Reference

•Publications

NBSIR 83-2648 NAT'L INST. OF STAND & TECH

Ainot Derimental and Analytical

Investigation of a Residential Hot

Water Boiler with Finned Copper

Tube Heat Exchangers

U.S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

September 1982

Issued April 1 983

Prepared for: —QC~

100

, U5o

d3-2648

1963

U.S. Department of Energy

__fice of Conservation and

Renewable Energy

ishington, DC 20585

ncrtace -Pel

m 5 1993

NBSIR 83-2648

cJC *°°

EXPERIMENTAL AND ANALYTICAL

INVESTIGATION OF A RESIDENTIAL HOT

WATER BOILER WITH FINNED COPPER

TUBE HEAT EXCHANGERS

William J. Mulroy

Cheol Park

U S. DEPARTMENT OF COMMERCE

National Bureau of Standards

National Engineering Laboratory

Center for Building Technology

Building Equipment Division

Washington, DC 20234

September 1982

Issued April 1 983

Prepared for: U.S. Department of Energy

Office of Conservation and

Renewable Energy

Washington, DC 20585

<• o*

U.S. DEPARTMENT OF COMMERCE, Malcolm Baldrige, Secretary

NATIONAL BUREAU OF STANDARDS, Ernest Ambler. Director

ABSTRACT

In response to a request by a manufacturer of a nontyplcal boiler, the

Department of Energy requested the National Bureau of Standards to perform

laboratory measurements under controlled conditions of the effect on seasonal

performance of several features Cfinned copper tube heat exchanger, water

circulating pump delay, and gas valve modulation) of this boiler that might

cause it to be unfairly treated by the existing test procedure. As a result

of this study, recommended changes to the existing test procedure to allow

rating tests with water circulating pump delay are presented. A recommended

change to the assigned cyclic jacket loss factor and a Simplified procedure

for experimentally determining this factor are also presented. No change to

the current test procedure treatment of gas valve modulation or flue gas mass flow as a function of temperature are recommended.

Key Words: Annual efficiency; annual operating costs; boilers; fossil fuel

heating systems; jacket loss; modulating control gas fueled;

part-load performance; rating procedures; seasonal efficiency

iii
I
TABLE OF CONTENTS

LIST OF TABLES vi

LIST OF FIGURES vii

NOMENCLATURE viii

1. INTRODUCTION 1

2. DESCRIPTION OF TEST BOILER 4

3. DESCRIPTION OF TEST APPARATUS 7

4. LABORATORY TEST PROCEDURES 11

4.1 Mass Flow 11

4.2 Modulation 13

4.3 Jacket Loss 13

4.4 Pump Delay 14

5. LABORATORY TEST RESULTS AND DISCUSSION 15

6. COMPUTER SIMULATION RESULTS AND DISCUSSION ON PUMP DELAY 35

7. CONCLUSIONS * 44

REFERENCES 46

APPENDIX A - Cyclic Jacket Loss Calculation 47

APPENDIX B - Recommended Change to the DoE/NBS Test Procedure for

Furnaces and Boilers to Accommodate Testing of Boilers

with Pump Delay 51

v
LIST OF TABLES

Table 1. Comparison of Cyclic and Continuous Pump Heat-Up

and Cool-Down Tests 31

Table 2. Input Data for Computer Simulation 37

vi
LIST OF FIGURES

Figure 1. Test apparatus schematic 8

Figure 2. Test configuration 12

Figure 3. Mass flow rate of flue gas through the stack . 16

Figure 4. Flue gas temperature at cool-down test 17

Figure 5. Reduced firing rate steady-state efficiency 19

Figure 6. Comparison of steady-state modulated efficiency to full-fired

cyclic efficiency with cyclic and with continuous pump

operation

Figure 7. Comparison of flue temperature during cool-down

from full and 1/3 fired steady-state operation 22

Figure 8. Side jacket temperature during cool-down 25

Figure 9. Side jacket temperature during cyclic operation 27

Figure 10. Flue temperature during continuous water pump

operation cool-down test 30

Figure 11. Flue temperature during cool-down with cyclic and

continuous pump and with 5 minute pump delay 33

Figure 12. Three modes of boiler pump operation 36

Figure 13. Fuel utilization efficiencies with respect to

fractional burner on-times and three modes of

pump operation 39

Figure 14. Fuel utilization efficiencies with respect to

dimensionless pump on-times with a parameter of

Figure 15. Total amount operation costs with respect to

dimensionless pump on-times with a parameter of

t

vii
NOMENCLATURE

A

2 2 jacket surface area, ft (m )

c

j

jacket loss factor

°t , OFF

cool-down temperature profile correction factor for the effect of

cycling

°t,ON heat-up temperature profile correction factor for the effect of

cycling

effya

annual fuel utilization efficiency

h

9 2 jacket heat transfer coefficient 3tu/ft h°F (W/m °C)

h

c

2 2 heat transfer coefficient of convection, Btu/ft h°F (U/m °C)

h i rl

o black body heat transfer coefficient of radiation, Btu/ft n°F

(W/tn °C)

m

F non-steady-state flue gas flow rate, lb /min (kg/min)

•“f.ss

flue gas flow rate at steady-state operation, lb/min (kg/min)

qin

fuel energy input rate at steady-state operation (including any pilot

light input) , Btu/h (W)

qout

fuel energy output rate at steady-state operation, the product of

QlN

x ^ss > Btu/h (W)

t t ime , min

+

t delay time between burner shut-off and blower or circulating pump

shut-off, min

t” delay time between burner start-up and blower or circulating pump

start-up, rnin

S.OFF burner off-time per cycle, min

S.ON

*

tB , ON

burner on-time per cycle, min

burner on-time as a decimal fraction of cycle total cycle length

t

f

flue gas temperature while the system is in cyclic operation, °F

(°C)

t

f,ss

flue gas temperature at steady-state, °F (°C)

T.

J jacket temperature while the system is in cyclic operation, °F

(°C)

viii
T • j ,ss

T

RA

X0/N co

2

y

e.

.

nss

nu

9

3

0

J,0,X

ton

t

off

^J,0,X

jacket temperature at steady-state, °F (°C)

laboratory room temperature, °F (°C)

concentration by volume of C0

2 present in dry flue gas, %

ratio of blower or pump on-time to burner on-time

emissHvity

steady-state efficiency, %

part-load fuel utilization efficiency , %

temperature difference between steady-state and on-cycle jacket

temperature, °F (°C)

temperature difference defined by equation 2 in Appendix A, °F

(°C)

heat-up time constant, min

cool-down time constant, min

temperature difference between steady-state and off-cycle jacket

temperature, °F (°C)

temperature difference defined by equation 1 in Appendix A, °F

(°C)

ix

1. INTRODUCTION

The residential boiler under study incorporated several nontypical features.

The most outstanding of these features were the use of a finned copper tube heat

exchanger instead of the traditional non-finned cast iron shell, and the incorporation of a modulating gas valve and a temperature sensing time delay

which controlled shut-off of the water-circulating pump.

Field data taken by the manufacturer indicated a seasonal efficiency

substantially greater for boilers of this type than that calculated by the

DoE/NBS Furnace Boiler Test Procedure. (Reference [1]* referred to as the

Test Procedure in the remainder of this report.)

In response to the manufacturer's request, the Department of Energy requested

the National Bureau of Standards to perform laboratory measurements under controlled conditions of the effects of various innovative features in these

boilers and to determine what changes, if any, should be made in the Test

Procedure [1] to treat these units fairly. This report contains results of

tests performed at NBS on a boiler supplied by the manufacturer as typical of

their production units. Tests and calculations in this report were directed

to the five following areas of concern:

a) Validity of DoE/NBS Test Procedure [1] Flue Gas Mass Flow Equation - The

purpose of these tests was to determine if there was something different about

the internal geometry of this boiler design that might alter the combustion air

mass flow rate during heat-up and cool-down from that which is assumed in the

Test Procedure [1],

*References listed on page 46.
b) Effect of Gas Valve Modulation on Seasonal Efficiency - The Test Procedure

[1] assumes that there is no advantage in seasonal efficiency to be gained by

modulating gas flow as opposed to cycling between a bumer-off and full-fired

condition.* Tests were performed at several reduced firing rates and compared

to the efficiency resulting from this assumption.

c) Value of in Test Procedure [1] - is a multiplier on the measured

steady-state jacket loss for calculating the jacket loss for units installed

outdoors during cyclic operation in a 42°F ambient. This operating condition

is assumed to result in an efficiency representative of the mean seasonal

value. Jacket temperature measurements were made which would allow calculation

of the cyclic portion of for several assumed operating conditions.

d) Value of Delayed Pump Operation - Cool-down and heat-up tests were performed

with the water pump running continuously to determine the maximum effect likely

on seasonal efficiency due to water pump delay as opposed to the simultaneous

cycling of the water pump with the gas burner as assumed by the Test Procedure

[1], It should be emphasized that continuous pump operation is not a

recommended control strategy for actual field applications but was employed in

this laboratory study as a limiting case.

* Work done in another study on development of a test procedure for modulating

fuel controls has been completed for vented household heaters and furnaces [6].

Procedures developed in that program can be applied where applicable as the

need arises.

2
e) Computer Simulation for Pump Delay - An analytical study was also performed

to see if the current DoE/NBS Fumace/Boiler Test Procedure [1] could evaluate

the performance of a residential hot water boiler system with a finned tube

heat exchanger, and with a control device for the pump delay. Based upon the

available information, computer simulations were made using an existing boiler

simulation program. The NBS DEPAB program [5] was used for this analysis after

some modifications were made. The analysis focused on the control strategy

related to the pump operation.

Sensitivity analysis of the pump delay parameters and cost analysis of the

boiler system were made. In the analysis, the annual costs in operation of

the boiler were obtained using the annual operation cost analysis routine in

the Test Procedure [1].

3
2. DESCRIPTION OF TEST BOILER

The name plate data of the tested boiler were:

Fuel : Natural Gas

Input Rate: 125,000 Btu/h (37 kW)

Output Rate: 100,000 Btu/h (29 kW)

Minimum Input Rate: 43,750 Btu/h (13 kW)

Maximum Water Pressure: 160 psi (1100 kpa)

2 2 Heating Surface Area: 11.4 ft (1.06 m )

The boiler had five burners running the length of the combustion chamber. The

boiler tubes were arranged in a single horizontal row, approximately 20 inches

(51 cm) above the burners. The combustion chamber connecting the burners and

the boiler tubes was insulated with a 2 inch (5 cm) thick cast refractory liner

and was surrounded by fiber insulation approximately 9.5 inches (24 cm) wide and

13 inches (33 cm) long. The boiler bottom jacket surface was approximately

7 inches (18 cm) below and the top jacket surface 27 1/2 inches (24 cm) above

the top surface of the burners. Top and bottom jacket surfaces were both

uninsulated.

There were 8 transverse boiler tubes with a finned length of approximately

9 inches (23 cm) and possessing approximately 67 integral fins each of 5/64 inch

(20 mm) thickness. The diameter of these circular fins was 1.5 inches (3.8 cm)

and the outside diameter of the boiler tubes was 13/16 inches (2.1 cm). The

8 transverse boiler tubes were divided into two groups of four; half crossing

from the water inlet of the boiler to a cast iron manifold and the other half

recrossing from the manifold to the water outlet. The inlet and outlet (return

4
and supply) fittings were combined in a single iron casting. Vee shaped flow

guides were fitted on top of the finned boiler tubes.

The boiler stack was 5 inches (12.7 cm) in diameter and 13 1/2 inches (34.3 cm)

in length from the boiler top to the relief opening of its integral draft hood.

In normal operation, gas modulation is provided by a 1/2 inch (1.3 cm), 130°F

(54 °C) to 180°F (88°C) , modulating valve with its sensing bulb in the hot water

supply line. Modulation temperature range is adjustable using a dial with settings from 1 to 9. At the lowest setting (1) the valve wodld be at its maximum

firing rate position for outlet water temperatures of 105°F (41°C) and below;

at its minimum firing rate at a 120°F (49° C) outlet water temperature; and off

at a outlet water temperature of 130°F (54°C). Similarly at its highest setting

(9) the valve would be at its maximum firing rate position for outlet water temperature of 165°F (74°C) and below; at its minimum firing rate at a 180°F

(82°C) outlet water temperature; and off at an outlet water temperature of 190°

F

(88°C). This modulating valve was disabled for all tests described in this

report. For tests in which a reduced firing rate was desired, the gas supply

to the boiler was throttled using a manually operated valve.

The water-circulating pump control of the boiler was designed to provide pump

start-up at gas cut-on and delay on pump shut-down after gas cut-off. The pump

cut-off delay switch senses outlet water temperature and had a factory-set differential of 15°F (8.3°C).

5
The manufacturer's literature recommended that the control be set at 20°F (11 °C)

to 30°F (17°C) below the outlet water temperature setting of the modulating

valve, but not less than 120°F (49°C) . This automatic pump delay feature was

disabled for all tests described in this report. The factory "on" setting for

the delay switch was at 120°F (49°C). Thus in operation, at the factory settings, the pump starts simultaneously with the burner by means of a relay in

parallel with the time delay switch. If the outlet water temperature exceeds

120°F (49°C) , the delay switch closes. When the burner is shut off this relay

switch opens immediately, but the delay switch keeps the pump on until the outlet water falls to 105°F (41°C) . For tests in which pump delay was desired,

pump operation was controlled by a manually operated switch.

Ignition for this boiler was by a continuously burning pilot light.

6
3. DESCRIPTION OF TEST APPARATUS

Testing was conducted using the boiler test apparatus schematically shown in

Figure 1. The test unit was mounted on four 7 3/4 inch (20 cm) legs allowing

' 3 room air circulation under the boiler. An insulated 125 gallon (0.47 m ) water

tank, circulating pump, hot water dump, auxiliary boiler, and city water makeup comprised the major support equipment. In order to achieve repeatable water

flow rates, pairs of manually operated valves were employed—one valve for

flow control and the other for open/shut (flow/no flow) operation.

The auxiliary boiler, a gas-fired 85 kBtu/h (25 kW) input unit, was used to raise

the temperature of the water in the storage tank between tests. Pump and boiler

operation was controlled manually.

A turbine meter in the inlet pipe to the boiler was used to measure the water

flow through the boiler. A manually read totalizer which counted pulses from

the turbine meter was used to determine the total water flow through the boiler

during each test. The accuracy of the turbine meter was verified using a weigh

tank and a timer at inlet water temperatures from 150 to 180°F (66 to 82 °C)

Occasional calibration checks were made throughout the period of time that tests

were being conducted.

A 32-junction thermopile made from 30 AWG type T ( copper/constantan) thermocouple

wire was used in stainless steel wells inserted in the inlet and outlet water

lines to measure the temperature differential across the boiler. The stainless

steel wells had an outside diameter of 3/8 inches (0.95 cm) and a 6- inch (15.2 cm)

7
Figure 1. Test apparatus schematic

8
immersion length, as recommended in ASHRAE Std. 41.1-74 [4], The flue gas

temperature was measured by a 9-in-l averaging grid of chromel-alumel , type K,

thermocouples 11 inches (28 cm) from the boiler dutlet. Additional type T

thermocouples were provided to measure the test room ambient, dry bulb and wet

bulb temperatures, the supply gas temperature, and the supply and return air

temperatures

.

Surface temperatures of the jacket were measured using 148 thermocouples bonded

to the surface with epoxy and painted to match the unit. Fifteen thermocouples

were on the top, fifteen on the bottom, and the remaining 118 on the sides.

Thermocouples on similarly facing surfaces were connected in parallel to allow

faster monitoring during transient tests.

The temperature data were recorded by an automatic data acquisition system

having a programmable time interval capability between data scans.

Flue gas samples were taken at the center of the flue, 11 inches (28 cm) from

the boiler outlet at the flue temperature measurement plane. The concentrations

of carbon dioxide and methane tracer gas in the flue gas were measured tising

instruments of the infrared absorption type. The full scale reading of the

carbon dioxide detector was 15 percent and, of the methane detector, 500 ppm. A

dessicant column was used to remove moisture from the flue gas before it entered

the detectors.

The methane tracer gas flow rate was measured with a 100 ml bubble meter. The

tracer gas was injected through a manifold beneath the burner tubes. The

9
percentage by volume of oxygen present in the flue gases was measured using an

electrochemical sensor with oxygen-sensitive electrolyte. The concentration

of oxygen provided a check on the CO2 measurement and was not used in the

calculation procedure.

The higher heating value of the natural gas used was continuously recorded on a

gas calorimeter located at a nearby NBS building. The dmount of gas consumed

during a test was measured with a dry-type positive displacement meter with a

one cubic foot per revolution register. The gas temperature at the meter was

measured using a type-T thermocouple and the gas pressure at the meter was

determined using an open tube manometer and a barometer.

The flue from the furnace top to the bottom of the draft hood was wrapped in 2

inch (5 cm) thick R-7 glass fiber insulation.

A five foot (152 cm) stack was installed on top of the draft hood. Tests were

performed both with the draft hood open and sealed. A vane anemometer was

mounted 1 inch (2.5 cm) above the top of the stack for stack flow rate measurement during some cool-down tests.

The anemometer had been calibrated under a geometrically similar flow condition.

Air velocity at the top of a calibration circular duct of the same diameter and

height as the stack was measured at room temperature. A high capacity dry gas

meter was installed between the air supply line and the calibration duct. The

anemometer used here was 4 inches (10.2 cm) in diameter with an eight vane rotor.

2 2 Its effective cross sectional area was 9.42 in (60.8 cm )

10
4. LABORATORY TEST PROCEDURES

4.1 MASS FLOW

A steady-state test at full load following the guidelines of the DoE/NBS

Fumace/Boiler Test Procedure [1] was first performed on the test boiler

(Figure 2(a)). The draft hood was then sealed as seen in Figure 2(b) and the

five-foot (152 cm) stack adjusted to produce the same flue concentration

as that measured during the first steady-state test. This second steady-state

test, with the draft hood blocked, was used as the initial condition prior to

cool-down for mass flow measurement.

It should be noted that the test condition of a sealed draft diverter and a

five-foot stack with sufficient restriction to reproduce the unsealed draft

diverter CO^ concentration is the DoE/NBS Test Procedure [1] preliminary test

condition for units which incorporate draft diverters instead of draft hoods.

The Test Procedure [1] flow equations are designed to fit this sealed draft

diverter test condition. For comparability in flow measurement, draft hood

units should also have their relief openings sealed and stacks restricted.

A cool-down test was performed during which the stack mass flow was measured by

two independent methods; vane anemometer mounted above the stack and methane

tracer gas. After the boiler was allowed to return to steady-state idle conditions (the following day) , an additional data point was taken representing the

flow caused by pilot light operation. The stack flow data were normalized by

division by steady-state values obtained during the pre-cool-down steady-state

period. Vane anemometer steady-state data were used for normalizing the vane

anemometer data. A mass flow rate value calculated from the stack CC^

11
Cool-down temperature

profile test

Cool-down temperature

profile and mass flow rate test

Figure 2. Test configuration

12
concentration and the fuel gas flow rate was used for the tracer gas cool-down

data normalization.

The draft hood was then opened, and steady-state and cool-down tests were

performed in accordance with the Test Procedure [1] to provide a cool-down flue

temperature curve for comparison to that obtained during the sealed draft hood

cool-down test (as in Figure 2(a))

.

4.2 MODULATION

Steady-state cool-down and heat-up tests were performed according to the Test

Procedure [1] at full load, and with the gas input rate reduced to 2/3 and to

1/3 of its rated value.

4.3 JACKET LOSS

The jacket temperatures were measured during all steady-state, cool-down, and

heat-up tests.

A full-fired, cyclic test was also performed. For this test the boiler water

was continuously circulated during both burner on-and off-cycles. The return

water temperature was kept at a nominal 165°F (73.9°C) by a thermostatically

controlled, auxiliary boiler in the test loop. These cycle periods were 9.68

minutes on and 33.26 minutes off as specified by the Test Procedure [1] for

boilers

.

Comparison of the jacket loss during the full fired, cyclic test and the full

fired steady—state test would allow direct calculation of C^, (exclusive of

ambient temperature effects) . The heat-up and cool-down from steady-state

13
jacket temperature data were used to see if a time constant method could be

used to calculate the cyclic jacket temperatures (eliminating the need for

cyclic testing)

.

4.4 PUMP DELAY

Full load steady-state, cool-down, and heat-up tests were performed with a 165 °F

(73.9 °C) steady-state return water temperature with the water circulating pump

cycling with the burner as specified by the Test Procedure [1], and were then

repeated with the pump running continuously.

The above comparison was then repeated with a 105°F (40.6°C) steady-state

return water temperature and, during steady-state operation, a proportionately

reduced supply water temperature (i.e., constant water flow rate).

During the off-portion of the cyclic water flow tests the make-up water to the

test loop was reduced to a minimum and the system insulation relied on to maintain the loop water temperature. Continuous water flow tests were conducted

both by operating a thermostatically controlled auxiliary boiler in the test

loop to maintain the return water at its steady-state value, and by allowing the

inlet water temperature to decay at a rate given by leaving the system make-up

water at the value used during steady-state operation and the auxiliary boiler

off.

Continuous pump tests were used instead of actual pump delay tests in order to

show the maximum effect possible due to pump delay and to eliminate the variable

of length of delay.

14
5. LABORATORY TEST RESULTS AND DISCUSSION

a) Mass Flow - The measured stack mass flow with the draft diverter sealed

divided by its full load value is shown as a function of normalized temperature

in Figure 3. Points at 0.0, 0.0 are included by definition rather than measurement. The low temperature points are those resulting from pilot light operation.

The conformance to the predictive equation of the DoE/NBS Furnace/Boiler Test

Procedure [1] shown as a solid line is quite good, deviating by 8 percent

at its worst.

In Figure 4 the flue temperature, as a function of time, measured during

cool-down from steady-state with the draft hood open (Figure 2a) is shown.

Also shown is the flue temperature measured during cool-down from steady-state

with the same draft hood sealed (Figure 2b) and the flow restrictor adjusted to

provide the same steady-state CC^ concentration as in the open hood test. Both

this agreement and that shown in Figure 3 are closer than would be expected for

a unit employing a draft diverter instead of the draft hood with which this

unit was equipped (see [2], [6]).

Only flue temperature profiles are normally measured during the cool-down test

in the Test Procedure [1], For ease of testing, draft hood units are allowed

to be tested with the draft hood open on the assumption that in this configuration the cool-down temperature curve will not be significantly different from

that which would have been measured with a sealed draft hood and the stack

restricted to provide the same CC^ value as the open hood condition. As can be

seen from Figures 3 and 4 this assumption is valid for this unit. Therefore,

a boiler of this internal geometry will receive a fair (consistent

15
Figure 3. Mass flow rate of flue gas through the stack

16

i'
LU

500 r

~ -

i 1 1 4 ,

o Draft hood open

4qq _ a Draft hood blocked

&°A

°Aa ^ _

*

200 - & * Z

100 i i i i 0 10 20 30 40 50

TIME (Min.)

A

0^A

°$z - %

A ft

T T

o Draft hood open a Draft hood blocked

a a

Figure 4. Flue gas temperature at cool-down test

17
with other designs) rating if tested in accordance with the current Test

Procedure [1].

It should be noted that comparability between units in the Test Procedure [1]

is based upon measurements with a 5 foot (152 cm) stack without draft relief.

If mass flow measurements were made with a substantially shorter stack (as

would be the case for the unit with the draft hood open) or in the flue and

compared to the coincident temperatures, the results would not be comparable to

units tested according to the Test Procedures [1], since the measured temperature

would not be representative of the height averaged temperature causing

flow. The Test Procedure [1] assumes that the flue flow is the result of convection in a stack at a uniform temperature, Tp, surrounded by an ambient

atmosphere at a lower uniform temperature, Tp^.

4 b) Modulation - Steady-state tests were performed with the unit full fired and

with the gas input reduced to 2/3 and to 1/3 of the full-fired input rate. The

results of these tests are shown in Figure 5.

The non-cycling efficiency achieved by modulating firing rate is compared in

Figure 6 to on/off cycling efficiency achieved by cycling with on- and off-times

varied to match load at the full-fired gas input rate both with and without pump

delay. The efficiency plotted in Figure 6 is modeled after the part load fuel

utilization efficiency, n , of the Test Procedure [1] with the difference that

the outdoor temperature was varied to produce a curve instead of a point value,

and that cycle lengths and time constants consistent with the outdoor ambient

temperature and the operational mode were used. The non-cycling efficiency is

18
Figure 5. Reduced firing rate steady-state efficiency

19
PART

LOAD

FUEL

UTILIZATION

EFFICIENCY,

0 10 20 30 40 50 60 70 80 90 100

PERCENT ON-TIME

Figure 6. Comparison of modulated steady-state efficiency to

full-fired cyclic efficiency with cyclic and with continuous

pump operation

20
the steady-state efficiency reduced by the infiltration loss. This comparison

indicates that while modulation would result in an efficiency improvement for

a cyclic pump boiler, the incorporation of pump delay would result in a still

greater improvement. It should be further noted that the improvement brought

about by these two features is not additive. In fact, for this unit, modulation

with pump delay is less efficient than cyclic burner operation with pump delay.

In an actual application it is presumed that this unit would not modulate at

light loads, since firing rate reduction occurs only in response to near maximum outlet water temperatures. Such temperatures would only occur near the

design outdoor temperature with the unit installed in a single loop system with

a 70 percent oversizing factor as presumed by the Test Procedure [1].

The flue temperature elevation above the test room ambient temperature during

cool-down from steady-state operation with the unit full fired is compared in

Figure 7 to that occurring during cool-down from steady-state operation with the

unit fired at 1/3 of its rated input. Time constants were calculated for the

flue temperature heat-up and cool-down curves as follows

:

Full-fired 1/3 fired

T 3.70 min. 4.13 min.

on T __ 18.02 min. 17.22 min.

off

In spite of the great difference between firing rates and steady-state flue

temperatures at these two first conditions, the flue temperature time constants

were changed very little.

21
FLUE

TEMPERATURE

-

AMBIENT

TEMPERATURE,

i//F

Figure

400 i r 1 r

o Full fired

a 1/3 fired

IQ

\

;°s

100 h» \

zK %,

- \

°bV

'A

^

'A.

'Q. o '4

A \

Os

'A ' s \ \ \

10 -

\ \ \ \

' \ _

VA \_

\ \ \ 5'

\ _

\

10 20 30

TIME, MINUTES

40

7. Comparison of flue temperatures during cool-down

from full and 1/3 fired steady-state operation

22
c) Jacket Loss - The DoE/NBS Fumace/Boiler Test Procedure [1] calculations

contain two correction factors for use with the laboratory measured jacket loss

in calculating the part load efficiency and operating cost of units that are to

be installed outdoors.

The first of these is CL in step 27 of paragraph 4.1 [1] which is the ratio of

the cycle jacket loss in a 42°F (6°C) ambient to the steady-state jacket loss

measured in the laboratory. It accounts for the fact that the unit is assumed

to be On only 22.5 percent of the time at this outdoor temperature and that the

off-period (and on-period) jacket losses can significantly reduce the part load

efficiency, ri^, of a furnace or boiler. Assigned values of CL (step 27,

paragraph 4.1) are 3.3 for furnaces and 4.7 for boilers.

The second of these factors is the value 3.3 in the equation for Q in

out

Section 4.2 which is the ratio between the steady-state jacket loss in a 5°F

(-15 °C) ambient and the steady -state jacket loss measured in the laboratory.

It is pure coincidence that for furnaces the value of CL happens to be numerically equal to the steady-estate jacket loss factor used in calculating Q

Laboratory tests of this boiler were directed only toward determining the

cyclic component of CL . It was assumed that, for a boiler of this type, the

variation of jacket loss with ambient temperature would be slight. The heat

exchanger covered only the top of the fire box, and the majority of the surface area, the sides and bottom, were exposed to the fire box temperature on one

side and the ambient on the other. In a conventional boiler or furnace with

large jacketing heat exchangers, the conditioned air or water temperature would

23
more nearly govern the Jacket loss than would the fire box temperature as in

this boiler. It is reasonable to assume that if the surrounding ambient temperature were reduced, the combustion gas (and fire box) temperature would be

reduced by the same amount, resulting in the temperature difference causing heat

transfer to be nearly constant with ambient temperature variation. It was

therefore assumed for the test plan and calculation procedures followed in this

report that the jacket loss would be invariant with ambient temperature. The

effects neglected in this assumption are the tendency of the top surface temperature to follow the water and flue temperatures and of the convective and

radiative heat transfer coefficients to decrease with decreasing temperature.

These two error sources would tend to be small and to cancel one another.

The tests performed to determine the cyclic portion of were a steady-state,

full load measurement of jacket loss, to serve as a reference, measurements of

jacket temperature during cool-down and heat up tests, to determine time constants, and measurements of jacket temperature during repetitive cyclic operation at the on-and off-cycle lengths assumed by the Test Procedure [1] as

typical for a boiler operating in a 42°F (6°C) ambient temperature (9.68

minutes on, 33.26 minutes off).

The average jacket temperature of the 118 thermocouples on

the sides during cool-down from steady-state is compared in Figure 8 to a

value calculated from the measured values 3.75 minutes and 25 minutes after

shut-off, using assumptions of exponential decay and a jacket temperature

at infinity equal to the room ambient temperature. The use of room air

as an infinity value instead of the value 45 minutes after shut-off, as

used by the Test Procedure [1] in flue and stack temperature calculations,

was felt to be more accurate because of the slow rate of decay of jacket

24
0 5 10 15 20 25 30 35 40 45

TIME, MINUTES

Figure 8. Side jacket temperature during cool-down

25
temperature as compared to the flue temperature, and because it was expected that

the infinite time value would, in fact, be close to the room ambient temperature.

Time constants for heat-up were calculated from experimental data taken 1 minute

and 5.5 minutes after burner cut-on, which occurred at 50 minutes after burner

cut-off as specified by the Test Procedure [1], The resulting equations were not

compared to the experimental heat-up data because of the assumption of the jacket

temperature at infinite time being equal to the room ambient instead of that occurring

45 minutes after shut-off. These time constants were then used to predict the

cyclic jacket temperature on each surface using a calculation procedure similar

to that used for the flue temperatures in the Test Procedure [1],

A comparison of the predicted cyclic jacket temperatures for the unit side

surfaces to the values measured in repetitive cyclic tests is shown in Figure 9.

These jacket temperature profiles predicted from heat-up and cool-down time

constants and steady -state temperatures were felt to be sufficiently close to

the experimental values to justify their use in cyclic jacket loss calculations.

This cyclic jacket loss calculation procedure is discussed in greater detail in

Appendix A.

The ratio of the cyclic to the steady-state jacket loss was calculated for

cyclic and continuous pump operation at nominal entering water temperatures of

165°F (74°C) and 105°F (41°C) , using measurements taken of jacket temperature

during steady-state, heat-up and cool-down tests in the normal laboratory

ambient (approximately 80°F (27°C)) and using the calculation procedure of Appendix

A with the following results:

26
Figure 9. Side jacket temperature during cyclic operation

27
Pump

Operation

Entering (Return)

Water Temperature

167°F

Cyclic Jacket Loss

Steady-State Jacket Loss

Cyclic

Continuous

Cyclic

Continuous

164 °F

105°F

105°F

1.21

1.09

1.18

1.02

Because of the small spread of these cyclic/steady-state jacket loss ratios and

because the manufacturer of this design boiler recommends low return water temperatures for most applications, a pump delay value of 1.0 is recommended for

CL instead of the value of 4.7 specified by the Test Procedure.

The intuitive explanation for the near unity value of CL is that the jacket of

this boiler surrounds the fire box where a more conventional design would surround the fire box with heat exchanger surface and then the jacket. Thus this

boiler has a high jacket temperature when operating which drops rapidly during

the off-cycle. The conventional design, because of the surrounding heat exchangers, would have a lower on-cycle jacket temperature resulting in the off-cycle

jacket loss being proportionately greater in comparison to on-cycle value.

The lower value of CL with continuous pump operation is the result of the

circulating water recovering heat that would otherwise have been lost by the jacket.

d) Pump Delay - The maximum effect likely from water circulating pump delay was

determined by comparing the results from a cool-down and heat-up test performed

according to the Test Procedure [1] (pump cycled with burner) to tests performed

identically except for continuous pump operation. The continuous pump tests were

28
performed (1) with the inlet water temperature held constant by an auxiliary

boiler in the test loop and (2) with the inlet water temperature allowed to

decay at an arbitrary rate given by allowing makeup water to continue to

replace test loop water during cool-down and heat-up The same water flow

rate was used as during the steady-state period to balance the unit heat capacity.

It was observed during the continuous pump operation tests that the flue

temperature during cool-down tended to approach the inlet water temperature

as is shown in Figure 10. This means that in boilers which have pump delay

the off-cycle loss is dependent upon the inlet water temperature. This is not

the case in boilers without pump delay, in which case the off-cycle stack

temperature approaches the test room ambient air temperature during off-cycle

cool-down. Steady-state efficiency is affected by the mean (average of inlet

and outlet) water temperature. The Test Procedure [1] specifies outlet water

temperature and a temperature rise which may be conveniently obtained in a

test laboratory, and which result in a mean water temperature typical of an

actual application. The off-cycle loss dependency of pump delay boilers on

inlet water temperature would require changing the allowable inlet water

temperature of the Test Procedure [1] to values typical of an actual application if the scope is to be expanded to cover this design feature.

Because of the importance of this variable, a second set of tests was performed

at a lower water temperature to clarify its effect. The test conditions and

results for these continuous and cyclic pump tests are summarized in Table 1.

29
Figure 10. Flue temperature during continuous water

pump operation cool-down test

30
Table 1. Comparison of Cyclic and Continuous Pump Heat-Up and Cool-Down Tests

Cyclic Pump Continuous Pump Continuous Pump

^Auxiliary Boiler\

on /

/Decaying

V Water

Inlet

Temp

.

Gas input, kBtuh 125.6 125.3 125.6 125.4 124.5 125.4

Steady-state flue temp.,

°F

497.9 467.1 496.1 468.4 492.6 468.2

Steady -state outlet water

temp . , °F

200.5 147.2 200.4 147.2 200.5 147.1

Steady-state inlet water

temp . , °F

167.1 105.2 163.7 105.0 159.2 105.1

Steady-state mean water

temp. , °F

183.8 126.2 182.1 126.1 179.7 126.1

Inlet water temp, at

cut-on*, °F

156.5 108.6 163.9 104.6 125.7 85.9

Heat-up time constant.

Ton

4.24 4.09 3.91 3.70 5.02 4.38

Cool-down time constant,

Toff

20.9 19.3 11.0 10.3 13.2 12.0

Steady-state efficiency, % 78.5 79.4 78.5 79.2 78.7 79.2

Part load fuel utilization

efficiency, 0 , %

60.9 62.5 66.8 71.0 68.6 71.7

Annual fuel utilization 60.3 61.8 66.1 70.2 67.9 70.9

efficiency, EFFY^, %

*The above listed inlet water temperature at cut-on was recorded one minute after

cut-on to account for the time lag resulting from the mass of the thermocouple wells

in the water circulating system.

31
The flue temperatures measured during the 165°F (74°C) return water temperature

cyclic and continuous pump cool-down tests are compared in Figure 11 and the

resulting efficiencies are compared in Figure 6.

The difference in part load efficiency between cyclic and continuous pump

operation tests is seen to be quite significant. When continuous pump operation is employed, the effect of the inlet water temperature is significant.

Decay of water temperature at the rates used in these tests had a comparatively

small effect.

An additional test was performed for comparison to the above cyclic and

continuous pump tests in which a 5 minute delay in pump shut-off after burner

shut-off was employed. This test resulted in a flue temperature which

followed the continuous pump curve until pump shut-off and which then rebounded

and approached asymptotically from below the cyclic pump flue gas temperature

decay curve (see Figure 11) .

It should be emphasized that continuous pump operation was employed instead

of actual pump delay because it served as an extreme case for showing the

effects of pump delay. An actual continuous pump operation design in field

operation may result in excessive pump power costs and, with sufficiently long

boiler off-periods, would eventually result in heat from the house being trans^

ferred to the boiler and out the stack if not fitted with a stack damper. This

latter occurrance would be most likely with intermittent ignition instead of

the standing pilot employed on the test boiler since in the former case the flue

temperature would approach the room air temperature, and in the latter case would

32
Jo ‘HniVHdMBi SV9 3mj

I

;

33

Figure

11.

Flue

temperatures

during

cool-down

with

cyclic

and

continuous

pump

and

with

5

minute

pump

delay
approach a temperature approximately 20° F (11°C) to 30° F (17°C) above the

room temperature.

34
6. COMPUTER SIMULATION RESULTS AND DISCUSSION ON PUMP DELAY

A control strategy analysis was made for general cases using the NBS DEPAB

program [5]. In the simulation model, the heat exchanger was assumed to be a

low mass shell type. Measured data was used as input to the simulation program.

When the measured data was not sufficient, an estimation was made. Input data

for the computer simulation is given in Table 2. The main difference of design

between the actual boiler and the boiler under simulation was the heat exchanger.

In addition to this, the circulating pump of the actual boiler operated by

sensing the water temperature, but the simulated boiler operated by

imposed time delay. This time delay feature was much easier to handle for

sensitivity analysis of pump delay operation.

In the absence of field test data, three likely modes of operation were

conceived as seen in Figure 12.

Mode I : The pump operates in the same phase as the burner

Mode II: The pump operates incorporating time delay

Mode III: The pump operates continuously regardless of the burner

operation

The Test Procedure calls for Mode I operation for all boiler performance tests.

When the burner is turned on, the water circulating pump is also turned on.

When the burner is off, the pump is off. In Mode I operation, no pump delay

is allowed.

In Mode II, the water pump operates in a similar way to an air circulation

fan of a forced warm air furnace. This mode was the focus of the study.

35
* I, OFF t B,ON

MODE

Burntr on

Pump on

Burner off Pump off

MODE II

t

l, ON

MODE III

Figure 12. Three modes of boiler pump operation

36
Table 2. Input Data for Computer Simulation

Fuel

:

Natural Gas

Burner Input Rate: 125.6 kBtu/h (36.79 kW)

Pilot Light, Input Rate: 426.0 Btu/h (124.8 W)

Pump Input Rate

:

250 W

Steady-State Efficiency: 78%

Heat Exchanger Weight*: 60 lb (27.2 kg)

Type of Heat Exchanger: Shell type

Water Flow Rate: 2,452 lbm/h (1,112.2 kg/h)

Water in the Heat Exchanger: 4.16 lb (1.89 kg)

Water in the Boiler*: 42.7 lb (19.37 kg)

Jacket Weight

:

23.3 lb (10.57 kg)

Overall Heat Transfer Coefficient of Jacket: 0.5 Btu/ft

2h°F (2.837 W/m2K)

Room Air Temperature: 70°F ( 2 1 . 1 ° C)

Outdoor Air Temperature 42 °F (5. 6°C)

Inlet Water Temperature 105°F (40.6°C)

Outlet Water Temperature at Steady-S tate

:

147°F ( 6 3 . 9

° C)

Off-Cycle Draft Factor for Flue Gas Flow: 1.0

Off-Cycle Draft Factor for Stack Gas Flow: 1.0

Bumer-On Time

:

9.68 min.

Burner Off-Time: 33.26 min.

*More water in the boiler was added to consider the effect of heat capacity of

the fire brick walls.

37
Computer simulations were performed for all three modes employing a constant

inlet water temperature. Since the inlet water temperature at a field installation depends upon many factors such as heat losses through the building envelope,

the flow rate of circulating water, outdoor temperature, infiltration loss, etc.,

the inlet water temperature was assumed to be constant at all times. Through

the computer simulations, which were based on the heat loss method, fuel utilization efficiencies, n , were obtained at each mode in terms of fractional burner

u

* on-times, t , as shown in Figure 13. As seen in the figure, higher efficienB , ON

cies were obtained at Mode II and Mode III operation than at Mode I. In Figure

13, two parameters, t and t+

, which are a pump delay time during on-period and

a pump delay time during off-period, respectively, were used. These parameters

are shown in Figure 12 - Mode II.

When the water circulation continued during off-period, the flue gas

temperature was lowered rapidly. The heat loss through the stack of the boiler

was thus reduced, resulting in higher fuel utilization efficiency. The analysis

was continued to obtain Figure 14 showing the effect of t +and t and ri at 22.5

percent fractional burner on-time. The dimensionless pump on-time, y, is

defined as in the revised version of the Test Procedure as:

i +

t+ - t_

y

- 1 + —

B ,0N

The value of y in Figure 14 was calculated using a constant tg equal to

9.68 min., as recommended in reference [1]. For a given y, the efficiency

decreased as the time delay during on-cycle, t , increased. Pump operation

38
*B,ON

B ’°N

t B,OH + t B,OFF

Figure 13. Fuel utilization efficiencies with respect to fractional

burner on-times for three modes of pump operation

39
Figure 14. Fuel utilization efficiencies with respect to dimensionless pump on-times with a parameter of

* _/t B,ON

40
without delay during on-period gave highest efficiency. Also, the prolonged

pump operation did not improve the efficiency of the low thermal mass boiler

(note that ri u did not show the effect of increased energy usage) . A control

strategy with proper choice of t +

, t , and y is needed to minimize the energy

loss as well as the equipment decay rate.

Now consider the annual operating cost of the boiler system which includes

the boiler itself and the circulating pump. Other associated devices like a

thermostat, electromagnetic valves and sensors are excluded. Annual operating

cost mainly depends on the fuel and electricity costs. Employing the annual

cost calculation procedure in the Test Procedure and results of computer simulations given in Figure 14, annual costs, in terms of the pump operating time, y,

for selected values of t /t_ were calculated. The results are shown in

B ,0N

Figure 15. To obtain Figure 15, the following numerical values were assumed

in addition to the boiler data:

Design Heating Requirement = 100 kBtu/h

Heating Load Hour (HLH) = 2080 h

Fuel Cost Rate (Natural gas) = 50q/10^ Btu

Electricity Cost Rate = 7c/kW-h

From Figure 15, there exists a minimum cost operation curve for a specific

condition. The total operation costs for y > 1 and small value of t / 1_ ...

are always less than that for y

= 1 and t / t = 0 (Mode I operation)

.

B , ON

One should realize that the annual operation cost figure is provided as an

example. Under normal field operating conditions, the operation cost of a boiler

system varies depending upon the boiler configuration, control strategy, heating

load hour, design heating requirement, fuel cost and electricity cost.

41
TOTAL

ANNUAL

OPERATION

COST

($/yr.)

y = 1 +

t B.ON

Figure 15. Total annual operation costs with respect to dimensionless pump on-times with a parameter of

* /X B.ON

42
The electricity cost for continuously running the circulating pump results in

the increase of operation cost. Unless the pump delay at the start-up during

on-period is considerably long, most cases of operation in Mode II ensure

lower cost than in Mode I.

43
7. CONCLUSIONS

The unit tested was found to conform to the assumptions upon which off-cycle

stack mass flow is calculated in the DoE/NBS Furnace Boiler Test Procedure [1],

Therefore, no change is recommended to the flue temperature measurement

procedure or to the flue flow calculation for this unit.

Modulation of the firing rate of this boiler resulted in a part load

efficiency that was greater than that resulting from full-fired cycling with

in-phase cyclic pump operation but less than that resulting from full-fired

cycling with continuous pump operation. Because (1) modulation reduced the

efficiency of the boiler when operated with continuous pump; (2) it is

questionable that much modulation would occur with the control system supplied

by the manufacturer (maximum firing at minimum water temperature) in the

70 percent oversized single loop system assumed by the Test Procedure [1]; and

(3) additional tests and calculation complexity are required for seasonal efficiency calculations for modulated equipment, it was decided not to recommend

modification of the Test Procedure to incorporate modulation at this time for

this boiler. It should be noted that continuous pump operation was used in

these laboratory tests as a simplified control system that would show the maximum effect likely from pump delay. It is questionable that continuous pump u

operation would ocdur for any significant time in the home. Units which also

modulate combustion air as well as the gas input would be expected to show substantial efficiency gain [6]. Possible amendment to the Test Procedure [1] to

foster the introduction to the market of such air modulating designs should be

considered. Parallel work covering the development of test procedures for

modulating vented household heaters and furnaces [6] has been completed.

Procedures developed in [6] may be applicable to such an amendment.

44
Laboratory measurements indicated that, for this class of a boiler, the value

of Cj in the Test Procedure should be changed from 4.7 to 1.0. A procedure

for calculating the cyclic portion of from jacket time constants which

could be measured during testing according to the current Test Procedure is

described in Appendix A.

Water pump delay was found to produce a substantial energy saving when applied

to this boiler. Changes to the Test Procedure [1] to allow testing of units

with pump delay are given in Appendix B. These changes require testing at the

same mean water temperature as the existing standard (thus producing comparable

steady-state efficiencies) but at lower outlet and higher inlet water temperatures. It was felt necessary to test at inlet water temperatures near

those actually occurring in the field because of the sensitivity of calculated

annual efficiency to inlet water temperature for pump delay boilers.

. ' I •.

It should be noted that the temporary exemption for this boiler to

be tested at a smaller temperature rise than specified by the Test

Procedure [1] results in an excessively high mean water temperature for comparability to other units. The above modification to the Test Procedure [1]

rectifies this problem as well as allowing fair credit to be given for pump

delay. These recommended changes do not increase the difficulty or time

required for testing and will produce results for conventional boilers that

will be consistent with previous tests.

45
REFERENCES

1. Kelly, George E. , Chi, Joseph, and Kuklewicz, Mark, "Recommended Testing

and Calculation Procedures for Determining the Seasonal Performance of

Residential Control Furnaces and Boilers," NBSIR 78-1543, March 1978.

2. Park, Cheol, Mulroy, William J. , and Kelly, George E. , "A Study of the

Dynamic Flue-Gas Temperature and Off-Period Mass Flow Rate of a Residential Gas-Fired Furnace, NBS Technical Note 999, July 1979.

3. American National Standard for Gas-Fired Central Furnaces, ANSI Z 21.47 - 1979.

4. ASHRAE Standard Measurement Guide: Section on Temperature Measurements,

41.1 - 74, American Society of Heating, Refrigeration and Air Conditioning

Engineers, New York, N.Y. 1974.

5. Chi, J. , "Computer Simulation of Fossil-Fuel-Fired Boilers," Conference

on Improving Efficiency and Performance of HVAC Equipment and Systems for

Commercial and Industrial Buildings, Purdue University, April 12-14, 1976.

6. Kweller, Esher, and Palla, Robert L. , "A Test Method and Calculation Procedure

for Determining Annual Efficiency for Vented Household Heaters and Furnaces

Equipped with Modulating-Type Controls, NBSIR 82-2497, May 1982.

46
APPENDIX A. CYCLIC JACKET LOSS CALCULATION

Cyclic Jacket temperatures were calculated using a procedure modeled after

that used by the DoE/NBS Test Procedure to calculate cyclic flue gas temperatures [1,2]. The derivation and nomenclature used here for the cyclic temperature calculation closely follows that given in greater detail in [2], The

primary difference between the calculation presented here for jacket temperatures and that of the Test Procedure for flue temperatures is the assumption

that the infinite time value of jacket temperature with the unit off is equal

to the room ambient temperature. This assumption was made instead of the Test

Procedure assumption of the infinite time value being equal to that at 45 minutes

after shut-off for cool-down from steady-state because the slow cool-down of

the jacket resulted in the jacket temperature being well above its true infinite

time value after 45 minutes and because, being relatively unaffected by the

standing pilot, the infinite time value was close to the room ambient. A different assumption may, however, be necessary for a fully jacketed unit. As

is shown in Figure 8, the jacket temperature during cool-down from steady-state

closely conformed to the assumption of exponential decay to the room ambient

temperature expressed as:

t e T of f

( 1)

T,RA

T is the steady-state and T the room ambient temperature, and

J jbb -tvA.

T „ is to be determined from cool-down test data,

off

47
A similar expression can be written for jacket temperature during the heat-up

period

:

0

J

(t> - ®J,0,X

e T

0N

( 2 )

where 0j.(t) T - T T (t) , and it is assumed that 0 T _ v = _ „ - T

J,SS J J,0,X J,0,X J,ss RA*

As shown in [2], if the jacket temperatures at time t , and t^ are experimentally

known, the time constant during the heat-up period, , can be obtained as:

r = r- .l2 ~ t

. 1 -,

°N

In

T

J,SS W _T

J,SS" Tj^ t

2^_

(3)

Similarly if the jacket temperature during the cool-down period is measured at

times tj and t^ after the furnace is turned off, T

Q pp

can be obtained using:

OFF

hx

h. - S

- V

T

J

(t4> - T

RA

(4)

When the cool-down and heat-up periods are finite, the pattern of the jacket

temperature changes with respect to time. Due to the finite length of time

for on- and off-cycle periods, the initial values of the exponential function,

0_ . „ and Y, ^ „ become smaller due to the fact that the unit may never heat

up to steady-state or cool down to equilibrium. It is therefore necessary to

introduce correction factors such that for the heat-up period,

48
t

®J

(t) = C

t,ON®J.O,X

6

« (5)

and for the cool-down period.

V£) " Ct,0FF^,0,x

e T°FF ( 6)

The quantities C and C , which are correction factors for the on- and

t y C/JM l j U r F

off-periods respectively, are derived in [2], If the assumption 0 =

J jU J A

¥ = T - T is substituted in the final formulae for C and C J,0,X J,SS RA t,ON t,OFF

given in [2] the resulting equations are:

1 - e

'OFF

''OFF

't ,0N

'ON OFF

1 - e \

r0N

T

OFF/

(7)

't,OFF

1 - e

'ON

r0N

''ON +

''OFF

1 - e \ ON OFF

(8)

Figure 9 compares the experimental cyclic jacket temperature and the predicted

jacket temperature given by

t

T

J

(t) ’ T

J,SS

- C

t,ON

(T

J,SS - Ve T

°N £or on- cy cle . < 9 >

t

T

J

(t) = T

RA

+ C

t,OFF

(T

J,SS • Ve °FF £or off -c5' cle " (10>

49
with the values of C and C obtained from equations (7) and (8) , t,ON t,OFF

respectively.

The jacket loss at a given temperature may be calculated from the equation

Q

= hAAT, where A is the surface area and AT is the temperature differential.

Values of h given as a sum of convective and radiative components are given in

[3] in graphical form for substitution into the equation:

. (a)

h - (h

c + E

where h

£

and h

^ are convective and radiative heat transfer coefficients,

respectively, and £ is the emissivity at the surface.

The following fits were prepared to the graphical presentations of [3].

h

c = 0.380 (Tj - T,^)

0 '^51 for t^e t0 p surf ace ( 12)

h

c = 0.273 (T - T^) 0,249 for the sides (13)

h = 0.200 (T t - T_ . ) °

251 for the bottom (14)

c J RA

h,

£

= (0.664 + 0.0026 T^e0, 00272 T

J (15)

Recommended emissivity (e

)

values are given in [3] for use in equation (11).

The cyclic jacket loss was calculated by using a subroutine to time integrate

the product of the surface area, jacket temperature and h factor for top,

side, and bottom surface using jacket temperature measured during steady-state,

and cool-down and heat-up tests.

50
APPENDIX B

Recommended Changes to the DoE/NBS Test Procedure

for Furnaces and Boilers to Accommodate Testing

of Boilers with Pump Delay

The following recommended changes are indexed to the DoE/NBS Furnace and Boiler

Test Procedure as presented in NBSIR 78-1543 titled "Recommended Testing and

Calculation Procedures for Determining the Seasonal Performance of Residential

Central Furnaces' and Boilers." These changes affect boilers without pump delay

only in specifying a mean water temperature for steady-state testing instead of

a temperature rise and leaving ('outlet) water temperature. This change increases the

range of allowable test conditions to allow closer matching of the test conditions to the manufacturer's design values and greater flexibility in test cell

design without sacrificing comparability. No changes have been made to the

test procedure for furnaces. The recommended changes or additions are shown

underlined in the following paragraphs

:

pg. vi, Nomenclature:

t+

delay time between burner shut-off and blower or circulating pump

shut-off, in minutes

t delay time between start-up and blower or circulating pump start-up,

in minutes

pg. vii, Nomenclature:

y ratio of blower or pump on-time to burner on-time

pg. 4, Definitions:

Mean Water Temperature - The arithmetic average of the inlet and outlet water

temperatures of a boiler.

51
pg. 8 :

2.5.4 Gas- and Oil-Fueled Low-Pressure Steam and Hot Water Boilers (Including

Direct-Vent Systems)

The water flow rate for hot water boilers shall be adjusted to produce a mean

water temperature (average of inlet and outlet), during steady-state operation,

between 115. 0°F (46.1°C) and 145°F (62.8°C) and a temperature rise greater than

or equal to 20°F (11.1°C). The inlet water temperature for units which have

pump delay after burner shut-off and/or which have counter-flow heat exchangers

shall be greater than 120°F (48.0°C). The selected water temperatures shall be

sufficiently high that condensation of the combustion products does not occur

during steady-state operation . For steam boilers, the steady-state performance

test described in 3.1 shall be conducted at atmospheric pressure or at a pressure not exceeding 2 pounds per square inch gauge.

Pg. Us

3.1.2 Oil-Fueled Forced-Air Central Furnaces and Oil-Fueled Low-Pressure Steam

and Hot Water Boilers (Including Direct-Vent Systems)

The furnace or boiler shall be set up and adjusted as specified in sections

2.1, 2.2 and 2.3.4. Begin the steady-state performance test by operating the

burner and the circulating air blower or water pump with the adjustments

specified by 2 . 4.2 and 2.5 until steady-state conditions are attained, as indicated by a temperature variation in three successive readings taken 15 minutes

apart, of not more than: [1] 5°F (2.8°C) in the flue gas temperature and [2]

4 op (2.2°C) in the mean water temperature for hot water boilers.

RcmcUndeA. ol PaAagsiaphA Unchanged

3.1.4 Electric Boilers

Flow conditions shall be as specified in section 2.5.6. Electrical supply

shall be as specified in section 2.3.5. The boiler shall be operated until

52
steady -state conditions are reached, as indicated by a temperature variation

in three successive readings taken 15 minutes apart of not more than 4°F (2.2°C)

,

in the mean water temperature for hot water boilers. Three measurements of the

total power input to the boiler shall be made at 10-minute intervals and

averaged to find the rated power input (E^) , in watts.

pg. 12:

3.2.2 Gas- and Oil-Fueldd Boilers (Including Direct-Vent Systems)

After steady-state testing has been completed, turn the main bumer(s) off and

measure the flue gas temperature at 3.75 (T __,_(t 0 )) and 22.5 (T_ ..--(t,))

F,Url< j F,(Jrr h

minutes after the burner shuts off, using the nine thermocouples described

above. During this off-period for units that do not have pump delay after

shut-off, no water shall be allowed to circulate through the hot water boilers.

For units that have pump delay on shut-off, the pump shall be stopped by the

unit control and the time, t+

, between burner shut-off and pump shut-off shall

be measured within one second accuracy. While the pump is operating the inlet

water temperature and flow rate shall be maintained at the same values as used

during the preceeding steady state test.

T ( t ) If temperature rebound after pump shut-off occurs such that F,OFF 4 is

T (t ) T equal to or greater than F,OFF 3 , or F,OFF (°°) is equal to or greater than

T (t ) +

F,OFF 4 , the time delay for pump shut-off after burner shut-off, t , shall

T

be manually increased or the pump cycled so that F,OFF (3) is greater than

^F,OFF ^4^ and ^F,OFF ^4^ is greater than ^F,OFF (°°) . The time delay t+

shall be increased by this additional pump operation time.

Remainder ofi PaAagaapki Unchanged

53
3.3.2 Gas- and Oil-Fueled Boilers (Including Direct-Vent Systems)

Fifty minutes after the main bumer(s) is turned off for the cool-down test,

the steam or hot water boilers shall be turned on and the flue gas temperature

measured, using the nine thermocouples described above, at 1.0 and

5.5 (T^ ^XT (t 0 )) minutes after the main burner(s) comes on. For units that do

F , ON ^ not have pump delay on start-up, the pump circulating the water through the hot

water boiler shall be started simultaneously with the main bumer(s) . For units

that have pump delay on started by the unit control and the time, t , between

burner and pump start-up shall be measured within one second . The water flow

rate shall be the same as that maintained during the steady-state test described

in section 3.1. During the heat-up test for oil-fired boilers, the draft in

the flue pipe shall be maintained within +0.01 inches of water column of the

manufacturer’s recommended on-period draft. Record the measured temperatures.

pg. 13:

3 . 6 Optional Procedure for Determining ^p* ^F and for Systems Equipped

Power Burners

On power-burner systems not employing automatic stack dampers or power-burner

systems with a stack damper and a draft diverter on draft hood, D^ shall be

measured during the cool-down test described in section 3.2. On systems for

which the flue or stack damper is to be closed during the cool-down test

described in section 3.2, D^ shall be measured during a separate cool-down test.

This separate cool-down test shall be conducted after the heat-up test described

in section 3.3 is completed. It shall be conducted by letting the unit run after

the heat-up test until steady-state conditions are reached, as indicated by

temperature variation in three successive readings taken 15 minutes apart of

not more than plus or minus 5°F (2.8°C) in the flue gas temperature and 4°F

54
(2.2°C) in the mean water temperature for hot water boilers, and then shutting

the unit off with the stack or flue damper controls by-passed or adjusted so

that the stack or flue damper remains open during the resulting cool-down

period. If a draft was maintained on oil-fueled units in the flue pipe during

the steady-state performance test described in section 3.1, the same draft

(within -0.001 and +0.005 inches of water gauge of the average steady-state

draft) shall be maintained during this cool-down period.

Pg- 16:

22. Enter value of y equal to 1 + (t - t )/3.87 or furnaces, y equal to

1 + (t ~ t )/9.68 for boilers and y equal to 1.00 for boilers without

pump delay or furnaces employing a single motor to drive a power burner

and an indoor-air circulating blower.

Pg. 23:

4.2 Recommended Procedure for Calbulating the Annual Cost of Operation of a Furnace or Boiler Located in Different Climatic Regions of the Country

and in Buildings with Different Design Heating Requirements

The annual cost of operating a gas- or oil-fired furnace or boiler located in

various geographic locations of the United States and in buildings with different design heating requirements shall be determined using the following

three-step procedure:

Step 1. Determine the number of burner operating hours using the equation:

Burner Operating House = A (HLH) (C) (design heating requirement) - B (HLH)

where the number of heating load hours HLH, may be obtained from Figure 9 for

the region of interest, the "design heating requirement" is the heating requirement to be met by the furnace or boiler in kBtu per hour at the 97 1/2 percent

55
outdoor design temperature, and C - 0.77 is an "experience factor" which tends

to improve the agreement between the average calculated burner operating hours

and the average burner operating hours found in the field. It is strongly

recommended that this "experience factor" be eliminated as soon as an improved

method is available to more accurately estimate residential heating requirements. Typical values for the design heating requirement are given in Table

for different furnace or boiler output capacities Q 0UT » where Q qut = Ogg (col. 30)

X (col. 4) rounded off to the nearest 1000 Btu/h for units intended for

installation in a heated space and

rounded off to the nearest 1000 Btu/h for units intended for installation out

of doors or in an unheated space. The constants A and B are unique to the

unit under tests and may be calculated using information contained in the work'

sheet and the following expressions:

q (col. 4)

Qout

- (n

ss (col. 30) - 3.3 Lj (col. 18)) rounded off

A = 341,300 (PE + y BE)"+"(Q

IN - Q

p

)n

u

100,000

B =

C2) (A) (Q

p

) (nu)

100,000

1

\ 3.87

for furnaces

where y

= for boilers

1.00 for boilers without pump delay or furnaces employing a single

motor to drive a power burner and an indoor-air circulating blower.

56
U.S. DEPT. OP COMM. 1 PUBLICATION OR 2. Performing Organ. Report No. 3. Publication bate

BIBLIOGRAPHIC DATA

REPORT NO.

SHEET (See instructions) NBSIR- 83-2648 March 1983

4. TITLE AND SUBTITLE

Experimental and Analytical Investigation of a Residential Hot Water Boiler with

Finned Copper Tube Heat Exchangers

5. AUTHOR(S)

William J. Mulroy and Cheol Park

6. PERFORMING ORGANIZATION (If joint or other than NBS, see in struction s) NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

7. Contract/Grant No.

S. Type of Report & Period Covered

9.

SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (Street. City, State, ZIP

)

U.S. Department of Energy

1000 Independence Avenue, S.W.

Washington, D.C. 20585

10.

SUPPLEMENTARY NOTES

| | Document describes a computer program; SF-185, FIPS Software Summary, is attached.

11.

ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant

bibliography or literature survey, mention it here)

In response to a request by a manufacturer of a non typical boiler, the Department

of Energy requested the National Bureau of Standards to perform laboratory measurements under controlled conditions of the effect on seasonal performance of several

features (finned copper tube heat exchanger, water circulating pump delay, and gas

valve modulation) of this boiler that might cause it to be unfairly treated by the

existing test procedure. As a result of this study, recommended changes to the

existing test procedure to allow rating tests with water circulating pump delay are

presented. A recommended change to the assigned cyclic jacket loss factor and a simplified procedure for experimentally determining this factor are also presented.

No change to the current test procedure treatment of gas valve modulation or flue

gas mass flow as a function of temperature are recommended.

12.

KEY WORDS (Six to twelve entries; alphabetical order; capitalize only proper names; and separate key words by semicolon s) Annual efficiency; annual operating costs; boilers; fossil fuel heating systems;

jacket loss; modulating control gas fueled; part^-load performance; rating procedures;

seasonal efficiency

13. AVAILABILITY

[XI Unlimited

|—

|

For Official Distribution. Do Not Release to NTIS

f | Order From Superintendent of Documents, U.S. Government Printing Office, Washington, D.C.

20402.

[X] Order From National Technical Information Service (NTIS), Springfield, VA. 22161

14. NO. OF

PRINTED PAGES

64

15. Price

$ 10.00

USCOMM-DC S043-P80
.

.