times the

measured cooling load at 35°F. No other extrapolations are recommended

.

11 . Conclusions

There is a definite need to account for solar heat gain in any

truck rating program. The test data shows that the solar gain

is consistent enough to enable the use of a 100°F steady ambient

temperature test and a multiplying factor in lieu of either an increased ambient temperature or an actual solar simulation test.

The test results and theoretical analysis indicate that, for a truck

tested at and rated for 0°F interior temperature, a multiplier of

1„22 times the base cooling load can be used to incorporate the

solar effect. For a truck tested at 0°F but rated for 35°F

interior temperature, a multiplier of 0.87 times the base cooling

load can be used, and for a truck tested at and rated for 35 °F

interior temperature, the multiplier is 1.34 times the basic

cooling load at 35°F interior temperature.

21
It was the recommendation of the Project Steering Committee that

the truck rating method be as consistent with that used for

trilers T 1] as possible,, The ram air simulation plenum used on trailers would, of course, be omitted. By using a multiplying

factor for solar heat gain, and by having air-cooling coils of

different sizes to adapt to the wide range of cooling loads, a test facility used to rate trailers can rate trucks under this

recommended technique. The different sized coils are recommended

to improve accuracy by limiting the amount of "bucking® 1 , or control heat required to hold the desired truck interior air temperature. Because the percentage errors in measurement of the total

cooling load absorbed by the air-cooling coil (and matched by

the comparison heater) are directly reflected in the determination

of the cooling load of the truck itself, the greater the ratio

of the bucking heat to the total cooling load absorbed by the

air-cooling coil, the less accurate the measurement of the

cooling load of the truck. For this reason, it is recommended

that in no case should the bucking heat be allowed to exceed

50 percent of the total cooling load absorbed by the air-cooling

coil

.

Experience gained in the tests at the NBS also pointed up the

absolute necessity of very steady brine temperatures. Fluctuations

in the temperature of the brine leaving the refrigerating unit

carry through the entire system, and cause transients in measured

cooling load. For this reason, the maximum cyclic variation of

brine temperature leaving the refrigerating unit should not ex- ceed 0.4 deg F during the rating period, for results consistent

with the desired accuracy of the rating.

Some types of construction are tight enough so that air leakage

and attendant weight gain may be below the accuracy of the weighing

facilities and therefore be meaningless. For this reason, it Is

recommended in the proposed rating method that any observed weight

gain less than the sensitivity requirement of an appropriate scale

(or equivalent sensitivity of other weighing techniques), or less

than the combined sensitivity requirements of the scales (if more

than one is used), be shown in the data as being less than that

sensitivity requirement.

Even though defrosting of the air cooling coil is not to be

done during the final weight gain portion of the proposed ratingtest, collection and weighing of the condensate provides a means for determining the percentage of the observed total weight

gain that is accumulated on the coil. To improve the accurac3

r of this determination, the air cooler, particularly the drain

pan, should be designed to drain freely during the defrost operation.

22
It was found that static air pressure leakage tests, helium

trace air leakage tests, and smoke bomb tests, do not relate

directly to cooling loads. For this reason, it is not recom- mended that they be included in the proposed rating technique.

These tests are useful in determining the relative effectiveness of air sealing techniques, and might possibly pinpoint

major leakage problem areas, and for this reason might prove

useful, as in-house tests for a manufacturer.

It should be emphasized that the proposed rating technique

with the solar load multiplier does not give the total cooling

load of a refrigerated truck and therefore cannot be used

alone for determining the capacity required of the refrigerating

unit. The service load caused by door openings and possible cargo

cooling loads for some trucks in some types of service may be

several times greater than the rated cooling load. Laboratory

work has been done on measuring door opening cooling loads, and

will be described in a forthcoming NBS-USDA publication.

Information on cargo loads, as from a warm cargo or one which

respirates, is available in various Department of Agriculture

publications, and in the Guide and Data Book of the American

Society of Heating, Refrigerating, and Air Conditioning Engineers.

23
References

Phillips, C. W. , W. F. Goddard, P. R. Achenbach, H. D.

Johnson, and R. W. Penney, A Rating Method for Refrigerated

Trailer Bodies Hauling Perishable Foods. U. S. Dept, of

Agriculture, Agricultural Marketing Service, Marketing

Research Report No. 433_. Washington, D. C., 1960.

Achenbach, P. R. and C. W. Phillips, Heat Sink Method for

Measuring the Cooling Loads of Refrigerated Structures.

Proceedings Xth International Congress of Refrigeration,

International Institute of Refrigeration. 177, Boulevard

Malesherbes, Paris (17e) France, Part 2 p 272-276,

illus . , 1959 .

Achenbach, P.iR., C. W. Phillips, and R. W. Penney,

Measuring the Cooling Load of Refrigerated Vehicles by

Free Evaporation of Liquid Nitrogen. Supplement 1965-3

Bulletin of the International Institute of Regrigerat ion

.

177, Boulevard Malesherbes, Paris France, p. 115-124.

Altieri, Mario, A Method of Testing by Internal Cooling

the Overall Heat Transfer of Insulated Vehicles, International Institute of Refrigeration. Proceedings of

Commission 7 Meeting in Padua, Italy, 1961. 177, Boulevard

Malesherbes, Paris France.

Coblentz, Carl W. and P. R. Achenbach, Design and Performance

of a Portable Inf iltration Meter . American Society of Heating,

Refrigerating and Air-Conditioning Engineers, ASHRAE

Transactions, Volume 63, 1957.

Achenbach, P. R. and C. W. Phillips, Laboratory Study of

the Effect of Solar Radiation on Cooling Load of Stationary

Refrigerated Vehicles, International Institute of Refrigeration, Proceedings of Commission 7, Munich, Germany, 1963.

177, Boulevard Malesherbes , Paris France.
Appendix I Multiplier Factors for Extrapolation

from 0°F to 35°F Interior Temperature

An equation for the total heat transfer under a periodic con- dition into a refrigerated vehicle parked in the sun can be

expressed as follows: *

q* - U (Tern - t + E) ( 1 )

where: q

U

Tern

t

.

1

« _

E

= total heat transferred into the vehicle, Btu/hr

overall heat transfer coefficient of the vehicle,

Btu/hr deg F

mean sol-air temperature (see below), °F

interior temperature of vehicle, °F

heat-sink characteristic factor (see below), °F

.

The mean sol-air temperature is an effective temperature that

relates the various modes in which the outside skin of the

vehicle reacts to its thermal environment. It is expressed by:

Tem

where: Tern

t o

$

t , “F

o h

mean sol-air temperature, °F

exterior ambient temperature, °F

insolation to which the vehicle is subject,

Btu/hr fta

surface absorptance factor

overall heat transfer coefficient between outside

surface and ambient, Btu/hr deg F

Tem thus represents the energy level of the exterior of the

vehicle, and is dependent only on the exterior surface and the

ambient thermal environment, and is independent of the interior

temperature

.

The heat sink characteristic factor, E, incorporates the

resistance to change in the heat flow through the structure,

* J. L. Threlkeld, Thermal Environmental Engineering, pp . 372-374,

Prentice-Hall, 1962.

1-1
as when the increase in insolation increases the heat flow

into the vehicle. It is dependent on the heat storage and

transient heat transmission characteristics of the vehicle

,

are in turn dependent on its construction. Thus it can be seen

that for any given vehicle subject to a given ambient temper:

ture and insolation for a specified time or time period, tern and

E will be constant.

The steady-state heat transfer into a vehicle not subjected

to insolation can be expressed by

q = U(t - t^), Btu/hr, (2)

where U, t and t^ are as previously defined.

The results of the comparative tests for any one truck, run

at 0°F interior temperature and with and without simulated

solar irradiation, can be expressed in the following equation

q^ U (Tem + E - tj.)

57

= u(t

o

- tp <3)

in which the prime denotes values corresponding to the case

of solar insolation, and the subscript i on q denotes the

truck interior temperature.

In Table 4, observed values of the ratio q|/ q^ for the individual trucks are given, for t = 100°F ana t, = t ! = 0°F.

o i i

from which the value of (Tem + E) for each truck can be calculated, as follows:

Truck ^ 1 q

o Tem + E, °F

A 1.24 124

B Not tested - C 1 . 20 120

D 1.19 119

E 1.25 125

Average 122

Assuming that for t Q = 100°F (Tem + E) is substantially

independent of the truck interior temperature, t, , and using

all of the trucks the average value 122 SF, (3) becomes

122 - t!

q i 1 q i " ioo - i[ W
Using (4), it is possible to obtain immediately the multiplier

factors for various conditions:

qo

: 122-0 , _

100 - 0

q

o 1,22 q o (5)

q 35

122 - 35 __

100 - 0

q o 0,87 qo (6)

q 35

122 - 35 , 100 - 35 q 35 1,34 q 35 (7)

Use of the factors given in (5), (6), and ( 1 ) will, when used

with the standard method of determining q, give values of q'

which may be used as ratings for comparing vehicles under

conditions of insolation.

1-3
Appendix II

Recommended Standard Method for Testing and Rating

the Cooling Load of Refrigerated Truck Bodies

1.0 Purpose

The purpose of this standard is to describe methods of testing and

rating refrigerated truck bodies with respect to cooling load under

selected standard interior and ambient conditions with adjustment

for insolation effect, and with respect to weight gain rates.

1.1 Scope

This standard applies to refrigerated truck bodies (or containers)

used for transporting frozen food or other materials requiring refrigeration. It describes a laboratory technique for measuring the

cooling load under assumed typical ambient operating conditions,

and at interior temperatures of 0®F and 35 °F. The distinction

"truck” is taken to mean a vehicle operating primarily on short haul

delivery routes, and thus the standard takes into account the influ- ence of solar loading, but does not assume ram air pressure on the

front of the vehicle such as highway use would impose. Also omitted

is the effect of door openings. Although significant, these fall

under a service load category, and should not be included in a basic truck body rating technique.

The test method described can also be used to measure the cooling

load and weight gain rate of a truck body at any time during its operating

life to evaluate changes in performance.

2 . 0 Basis for Rating

2 . 1 Ratings

Results to be determined from the rating test shall consist of the

cooling load in Btu/hr, and weight gain rate in lb/hr, all under specified conditions.

The average of two methods of simultaneously determining the cooling

load shall be used for rating. Results of the two methods must

agree within 5 percent.

Because the ratings will be in error if measurements are improperly

made, or if conditions are not properly maintained, all instruments

and readings shall meet the accuracy requirements of this standard.

II-l
2 . 2 Standard Rating Conditions

Tests to determine Standard Ratings of all truck bodies shall be

measured under one or both of the following standard rating

conditions (low or medium interior temperatures):

2.2.1 Cooling Load and Weight Gain Tests

Test room ambient air temperature

Dry-bulb 100. 0°F

Wet-bulb (at standard barometer)* 83.5°F

Truck interior air temperature

Dry-bulb: Low temperature

Medium temperature

0 . 0°F

35 . 0°F

2.2.2 Deviations

Deviations allowed in test conditions from Standard Rating

Conditions

Reading

Maximum deviation

of arithmetical

average of all

readings from

standard conditions

Maximum

deviation

of individual

readings

Test Room Air

Dry-bulb

Temperature

±1 . 0 deg F ±2.0 deg F

Wet-bulb ±1 . 0 deg F ±2.0 deg F

Truck Interior

Dry-bulb

Air Temperature

±0.5 deg F ±1 . 0 deg F

2 . 3 Standard Ratings

There shall be three allowable ratings that can be published

under this standard. The first is the basic rating of 0°F

interior temperature, 100°F and 50% RH ambient temperature and

humidity. The second rating is for the same ambient conditions

and a 35°F interior temperature, extrapolated as specified from

a test run at 0°F interior temperature. The third rating is

* For barometric variations from standard (29.92 in. Kg) of 1 in.

Hg or more, the standard wet-bulb temperature shall be lowered

1°F for each in. Hg decrease in barometric pressure.

II -2
for 35°F interior temperature, at the same ambient conditions,

with the test run at 35*F interior temperature . The published

rating shall incorporate the specified multiplier to account for

solar load (see 8.0) and shall state the interior temperature of the

rating, and the interior at which the truck was tested.

3.0 Instruments

3.1 It is suggested that temperatures be measured by one of the

following methods;

a. Thermocouple systems

b. Electric resistance thermometer systems

Accuracy of the measurements obtained with the system shall be

within the following limits;

a. Wet- and dry-bulb air temperatures

b. Brine temperatures

c. Brine temperature differences across coil in

truck body and across external comparison

heater

d. Other temperatures . The smallest scale division of the temperature measuring instrument shall not exceed twice the specified accuracy.

iU.4 r +0.4°F

0.05°F

±0 .5 °F

The temperature measuring system used for measuring temperatures

shall be calibrated, or monitored during the test, by comparison

with a certified standard temperature measuring instrument calibrated in the appropriate temperature range.

Wet-bulb temperatures shall be read only under conditions which

assure an air velocity of 1,000 ±250 ft/min over the wet-bulb,

and only after sufficient time has been allowed for evaporative

equilibrium to be attained. Care must be exercised in obtaining

wet-bulb temperatures to use distilled water on the wick, and to

have the wick damp at the time of observation. The wick must be

kept clean.

Relative humidity measurements, if used, shall be made with sufficient accuracy to obtain compliance with the accuracy requirements

for wet-bulb temperature as stated in paragraph 3.1 of this

standard. Relationship of wet-bulb and dry-bulb temperatures to

relative humidity shall be based on U. S. Weather Bureau tables.

Temperatures of brine in conduits shall be measured by inserting

the temperature measuring element within a well inserted not less

II -3
than 25 times the outside well diameter into the circuit.

Instruments or systems used to measure the temperature differences of the brine across the cooling coil in the truck and across the comparison heater shall be compared with each other

before they are installed and they shall agree within 0.05°F when

immersed in the same baths at temperatures approximating those

of use.

3.2 Brine flow shall be measured with an integrating liquid

flowmeter having an accuracy within ±0.5 percent of the volume

flow rate measured.

3.3 Electrical energy usage should be determined preferably

with watt-hour meters. On steady loads, a wattmeter may be

used in lieu of a watt-hour meter; and on steady resistance loads,

an ammeter and voltmeter may be used.

Accuracy of instruments used to measure the electrical input to

heaters in the truck or in the comparison heater shall be within

±1.0 percent of the quantity measured. Accuracy of instruments used to measure other electrical quantities shall be

within ±2.0 percent of the quantity measured.

3.4 Instruments used to measure the change of weight of the

truck being tested shall have a sensitivity requirement (see

NBS Handbook 44) of 0 o 5 pound maximum under actual test loads.

4 . 0 Test Room

An insulated test room approximately 16 ft wide and 35 ft long

and 14 ft high is required for testing trucks. A door at least

9 ft wide and 12 ft high is required at one end. The walls and

ceiling of the test room should be insulated sufficiently to

prevent condensation on the inner wall surface at standard test

conditions during cold weather. A low heat loss for the test room makes it easier to keep the temperature uniform throughout the room. A good vapor barrier material should be applied

at the inner wall surface or at the inner surface of the insulation. Separate rooms for the refrigerating equipment and instruments are desirable because of the high temperature and humidity

maintained in the test room.

Distributed heating and humidity sources are desirable in the

test room to provide uniform conditions around the test specimen

with the minimum air motion. However, some mixing of the air

with fans will probably be required to attain the specified uniformity. If fans are used, they shall be directed so that they

II-4
do not blow air against the exterior surface of the truck at

a velocity in excess of 400 ft/min. High air velocities around

the truck affect the air leakage of the vehicle and also make

precise weighing more difficult. Scales or other weighing

mechanism may be portable or incorporated in the floor con- struction.

4. 1 Cooling Load Test Apparatus

A diagram of the refrigerating equipment and temperature

measurements required to determine the cooling load is shown

in figure II-l. As indicated in the figure, the equipment con- sists of a refrigerating unit and brine chiller, a brine pump,

an air-cooling coil and fan inside the truck, and a comparison

heater and flowmeter in the brine circuit outside the truck

or test room.

The refrigerating unit and brine chiller may be single stage or multistage, and it must have a capacity of not less than twice

the cooling load of the largest test specimen, with brine leaving

the chiller at about -25°F. Capacity control is required to

adjust the cooling capacity to the cooling load of particular

specimens. The refrigerating unit and its controls should be

of a type that will produce a steady cooling effect during the

test period. Cyclic variations in brine temperature entering

the truck cooling coil or comparison brine heater should not

exceed 0.4 deg F.

The cooling coil inside the truck shall be designed without

fins or shall have fin spacing of 1/4-inch minimum to prevent

rapid stoppage w^th frost or ice. Provision should be made

for rapid defrosting of the coil. The heat transfer surface

of the coil should be adequate to absorb the cooling load of

the truck to be tested plus fan loads and a limited amount

of controlling heat with a mean temperature difference between coil brine and truck air temperature of 20 deg F or

less. The blower in the cooling unit shall deliver sufficient

air to produce a temperature difference of 10 deg F or less

between the air entering and leaving the coil.

The total electric heat input to the truck shall not exceed

one-half the measured cooling load absorbed by the air-cooling

coil. Because of this, a range of air cooling coil sizes may

be necessary, also some means of varying both the capacity of

the refrigerating unit and the brine flow rate will be needed.

II-5
The comparison heater outside the truck shall be insulated

sufficiently to reduce the heat transmission from the sur- roundings to the brine to 1/2 percent or less of the electric

heat input to the brine heater. The heating capacity of the

comparison heater shall be approximately equal to the maximum

total heat absorption of the air cooling coil in the truck,

and the heater capacity should be adjustable. Voltage regulation shall be provided for the power supply to the comparison

heater and to tne heaters inside the truck that will prevent

voltage fluctuations in excess of ±1 percent.

The brine pump shall be of a type that has an essentially flat

volume versus pressure performance curve and a pressurized shaft

seal or other means to minimize inward leakage of moist air.

The capacity of the pump shall be such that the temperature rises

of the brine through the cooling coil in the truck and through

the comparison heater shall be about 8 deg F (not less than 6 deg

F) each for the particular brine used. The brine piping circuit

shall be designed to suit the head characteristics of the pump

at the selected flow rate of the brine, and shall be insulated

to reduce heat gain.

The brine shall have suitable toxicity, viscosity, and vapor

pressure characteristics at temperatures ranging from room

temperature to -30°F. Its density shall not vary more than

0.08 percent per degree F, and its specific heat shall not vary

more than 0.02 percent per degree F, in the range of temperature used in the brine circuit. Methylene chloride meets the

density and specific heat tolerances specified, and has most of

the other desired characteristics, but other brines may be

found that are equally satisfactory.

Electric heaters of a capacity slightly greater than the increments in refrigerating capacity should be installed either in

the cooling coil or in the air discharge from the cooling coil,

and should be controlled to maintain the required truck temperature. All electric power to fans, motors, heaters, etc., in

the truck shall be measured, and the total shall not exceed

50 percent of the cooling load absorbed by the air cooling coil

during any test.

Brine lines, power cables, instrument leads, etc., may be brought

into the truck at any convenient point. Where no opening is

available, it is recommended that a suitable sleeve be installed

in one of the doors. These necessary lines must be flexible

and must be supported in such a manner that their effect on the

measured weight is minimal and constant throughout the test.

II-6

5.0 Cooling Load Test Procedure

Test methods incorporated in this standard are intended to

produce heat transfer determinations accurate within ±5 percent of the quantity measured. To achieve this overall

accuracy the test must be conducted in strict conformance

with the limitations and methods outlined in the standard.

When improved techniques and instruments are available, their

use is encouraged; but they should be approved by the organization sponsoring this standard before being substituted for

methods or instruments presently required.

The truck to be tested shall be placed on the weighing

mechanism and the test equipment and measuring devices

installed. After the truck body temperature and ambient

conditions of temperature and humidity required for a rating

test have been attained, they shall be maintained for not less

than 48 hours. The rated cooling load shall be determined from

the average of the data taken during the last 12 hours of the 48-

hour test period, and the weight gain rate shall be determined

from the measurements taken during the last 24 hours of the 48-

hour period. No interruption of steady state conditions, such as defrosting the cooling coil, shall be allowed during the final

24 hours of a test.

The temperature difference of the brine across the cooling coil

in the truck and across the comparison heater shall each

be held at a constant value between 6 deg F and 10 deg F

during the rating test.

The ambient dry-bulb air temperature shall be the average of

the observations of not less than six stations, one approximately 1 foot from the center of each surface of the truck.

The temperature difference between any two of these stations

at a given time shall not exceed 3 deg F during the test period.

The ambient wet-bulb temperature shall be the average of not

less than two points, one at the rear and the other at the

front of the truck. The difference in wet-bulb temperature

at these points of measurement at a given time shall not

exceed 2 deg F during the test period

„

The air temperature inside the truck shall be the average of

the observations at 12 stations located as follows: four at

the front, one in each corner suspended 6 inches from each

adjacent surface; four similarly located at the rear; and

II -7
four at the -middle of the truck, one at each corner 6 inches

from each adjacent surface. If desired each group of four

temperature sensing elements may be connected in parallel and

read as a single temperature, reducing the number of readings

to three. If the twelve elements are read individually, no

two readings at a given time may differ by more than 3 deg F;

if the groups of four are used, no two readings at a given time

may differ by more than 2 deg F.

During the portion of the test used to determine the cooling

load rating, all observations shall be made at not more than

30-minute intervals.

Trucks equipped with removable plug-type refrigerating units in

the front wall shall be tested for standard rating with the unit

removed and the opening carefully closed with an airtight insulated plug.

All floor drains shall be plugged during the cooling load test.

The cooling coil may be mounted at any point in the truck. Care

must be taken that air discharged from any fan does not blow

directly on joints, cracks or seams of the interior surfaces.

The brine lines within the truck must be well insulated.

II-8
6.0 Data to be Recorded

The following items must be recorded:

Item Unit

1. Date and time of test

2. Observer

3. Barometric pressure in. Hg.

4. Power input to comparison brine heater watts

5. Power input to heater in truck watts

6. Power input to fan motors, etc., in truck watts

7. Applied voltage to comparison brine heater volts

8. Applied voltage to heater in truck volts

9. Applied voltage to fan motors in truck volts

10. Electric current to comparison brine heater amps

11. Electric current to heater in truck amps

12. Electric current to fan motors in truck amps

13. Dry-bulb temperatures of air inside truck °F

14. Dry-bulb temperatures of air in test room °F

15. Wet-bulb temperature of air in test room °F

16. Temperature of brine at inlet of cooling coil °F

17. Temperature of brine at outlet of cooling coil °F

18. Temperature difference of brine in and out of

truck deg F

19. Temperature of brine at inlet of comparison

brine heater °F

20. Temperature of brine at outlet of comparison

brine heater °F

21. Temperature difference of brine in and out of

comparison brine heater deg F

22. Temperature of brine entering flowmeter °F

23. Brine flow rate lb/hr

24. Weight of truck, or change in weight lb

7 . 0 Calculations of Observed Cooling Load

Two simultaneous methods are used to determine the cooling load.

One method uses the comparison between the temperature rise of

the brine in the truck and the temperature rise in the comparison

brine heater; the other method uses the temperature rise of the

brine in the truck and the mass flow rate of the brine as measured

by the flowmeter. The results of the two methods must agree within

5 percent for a given test to be acceptable as a rating test.

Because both methods rely on the temperature rise of the brine

in the truck, two separate sets of measuring elements shall be

used to measure this brine temperature difference and must agree

within 0.1 deg. F.

II-9
The cooling load measured by the comparison method shall be

computed for the standard temperature difference of 100 degrees

by the following equation:

Cooling load, Btu/hr ATi (Ha ) - AT2 x 100

AT3

where: ATi

AT2

H2

Temperature rise of brine in the truck, deg F

Temperature rise of brine between inlet and outlet

of brine heater, deg F

Heat input to comparison brine heater, Btu/hr

AT3 = Temperature difference between air in truck and

air in test room, deg F

Hi = Heat input to heater, fan motors, etc., inside

the truck, Btu/hr.

The cooling load measured by the flowmeter method shall be

computed for the standard temperature difference of 100 degrees

by the following equation:

Cooling load, Btu/hr = ^——X x 100

AT3

where: ATi = Temperature rise of brine in the truck, deg F

M = Brine flow rate, lb/hr

C = Specific heat of brine at mean temperature in

the cooling coil, Btu/lb deg F

Hi = Heat input to heater, fan motors, etc., inside the

truck, Btu/hr

AT3 = Temperature difference between air in truck and

air in test room, deg F.

The observed cooling load shall be the average of the values

determined by the two methods.

8 . 0 S tandard Cooling Load Rating

The standard cooling load rating shall be the product of the

observed cooling load and the appropriate multiplier to account

for solar load, expressed to the nearest even 100 Btu/hr, e.g.

1200, 1600, 2200, etc.

II- 10
Rated interior

temperature

Test interior

temperature

Multiplier

0°F 0°F 1.22

35 °F 0°F .87

35 °F 35 °F 1.34

9 .0 Standard Weight Gain Rating

The standard weight gain rating is the average weight gain

rate in lb/hr determined for the final 24 hours of the test,

and shall be expressed to the nearest 0.1 lb/hr.

Weight gain rates less than 1/24 of the sensitivity requirement

of the weighing system used shall be simply listed as less

than that amount (e.g. weight gain rate = < y lb/hr, where

y = sensitivity requirement).

10 o 0 Published Ratings

Published ratings, in order to conform to this standard, shall

be identified as follows: "( sponsor’s designation ) Standard

Cooling Load Rating ( test result ) Btu/hr, Standard Weight

Gain Rating ( test result ) lb/hr, Rated at (0°F) (35°F)

interior temperature, tested at (0°F) (35°F) interior temperature.

Tests conducted in accordance with ( sponsor's designation ) Standard Method of Testing and Rating the Cooling Load of

Refrigerated Trucks."

The terms "( sponsor's designation ) Standard Method" or "( sponsor's designation ) Standard Conditions" shall not

be used in connection with published ratings unless such ratings

have been determined in accordance with this standard.

11-11


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