NBSIR 80-2145 (DoE) (£) 9

so not mi;:

MAR 2 0 /t/g/

A Study of an Integrated

Appliance, the Air Conditioner/Heat

Pump-Heat Recovery Unit-Water

Heater

King-Mon Tu

Alan Davies

Sidney Fischler

Center for Consumer Product Technology

National Engineering Laboratory

U.S. Department of Commerce

National Bureau of Standards

Washington, DC 20234

Final Report

November 1980

Issued February 1981

repared for

United States Department of Energy

and the

Oak Ridge National Laboratory

Oak Ridge, Tennessee 37830

operated by Union Carbide Corporation

the United States Department of Energy

ontract No. W-7405-eng-26)

NBSIR 80-2145 (DoE)

A STUDY OF AN INTEGRATED

APPLIANCE, THE AIR CONDITIONER /HEAT

PUMP-HEAT RECOVERY UNIT-WATER

HEATER

King-Mon Tu

Alan Davies

Sidney Fischler

Center for Consumer Product Technology

National Engineering Laboratory

U.S. Department of Commerce

National Bureau of Standards

Washington, DC 20234

Final Report

November 1980

Issued February 1981

Prepared for United States Department of Energy

and the

Oak Ridge National Laboratory

Oak Ridge, Tennessee 37830

Operated by Union Carbide Corporation

for the United States Department of Energy

(Contract No. W-7405-eng-26)

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

NATIONAL BUREAU OF STANDARDS, Ernest Ambler, Director

TABLE OF CONTENTS

Page

Abstract iv

Summary v

I. Introduction 1 II. Performance of Existing Systems 3

A. Typical AC/HP-HRU-WH System Operation 3

B. Apparatus and Test Setup 4

C. Test Procedure 7

D. Test Results and Analysis 7

1. Figures for System 1 8

2. Figures for System 2 10

3. Tables for Systems 1 and 2 11

E. Conclusions 12

III. Estimation of Seasonal Energy Costs and Savings 12

A. Introduction and Principal Findings . 12

B. System Description and Operation 14

C. Simulation Methodology 17

D. Results 19

E. Alternative Estimation Methods 20

Tables

II— 1 : Time-Averaged Conditions and Thermodynamic Performance Data for

Heat Pump 1 by Itself and as Part of Integrated HP-HRU-WH

Systems 1 and 2 22

I I -2 : Comparison of Heat Pump Capacity, Power and Efficiency Performance

Data for Systems 1 and 2 with Available Data Supplied by

Manufacturers 23

I

I3 : Heat Recovery Rate, Thermal Energy Rate, Energy Consumption Rate,

and Heat Recovery Efficiency for Integrated Systems 1 and 2 . . 24

I I

I

1 : Tank Size Simulation Results 25

1 1 1-2 : Alternative Calculation Comparisons 26

Figures

1-1 : Major Components of a Heat Pump 27

I- 2: Pressure versus Enthalpy Schematic for Refrigerant Cycle in a Heat Pump 27

I

I—

1 : Detailed Sample Pressure versus Enthalpy Diagram 28

1 1 -2 : Schematic of Integrated Systems 1 and 2 29

I I -3 : Schematic Representation of the Heat Recovery Unit No. 1 . . . . 30

l
Figures (cont.)

1

- 4 :

1 1 -5

1

-6 :

I I-7A-C

I I-8A-C

I I-9A-C

II-10AI 1-11 AII-12AII-13AII-14AII-15AII- 16AIII- l:

III-2a1 1

- 3 :

1 1 1 -4

Schematic Representation of the Heat Recovery Unit No. 2 . . . .

Schematic for Integrated System 3

Schematic Representation of the Heat Recovery Unit No. 3 ... .

: Water Temperatures at Various Locations in the WH versus Time

of Operation of the HP-HRU-WH Integrated Appliance, using

HRU No. 1

: Heat Recovery versus Time of Operation of the HP-HRU-WH

Integrated Appliance, using HRU No. 1

: Refrigerant Temperatures versus Time of Operation for the

HP-HRU-WH Integrated Appliance, using HRU NO. 1

C: Refrigerant Pressures versus Time of Operation of the

HP-HRU-WH Integrated Appliance, Using HRU No. 1

C: Energy Consumed versus Time of Operation of the HP-HRU-WH

Integrated Appliance, using HRU No. 1

C: Water Temperatures at Various Locations in the WH versus Time

of Operation of the HP-HRU-WH Integrated Appliance, using

HRU No. 2 ,

C: Heat Recovery versus Time of Operation of the HP-HRU-WH

Integrated Appliance, using HRU No. 2

C: Refrigerant Temperatures versus Time of Operation for the

HP-HRU-WH Integrated Appliance, using HRU No. 2

C: Refrigerant Pressures versus Time of Operation of the

HP-HRU-WH Integrated Appliance, using HRU No. 2

C: Energy Consumed versus Time of Operation of the HP-HRU-WH

Integrated Appliance, using HRU No. 2

Modified AG/HP-HRU-WH Integrated Appliance System

c: Heat Status versus Time of Day for Washington, D.C.; a and b:

different days; b and c: same day

Heat Recovery Factors versus Outdoor Temperature

NBSLD House Heating and Cooling Loads versus Outdoor DryBulb Temperature

1 1

- 5 : Simulation Outline Diagram
Page

Bib! iography

Appendix A:

AI:

All:

AI II:

A I V

Simulation Background and Related Data . . House Heating, Cooling, and Hot-Water Loads

Heat Pump and Heat Recovery System . . . . Simulation Processing

Additional Temperature Profiles

Table AI : Water Draw Schedule

Table A2: Heat Pump Capacities (CAP), Heat Pump Coefficients of

Performance (COP) and Water Heating Factors (WHF) . ,

Figures

AI : Computational Flow Diagram for Hourly Simulation Method

A2: Computational Flow Diagram for Bin, Average and Bin, Linear

Methods

A3-A13: Bin Hours per Year versus Dry-Bulb Bin Temperatures for Andrews

AFB, MD: Dobbins AFB, GA; O' Hare IAP, IL: Everett, WA; Hanscom

Field, MA; Homestead AFB, FL; Davis-Monthan AFB, AZ; Kansas City,

MO; Carswell AFB, TX; Monterey NAF, CA; and Los Angeles, CA . . . .

\

i i i
Abstract

Three integrated heat pump-heat recovery unit-water heater appliances were

tested under various environmental conditions to measure the functional

parameters and study the operating characteristics of these systems. It was

found that the heat recovery, heat recovery rate and heat recovery efficiency

were dependent on the heat recovery unit's characteristics. The use of the heat

recovery unit in the system resulted in a reduced work load for the heat pump's

compressor and slightly improved the heat pump's performance.

A computer simulation model of the integrated system was developed to study

the interactions between several of the pertinent system variables on an hourly

basis for selected situations and to estimate energy savings. Two alternative

estimation methods that utilize five-degree temperature bin data were also

developed. The estimated savings determined by using the alternative methods

were about the same as those estimated using the hourly data. Conclusions were

also reached concerning the use of water heaters with different tank capacities

and on methods of increasing potential energy savings.

i v
Summary

Three different integrated heat pump-heat recovery unit-water heater

appliances were tested under various environmental conditions in order to

measure and study the operating characteristics and parameters of these systems.

Tests were conducted on the same heat pump and water heater with two different

heat recovery units to determine the heat recovery, heat recovery rate and heat

recovery efficiency. For these same two integrated systems, the heat recovery,

heat recovery rates and heat recovery efficiencies were substantially greater

during the cooling mode operation than during the heating mode operation. The

heat recovery units were installed and used in the systems as received from the

manufacturers. No adjustments were made to optimize or otherwise alter the asreceived performance, Since the third system was discovered to he operating

improperly, only a limited discussion of this system is included in this report.

It was also determined that the heat recovery units had no adverse effects

on the heat pumps' performance. In fact, they slightly improved the performance

by reducing the work load on the compressor.

A modified integrated appliance system that utilizes a water heater with an

upper electrical resistance heating element and a heat pump with supplemental

resistance heat was analyzed to estimate energy savings potentials. A computer

simulation model was developed to study the interactions among several of the

pertinent system variables on an hourly basis for selected situations.

Information generated by the computer simulations for four locations in the U.S.

indicated yearly point-of-use savings in the range 800-1200 kWh, representing

40-60 dollars for an electricity cost rate of 5 cents per kWh, for a particular

residential application. Since hourly data are available for few locations,

specific house characteristics and operating strategies, two alternative

estimation methods were developed that use five-degree temperature bin data,

which are available for numerous locations throughout the U.S.

It was also concluded from the analysis that: (l) the use of a 52 gallon

water heater tank would permit greater flexibility in scheduling residential hot

water usage than would a 40 gallon tank but it would add about 5 dollars a year

in operating costs due to jacket losses; (2) the use of tanks with more than a

52 gallon capacity increase energy costs without providing sufficient additional

benefits; (5) energy savings could be increased by controls which account for

the thermal efficiency of the heat recovery unit and utilizing ventilation when

it can achieve and/or maintain satisfactory indoor conditions; and (4) the use

of averaged weather data for energy estimation and decision purposes, by

consumers, is justified.

v

I. Introduction

The work described and analyzed in this report was performed by the Product

Performance Engineering Division (PPED) of the Center for Consumer Product

Technology (CCPT), National Bureau of Standards (NBS), in partial fulfillment of

"Task B, Air Conditioner/Heat Pump - Water Heater Project." Funding for and

management of the project, in accordance with Interagency Agreement DOE NO. EW78-1-05-5750, was provided by the Department of Energy through the Oak Ridge

National Laboratory (ORNL).

The purposes of the project were to gain a better understanding of AC/HPHRU-WH systems and how they may contribute to energy conservation, and to

identify areas in which potential improvements may result in increased energy

savings.

The "Air Conditioner/Heat Pump - Water Heater" specified in the task title

is a new and currently available integrated major heme appliance whose principal

components are a central air conditioner (AC) or heat pump (HP), a heat recovery

unit (HRU), and a storage type water heater (WH) . In view of these principal

components, the integrated appliance will hereinafter be abbreviated as AC/HPHRU-WH or, when a heat pump is used, as HP-HRU-WH.

The purpose of integrating the three principal components is to construct

an appliance that will use less energy than would ordinarily be used by the

AC/HP and the WH. In the AC/HP-HRU-WH, the AC performs its normal cooling

function and the HP its normal heating and cooling functions, and the WH stores

and provides potable hot water for normal use. The HRU is the added unifying

component that makes energy savings possible; its function is to provide the

means for and to control the transfer of thermal energy from the superheated

refrigerant gas discharged by the compressor of the AC/HP to the water in the WH

tank. Figures 1-1 and 1-2 show, schematically and respectively, the major

components of an AC/HP and a typical pressure - enthalpy diagram for the

refrigerant cycle in an AC/HP.

The HRU has two inlets,' one for superheated refrigerant gas that is

discharged from the compressor of the AC/HP and the other for water from the

drain outlet of the WH, and two outlets, one to allow the refrigerant gas to

flow to the condenser of the AC/HP, and the other to allow the water to flow to

the hot water outlet of the WH. Inside the HRU are a counterflow heat exchanger

(HX) and the gas and water .flow control system.

Utilization of thermal energy that would otherwise be wasted by discharging

it into the outdoor air during cooling mode operation results in energy savings

as less conventional energy is required to attain and maintain the desired water

temperature in the WH tank. As a result of the advertized claims of energy

savings and user interest in saving energy, thousands of AC/HP-HRU-WH systems

have been installed in residences, supermarkets, restaurants, hotels, hospitals

and other structures. Some of the installed AC/HP-HRU-WH systems are of the

retrofit type, and others are supplied as complete or almost complete systems.

In the retrofit installation, an appropriately sized HRU is added to interface

with a previously installed AC/HP and WH. In the complete or almost complete

1
system installation, an entire AC/HP-HRU-WH system is installed, or an integrated AC/HP-HRU system is interfaced with an existing WH. Few companies

manufacture the entire AC/HP-HRU-WH system or an integrated AC/HP-HRU system.

Many companies manufacture an assortment of HRUs for retrofit purposes.

The obvious goal in using the AC/HP-HRU-WH integrated appliance in

commercial or residential applications is to optimize energy usage, i.e.

maximize energy and dollar savings, while performing the required cooling or heating functions and satisfying normal hot water needs. However, there are

many technical factors involved in the operation and assessment of potential

energy savings of such systems that have not been fully addressed and,

therefore, require investigation. Some of these factors are ambient conditions;

incoming, storage and operating water temperatures; hot water draw schedules;

water storage tank size; and the heat recovery unit design.

Ambient conditions have a significant effect on the thermal energy

available for heating water. During the cooling season, for example, higher

outdoor temperatures result in greater quantities of thermal energy available

for recovery. During the heating season, thermal energy available for heating

water decreases.

Incoming water temperatures are higher in southern than in northern states;

storage temperatures, in the water heater, are dependent upon the settings of

the water heater thermostatic control; and operating water temperatures will

depend upon the user.

Hot water draw schedules vary from application to application dependent

upon the number of users and appliances. Furthermore, even draw schedules for a

particular application may vary daily, and from month to month, depending upon

any number of water usage variables, such as for bathing, cooking, dishwashing

and clothes washing, and different hot water temperature preferences.

Water storage tank size will also affect the energy savings potential.

Since the storage tank is to provide adequate amounts of water at sufficiently

elevated temperatures, it is important to select a WH tank size that will have

sufficient capacity with an appropriate backup heater to handle ma.ior draws. If

the heat recovery unit in the AC/HP-HRU-WH is designed to heat large quantities

of water using a high flow rate, the increase in water temperature will be

slight. However, if the design is to handle smaller flow rates, the increase in

water temperature will be large. In either case, a prolonged draw rate may

result in depletion of the hot water if the storage tank capacity is too small

for the application.

The HRU must be properly designed for effective heat recovery. Some

manufacturers of HRU have claimed that their use in the integrated appliance not

only provides "free heat” for heating water, but it also improves the capacity

and efficiency of the AC or HP. However, others have indicated that such

improvement is marginal at most. Manufacturers of compressors used in AC and HP

have expressed concerns about the possibility that use of the HRU may increase

compressor head pressure which may lead to premature compressor failure.

2
Two different types of HP and three different HHJ were selected for this

investigation because they were judged to be typical of the integrated AC/HPHRU-WH systems presently in the field. When operated in the cooling mode, the

HP is considered to be an AC. These systems are described in the section

dealing with the apparatus and test setup.

The investigation consisted of two major phases. One was an experimental

phase during which the HP was operated with and without the KRU and WH. This

was to verify the manufacturers data for the HP, to determine how the integrated

appliance performed, and to provide the technical basis on which a test

procedure for evaluating the performance of the HHU may be developed. In the

other phase a computer model was developed to study the interactions between

several pertinent system variables for selected situations. Analytic methods

were developed to utilize the performance data in estimating and optimizing

energy savings.

II. Performance of Existing Systems

A. Typical AC/HP—HHJ—-WH System Operation

A pressure-enthalpy diagram for a refrigeration cycle is presented in

Figure II-l. The enthalpy differences h

c-h

B

, h^-h^, , and h^-h^ represent,

respectively, the cooling capacity, the compressor work input and the heating

capacity. If losses are ignored, the thermodynamic coefficient of performance

for cooling, C0Pp , is calculated as

cop

c

= h

c h

B

h

D~

h

C

and for heating, COP^, is calculated as

h

D~h

A

h

D“

h

C

The superheat portion for the refrigerant in this cycle, following

discharge from the compressor, is between points D and F in the figure. Typical

refrigerant temperatures are 220°F at point D, the compressor discharge, and

115 F at point F, the condensing temperature. The amount of thermal energy that

might be utilized in heating water is determined by the enthalpy difference,

(h -hj , times the refrigerant mass flow rate. If the controls in the HHJ

confine the water flow only to desuperheating the refrigerant gas, then the

quantity of thermal energy available for heat recovery would be reduced to (h^-

h-j-,) times the refrigerant mass flow. An important parameter that is used in

Section III of this report concerns the thermal energy available for water

heating and is called the "water heating factor": it is defined as the ratio of

the heat available for water heating per pound of refrigerant to the

3
corresponding cooling or heating effect. Its variability with respect to

outdoor dry-bulb temperature is shown in Figure III-3a.

That part of the cycle between points G and A represents refrigerant

subcooling. In passing through a metering device, generally a capillary tube or

thermostatic expansion valve, the refrigerant pressure drops from the high side

(or discharge) pressure, at point A, to the low side (or suction) pressure, at

point B. From point B to point B' , the refrigerant passes through the

evaporator absorbing thermal energy and vaporizing. Eetween points 3' and C the

refrigerant is superheated (about 15°F in the example shown) before entering the

compressor. The compressor raises the refrigerant pressure and temperature from

point C to point D.

B. Apparatus and Test Setup

The performance of each of three HP-HRU-WH integrated appliances was

investigated in this phase of the study. The common major components of the

first two integrated appliances, or systems, are a HP (cooling capacity of

35. 000 and 33, OCX) Btu/h at outdoor temperatures of 95 and 75°F, respectively;

and heating capacity of 34,000 and 39,300 Btu/h at an outdoor temperature of 40

and 50°F, respectively) and a nominal 52 gallon storage type V/H whose electric

heating elements were disconnected for this work. The third major component of

the first system was a retrofit HRU intended for use with an AC or HP rated for

18.000 to 60,000 Btu/h. The third major component of the second system was an

HRU produced by the manufacturer of the HP described above and intended for use

with any AC or HP rated for 18,000 to 42,000 Btu/h.

A schematic representation of the these two integrated systems, referred to as System 1 and System 2, are presented in Figure I 1-2. The figure shows an

indoor unit, called an air handler, in the "indoor room" of a three room

environmental chamber. This chamber was set and maintained at a nominal 80°F

and 50^ relative humidity (RH) when the HP was operated in the cooling mode, and

a nominal 70°F and 50/6 RH when the HP was operated in the heating mode. The

coil in the indoor unit served as the evaporator coil during cooling mode

operation and the condenser coil during heating mode operation. The outdoor

unit was located in the "outdoor room" of the environmental chamber. This

chamber was set and maintained at nominal 95, 85 and 75°F, and 50^ RH, during

cooling mode operation, and a nominal 20°F during heating mode operation. The

reversing valve, accumulator, compressor and capillary tube of the outdoor unit

are identified in the figure. The coil in the outdoor unit served as the

condensing coil during cooling mode operation, and the evaporator coil during

heating mode operation. The ,rbasement room" of the environmental chamber, which

was set and maintained at a nominal 65°F and 50^ RH, contained the HRU and the

V/H. Water flow direction is shown in the "basement room", and refrigerant flow

direction is Shown in all rooms for cooling mode (black arrows) and heating mode

(white arrows) operations.

Schematics of the HRUs used in System 1 and System 2 are presented in

Figures II-3 and II-4, respectively. The water and refrigerant inlets and

outlets of the HRU are appropriately identified. For both systems, the

refrigerant flow through the heat exchanger in the HRU is governed by the

condenser and evaporator temperatures seen by the compressor.

4
In System 1, the water circulating pump of the HRU operates when the HP

compressor operates and the inlet water temperature sensed by the thermally

controlled switch is less than about 150°F. When the inlet water temperature

rises to about 150°F, the switch deactivates the circulating pump. When the

inlet water temperature decreases to about 140°F, the switch reactivates the

circulating pump. Although the circulating pump is the driving force for

circulating water, the water temperature at the modulating' flow control valve

must exceed some set value which in this case was 140 + 5 F before the valve

opens allowing the water to flow^ For this HHU the control valve temperature is

settable in the range of 120-160°F. The circulating pump has a sealess type

impeller with magnetic coupling to the drive motor which has thermal overload

protection; it is rated at 1/200 HP, 1725 RPM, 230 V, 60 Hertz, 1 phase and 0.26

amps. Although the maximum rated flow rate of the circulating rump is 3 gallons

per minute, the actual flow rate is intermittent and modulated from less than

0.1 gallon per minute to the maximum by the temperature sensitive flow control

valve. The heat exchanger in HHJ 1 is well insulated, being completely

encapsulated in polyurethane foam insulation.

In System 2, the water circulating pump of the HRU operates whenever the HP

compressor operates. The flow of water is controlled by a temperature sensitive

modulating flow control valve. This valve, commonly called a reverse acting

valve, opens when the temperature at its sensor reaches the control temperature,

which can be set between 130 and 160°P. When the water temperature at the valve

equals or exceeds the set temperature, the valve allows the water to flow. For

more rapid thermal response, the valve contains a small internal bleed port,

through which a small steady stream of water flows over a temperature sensitive

bulb. Another important component in this HRU is the 3-way flow control valve,

which directs the incoming water to either the heat exchanger or the heat

exchanger by-pass loop, or both, depending upon the temperature of Jhe inlet

water. The recommended temperature set point for this valve is 160 F. When the

inlet water temperature is below 160°F, it flows through the heat exchanger;

between about 160-170°F bgth paths are open; and when the inlet water exceeds

the set-point by about 10 F, the water by-passes the heat exchanger and will not

extract any heat from the hot refrigerant gas. The heat exchanger in HRU 2 is

enclosed in a fiberglass insulated compartment that is open on one side and

allows venting of the HRU through louvers in the HRU casing.

In both Systems 1 and 2, the heat exchanger is of the counterflow tube-intube type. Hot refrigerant gas passes through the outer tube and water flows

through the concentric inner tube which has a convoluted outer surface for

increased heat transfer on the refrigerant side.

The schematic representation of the third integrated system investigated is

shown in Figure 1 1-5. The HP in this system contains a thermal expansion valve

and a capillary tube for metering the refrigerant flow during heating and

cooling mode operation, respectively. In addition, the heat pump is a single

package unit that mounts completely within the outdoor chamber. The rated

cooling capacity of this HP is a nominal 36,000 Btu/h. A schematic of the HRU

is shown in Figure II-6. The water circulating pump of the HRU operates only

when the HP compressor operates. When water is circulating at a temperature

below the set temperature, the two solenoids valves are normally open (N.O.) and

5
normally closed (N.C.), as shown in Figure II-6. When the temperature of the

leaving water exceeds the set temperature it causes the temperature sensors on

the water outlet line to reverse the two solenoid valves positions and the hot

refrigerant gas to by-pass the heat exchanger. When the temperature sensor on

the refrigerant gas outlet senses the set temperature of the gas in the gas

outlet line it deactivates the circulating pump. It should be noted that all

flow control set temperatures may be manually adjusted. The heat exchanger in

this HRU is of the counterflow 2-tube type.

For Systems 1 and 2 a twelve foot 0.5 inch O.D. copper tube was used to

connect the refrigerant outlet from the HP compressor discharge to the

refrigerant inlet of the HRU. A similar length of tubing was used to connect

the refrigerant outlet of the HRU back to the HP reversing valve. By-pass

valves were installed near the compressor discharge to permit the HRU to be

switched either into or out of the system. For System 5, two 25 foot 0.5 inch

O.D. copper tubes were used to connect the inlet and outlet of the HRU to the

refrigerant lines of the HP; no HRU by-pass provisions were made for this

system, since such provision is an integral part of the HRU used in this system.

For all systems a six foot 0.5 inch O.D. copper tube was installed to

connect the water inlet of the HRU with the drain outlet of the WH, and a ten

foot 0.5 inch O.D. copper tube was installed to connect the water outlet of the

HRU with the hot water outlet of the WH. All of the interconnecting tubing

between the HP, HRU and WH was insulated with RF-16 Rubatex FS 4005, 5/8 ID x

5/8” wall thickness.

For Systems 1 and 2, sheet metal ducting was installed on the indoor air

handling unit to obtain the manufacturer's specified airflow rate and pressure

drop across the indoor coil. For System 5, the airflow rate through the indoor

side was adjusted to the manufacturer's specifications by use of a baffle

installed in the indoor section. The proper airflow pattern was obtained by

using an air deflector.

Standard copper-constantan thermocouples were used to measure the

following: (l) dry-bulb and wet-bulb temperatures of the air on the downside of

the coil within the indoor unit and at the air inlet and outlet of the HP indoor

unit, (2) dry-bulb temperatures in each of the three environmental compartments

("outdoor", "indoor" and "basement"), (5) water temperatures entering and

leaving the HRU, (4) water temperatures at 6 different levels in the WH, and (5)

temperatures at various locations along the refrigerant flow line.

Calibrated pressure gauges were used to measure high side and low side

refrigerant pressures in the refrigerant line-. Calibrated watt-hour meters were

used to record the electrical energy used by (l) the compressor and outdoor fan

and (2) the indoor side blower. The locations of the thermocouples, pressure

gauges and air flow meters are indicated in Figure II-2 for Systems 1 and 2.

The six thermocouples in the WH are attached to a "thermocouple tree"; their

junctions were located at the level of the central horizontal plane of each of 6

non-overlapping but approximately equal volumes from the top to the bottom of

the WH tank.

6
Finally, after making all refrigerant line connections to the HRU,

refrigerant R22 was added to the system in accordance with the HRU

manufacturer's instructions. During this refill, all valves between the HP

compressor and reversing valve were open and the "indoor", "gutdoor" and

"basement" room temperatures were equal, at approximately 75°F.

C. Test Procedure

The test procedure generally followed during the experimental phase of this

study was to begin operation of the HP with the HRU in the line (not by-passed)

and record the relevant data every half during the first 5-5 hours of HP

operation, after which time the HRU was by-passed. With the HRU by-passed, the

HP operation was continued for an additional 1.5 hours with measurements made

each half hour before the run was terminated.

Heating elements in the water heater were disconnected and no water was

drawn during the test periods. The indoor conditions were controlled to

maintain nominal values of 50 percent relative humidity with an indoor room

temperature of 80°F during cooling mode operation and 70°F during heating mode

operation. The series of tests for each of the systems was

orun in order with a

nominal "outdoor" temperature of 20, 95, 85, and finally 75 F.

D. Test Results and Analysis

The data recorded during the experimental investigation of integrated HPHRU-WH Systems 1 and 2 were used to develop Figures II-7A-C through II-16-A-C,

which depict water temperatures, heat recovery, refrigerant temperatures,

refrigerant pressures, and energy consumption as a function of time during

continuous HP operation. In addition, these data were also utilized in the

development of Table II-l, which presents thermodynamic performance data for the

HP with and without the HRU in the refrigerant line; Table I 1-2, which presents

HP capacity, power, and efficiency performance data with and without the HRU in

the refrigerant line; and Table II-5, which presents mean heat recovery rate,

thermal energy rate, energy consumption rate and heat recovery efficiency.

Furthermore, it should be noted that due to time constraints of stabilizing

the environmental chamber in which the tests were conducted and limitations on

personnel availability it was not possible to carry out any runs to the point

where the HRU control systems would shut down the heat recovery operation.

Thus, the design cutout temperatures were never determined. This problem could

have been alleviated by using a smaller water storage tank, but this would not

have been consistent with the objective of determining the time required for a

complete recovery with a standard 52 gallon storage type water heater. The

recovery time for the two systems tested is obviously greater than 5*5 hours,

the maximum of any of the runs.

It was found that there were unexplanable inconsistencies in the data for

the cooling mode run with nominal outdoor temperature of 85°F for System 1 and

the heating mode run for System 2; thus, no data for these runs are presented.

Eefore discussing the figures and tables, an explanation of why they do not

include System 5 is appropriate. While running System 5, it v.as apparent that

7
the heat recovery was much less than expected. However, the runs were continued

in the belief that the system was performing as it should and that its heat

recovery performance was just poor. This unsatisfactory performance, which was

highlighted during the analysis of data, led to a trouble shooting investigation

of why the performance was so poor. From this investigation it was determined

that the automatic by-pass control, which by-passes the refrigerant around the

heat exchanger when the water in the tank is sufficiently hot, operated

incorrectly during these runs, resulting in invalid test data. In other words,

the poor heat recovery had resulted from the refrigerant by-nassing the heat

exchanger when it should not have, due to a faulty automatic by-pass control.

Since the data taken for System 3 are invalid they are not presented or further

discussed in this report, nor is it planned to repeat these tests at this time.

1. Figures for System 1

Figures II-7A-C show the water temperatures within the storage type

WH as a function of time during continuous HP operation over a period of 5 or 5*5 hours. As may be expected, the system provides more hot water at

usable temperature when the outdoor temperature (T ) is higher than when it

is lower. When T = 20°F, little usable hot water is provided. In

addition, the figures also show the thermal stratification in the water

tank and the decrease in the time required for all of tbe water in the tank

to reach an "equilibrium" tempera.ture, of about 140-145°F, as the T

q increases. The cut-in and cut-out temperatures are the maximum and minimum

temperatures respectively, at the water outlet of the FRU. These

temperatures are influenced by the location and temperature setting of the

flow control valve, its operational temperature tolerances, the heat

exchanger insulation and heat transfer characteristics, the lengths of

piping between the heat exchanger and the flow control valve, and the

insulation on the copper tubing within the HHJ. As the water in the heat

exchanger increased in temperature, the water near and at the flow control

valve also increased in temperature, by thermal conduction. When the flow

control valve sensed a high enough temperature it opened. Since the

temperature of the water between the flow control valve and the heat

exchanger and that in the heat exchanger was higher than that at which the

flow control valve opened, flow continued until ^ter passing through the

heat exchanger gained insufficient thermal energy to maintain a temperature

high enough to keep the flow control valve open.

Figures II-8A-C show the increase in heat recovery as a function of

time of system operation. The values used to plot the curves were

calculated using the water heater thermocouple tree temperature data taken

each half hour. It is apparent that heat recovery during cooling mode

operation is substantially larger than during heating mode operation.

These curves also serve as a basis for determining some of the heat

recovery rates shown in Table II-3 determined for the period between 1 and

3 hours following the start of system operation, as will be described

below.

Figures II-9A-C show the temperatures of the refrigerant gas

discharged from the compressor (Trnf,), entering the HRU (T^t,), leaving the

HRU , and entering the compressor (T^-g) as a function^of HP

8
continuous operating time. In the heating mode of operation with T = 20°F, the temperature of the compressor discharge gas at the inlet to the

HRU is relatively low, approximately 155°F; thus, the relatively small

magnitude of the refrigerant gas enthalpy in the superheat region results

in a low heat recovery rate. In the cooling mode of operation the heat

recovery rates are much larger than for the heating mode, since the higher

refrigerant gas temperature at the compressor discharge provides a much

higher enthalpy and, consequently, greater quantities of thermal energy are

available for the HRU. After some time of operation, the rise in the

temperature of the refrigerant gas leaving the HHJ (TTM-,-r) reflects a

decrease in thermal energy transfer due to the rise in the temperature of

the water returning to the HRU from the WH tank. This occurred after

approximately 2. 0-3- 5 hours for this system when operating in the cooling

mode, Figures II-9B-C. When the HRU was removed from the refrigerant line,

there appeared to be a slight initial drop, followed by a slight rise, in

the refrigerant gas temperature at the compressor discharge port, Tq^q.*

Figures II 10A-C show the refrigerant pressures at the compressor

discharge (P^), downstream of the 12 -foot gas line from the HRU (Ppp),

just downstream of the condenser (Pp

T

~,), and in the suction line touhe

compressor (P_~) as a function of time of continuous HP operation. The

discharge pressure is more than 50 psi higher for T

q

= 95 F than for T

q = 75°F, while the suction pressures differ by only 10°psi for this 20°F

variation of outdoor temperatures. The pressure drop due to the HRU and

the two 12 foot tubing lines connecting the HRU to the HP is about 5-8 psi

at T = 75°F and 8-10 psi at T = 95°F; it is only 2-4 psi at T

Q = 20°F.

The pressure drop between the compressor discharge and the condenser outlet

decreased from an average of about 17-4 psi with the HRU in the refrigerant

line to an average of about 15-3 psi with the HRU by-passed, during cooling

mode operation. During heating mode operation the pressure difference

averaged about 40.8 psi. Suction pressure generally rose slightly when the

HRU was by-passed.

Figures II-11A-C show the total electrical energy consumed by the

indoor blower, condenser fan and compressor as a function of HP continuous

operating time. As may be noted, the rate of energy consumption shorn in

Table I 1-3 remained essentially constant whether or not the HRU is in the

refrigerant line.

Mean heat recovery rates for the runs at nominal T = 20, 75 and 95°F

for System I were determined from the difference between the total heat

recovery at 1 and 3 hours following the start of system operation, see

Figures II-8A-C, and are presented in Table I 1-3. The thermal energy rate

of the refrigerant gas was determined from the difference between the inlet

and outlet enthalpy and the refrigerant mass flow rate at 2 hours following

the start of system operation. The refrigerant pressure used in the

determination of the inlet and outlet enthalpies was taken as the mean of

the compressor discharge pressure, PpQ

, and the outlet pressure of the HRU,

PpO_, at the 2 hour mark, see Figuresnl-IOA-C. The inlet and outlet

temperatures used in determining the enthalpies were TTM^ and Tp^r

, were

taken at the 2-hour mark, see Figures II-9A-C. Refrigerant mass flow rate

9
was determined from the manufacturer's data for this compressor at the

appropriate evaporator and condenser temperatures.

2. Figures for System 2

Figures II-12A-C are similar to Figures 11-73, C, showing that the

higher the T , the more hot water was nrovided at usable temperatures. The

hot water in°£he WH tank generally reached a firgt pass mean temperature of

about 135-138°F with a cut-in temperature of 140°F as compared to the 140-

145°F mean temperature and 149°F cut-in temperature for System 1.

Obviously, the temperature settings of the flow control valves in System 1

and 2 as set by the two HRU manufacturers, were not identical.

Figures II-13A-C are similar to Figures II-8B,C and were used in the

same manner to calculate the experimental mean heat recovery rates shown

for System 2 in Table 1 1-3.

Figures II-14A-C are similar to Figures 11-93, C. The compressor

discharge gas temperature, T^p, for the T = 75°F run was higher than

expected because of the increased load tha't; resulted from an excessively

high indoor temperature (Table II-l) . The result of this condition was an

increased heat recovery for the T = 75°F run to about the same magnitude

as for the T =

085°F run, as may $e seen by comparing Figures II-13A-B.

For the T =°95°F run, the magnitude of T^py at 5 hours of operation is

much greater than for the other runs, due uo the increased temperature of

incoming water (Figure II-12C) that resulted from a high heat recovery

rate. Thereafter, the small difference between Tpp-^ and T^x resulted in a

decrease in heat recovery, as seen in Figure 11-13(5. Abouts.5-3-5 hours

of cooling mode operation were required before T^^ began to increase.

Figures II-15A-C are similar to Figures II-10B,C. As occurred in

System 1, the discharge pressure of System 2 increased substantially with

T

q

while the suction pressure increased only slightly. The pressure drop

due to the HRU and the two 12 foot lines connecting the HRU to

Q

the HP was

relatively constant a£ 8-10 psi for the runs at T

q = 75 and 85 F and 6-8

psi for the run at 95°F. Removal of the HRU from

0 the refrigerant line

resultgd in an apparent increase in the discharge pressure for the T = 75

and 85 runs and little change for the T = 95 F run. The pressure crop

between the compressor discharge and the condenser outlet, during cooling

mode operation, decreased from an average of about 16.8 psi to an average

of about 8.4 psi, when the HRU was by-passed. Suction pressure generally

increased slightly when the HRU was by-passed.

Figure II-16A-C are similar to Figures 11-113, C. The total energy

consumed by the indoor blower, condenser fan and compressor increased

linearly with time and directly with T . The rates of energy consumption

for cooling mode operation did not noticebly change when the refrigerant

gas was rerouted to by-pass the HRU.

The same methods used for calculating the mean heat recovery rates

for the WH and thermal energy rates for the refrigerant for System 1 were

10
used for System 2. The results of these calculations are also summarized

in Table II-3-

3- Tables for Systems 1 and 2

Table II-l shows the time averaged ambient conditions and

thermodynamic performance data for the HP operating by itself and as part

of an integrated HP-HRU-WH system. As may be noted, the condenser and

evaporator pressures and temperatures, the compressor discharge pressures,

the compression ratios, and the refrigerant mass flow rates were

consistently lower when the HRU were in the refrigerant line than when they

were by-passed. The compressor discharge temperatures, work of

compression, and refrigeration effect were almost always larger when the HP

operated as part of the HP-HRU-WH system rather than as an independent

unit. Finally, the values calculated for the thermodynamic coefficient of

performance during cooling mode operation and the cooling capacity when the

HP operated with the HRU in the refrigerant line were about equal to those

calculated for the HP when it operated as an independent unit.

Table II-2 presents overall performance data, which includes all fans

and pumps, and shows the capacity, power and efficiency of Systems 1 and 2,

and the HP data provided by the IIP manufacturer. Agreement of the

experimental data with the manufacturer’s data for the HP was almost

spectacular when it is considered that this was a reproducibility

experiment which compared a specific unit with a representative average for

the manufacturer's product. As for the average power, it may be noted that

using the HRU resulted in a decrease in power usage. The EER, calculated

by dividing the capacity by the power, almost always increased (mainly due

to the decreased power requirement) when the HRU were used.

These data indicate that the HRU effectively increased the capacity

of the condenser, which led to an improvement in the performance of the HP,

since the increase in pressure drop through the condenser was not

excessive. In addition, the reductions in the compressor discharge

pressure and compression ratio indicate less loading on the compressor,

which would tend to improve the operation and prolong the useful lifetime

of the compressor.

Table II-3 summarizes the mean heat recovery rate, thermal energy

rate, energy consumption rate, and heat recovery efficiency for Systems 1

and 2. The table shows consistently better performance, i.e., greater heat

recovery rate and heat recovery efficiency, for System 1 when the HP is

operated in the cooling rather than the heating mode. The mean heat

recovery rates for the WH and the thermal energy rates for the refrigerant

also appear to increase with increasing outdoor temperature. Furthermore,

it is apparent that the thermal energy rates for the refrigerant are

consistently greater than the mean heat recovery rates for the TwH; the

differences, which average 7.2 and 8.3 percent during cooling mode

operation for Systems 1 and 2, respectively, are to be expected and may be

attributed to generalized heat loss characteristics of the HRU. The

sizable percentage difference for heating mode operation, for System 1, may

11
be attributed to the uncertainty in the determination of the mean heat

recovery rate and the heat loss characteristics of the HRU. Regarding the

heat recovery efficiency, which is the ratio of the mean heat recovery rate

to the energy consumption rate, there appears to be no consistent trend

with varying nominal outdoor temperature: although the heat recovery

efficiency increased from 1.86 to 1.96 Btu/Wh for System 1 operation at T - 75 and 95°F, respectively, it was an average of 1.68 Btu/Wh for System 2

operation at T

q - 75,85 and 95 F.

E. Conclusions

The major conclusions drawn from the experimental phase of this work

are as follows:

1. The HRUs used in HP-HRU-WH Systems 1 and 2 were adequately

controlled so that they produced no discernable adverse effects on

either the HP or the WH.

2. The use of the HRU in Systems 1 and 2 resulted in improved

performance of the HP, mostly due to a decrease in the power

requirement for its compressor.

3- The use of the HRU in Systems 1 and 2 resulted in a reduced work

load for the HP compressor, as shown by the decreased discharge

pressure and compression ratio.

4. The heat recovery rate and heat recovery efficiency for System 1

is substantially lower during HP heating mode operation than during

HP cooling mode operation.

5- To increase the water temperature from 50 to 140°E for a 52

gallon WH tank using a continuously operating HP-KRU-WH system the

equivalent of Systems 1 and 2 would require in excess of 6 hours

during the cooling season and much longer during the heating season.

1

III. Estimation of Seasonal Energy Costs and Savings

A. Introduction and Principal Findings

Seasonal energy costs and savings refer to a full annual cycle of

heating and cooling a house with a HP and meeting hot water requirements.

The savings are the differences between total household electricity usage

with and without heat recovery, expressed in kWh per year. Energy savings

may be converted to dollar savings using local rate schedules for the time

frame of the estimate. Note that marginal cost rates vary with locality,

amount of energy used, power company success in obtaining hi^ier rates

during the estimation period, and in some cases the type or class of

service provided to the particular residence. Hourly simulations were

conducted in which house construction and operation -were held fixed.

12
Weather was varied by using data from Washington, D. C., Atlanta, GA,

Chicago, IL, and Seattle, WA. Estimated savings are in the approximate

range of 800 to 1200 kWh per year, or $40 or $60 oer year at $6.05 oer kWh.

(See Table III-l)

Simulation details about the house, the heat recovery system and

house operation are given below and in Appendix A. Significant operating

assumptions were to set the thermostat for cooling at 10°E above the

setting for heating and to use ventilation where the results would be

acceptable. These assumptions led to a zero house load for more than a

third of the hours per year at each of the tested locations. This

consideration is important because heat recovery is possible only if the

house places a demand on the HP.

Tank size was varied at Washington and Atlanta. Different analytical

methods were compared to test the need for detailed simulations in making

decisions on heat recovery systems.

Principal Findings

1. Tank size selection involves compromise. Increasing the tank size

increases the ability to accept recovered thermal energy whenever it

is offered and increases the amount of "reserve” hot water that is

available when no recovered heat is being provided, but decreases

efficiency by increasing heat losses through the storage tank's

insulating jacket. The ability to accept offered thermal energy was

not a problem under the simulated conditions using tanks of 40

gallons or larger. Eased on a auarter tank reserve, a 40 gallon

system with a good heating rate from its backup resistance heater

might meet dishwasher or clotheswasher demands. A 52 gallon system

would permit more flexibility in scheduling hot water use, at an

additional cost of about $3 per year (at 5 cents per kWh) in jacket

losses. Larger tanks are not particularly attractive, increasing

energy costs without providing important benefits.

2. The fraction of the offered thermal energy that was accepted was at

or near 1.0 for all the cases covered in the simulations. This is

the "decoupling factor" shown in Table III-l. This means that the

tanks were quite successful in accepting thermal energy whenever it

was offered and making it available for later use. In this sense the

time sequence of offerings is decoupled or rendered independent of

the time sequence of water uses. Where time sequences are

unimportant, the hourly data may be arranged in ways that are

analytically more convenient and the benefits of hourly simulations

for system performance estimation purposes are minimal. Holding the

house load profiles fixed, the decoupling factor will remain high for

higher hot water usage. Fixing hot water usage, the decoupling will

also be high for lower house load profiles and for lower HP

capacities.

3- It should be expected that thermal energy recovered during the

heating season will not meet hot water requirements even under

13
favorable conditions. This means that heat from the backup heater in

the WH will be needed on a regular basis and that the tank will

commonly be at or near its '’reserve" status during this season.

4* Some thermal inefficiencies must be expected in the loop connecting

the HRU and the WH. If this loop is less efficient than the heating

element in the tank in providing thermal energy to the water, the

heat recovery system should be shut off when supplemental resistance

heat is being provided to the house.

5. The operating cost of the HP is unchanged by heat recovery operations

during the cooling season, except for the possibility of sane second

order effects involving performance of the refrigerant coils. Total

operating costs during the heating season for the HP, supplementary

heater and backup heater vary with tank size because .jacket losses

depend upon tank size. However, since all the offered heat was

accepted during the heating season, there was no difference in the

combined operating costs of the HP and the supplemental heater with

respect to tank size in the simulated cases. (Coil defrosting costs

were not estimated. The modeling assumption is that heat recovery

operations would be suspended during defrosting and that defrosting

cycles do not interfere with HP heating cycles.)

6. It is possible during some parts of the cooling season to achieve

satisfactory indoor conditions by ventilation instead of keeping the

house closed and depending on the HP. Ventilation is economically

sound since more energy is saved by not running the HP than would be

provided "free" to the WH.

7. The weather at any given location can vary considerably from year to

year and within the service area of a weather data source, with

corresponding effects on HP loads and heat recovery savings. Since

system paybacks will take years, the data and analysis should cover a

significant number of years and the use of averaged weather data in

the estimating procedures is justified.

8. Simplified estimation methods that would be adequate for consumer

decision purposes appear practical. Development of such methods is

suggested.

B. System Description and Operation

A schematic diagram of the integrated HP-HRU-WH system to be discussed is

shown in Figure III-l. The main elements are a HP, a KRU and a conventional

electric WH with its lower heating coil disconnected (and omitted from the

diagram) . The upper electric heater is called the backup heater in the

discussion. A supplemental resistance heater is provided to augment the HP when

the overall demand for heat by the house and the heat recovery system exceeds

the HP capacity. House thermostats control the HP and the supplemental heater.

The circulating pump in the HRU operates steadily when the HP is operating.

Water is allowed to flow from the HRU only when it is hot enough (nominally

145 F in these simulations).

14
The symbol, 0, on the scale next to the tank represents the heat status of

the tank. There will be some mixing between the hot water at the top and the

cold water at the bottom of the tank, but a sharp temperature transition is used

for purposes of exposition. Q increases (i.e., moves down in the diagram) when

the HRU is providing hot water and moves up when hot water is drawn or thermal

energy is lost during standby. The backup heater will come on when Q is less

than the "reserve" hot water level, Ql, and will turn off when 0 reaches 01. If

the HRU supplies enough thermal energy to increase Q to Q9, then the pump will

draw hot water, signaling that the tank has reached its maximum heat capacity,

and no further thermal energy offerings by the HRU will be accented.

Figures III-2a-c show the hour-by-hour heat status of the tank in percent

of heat capacity, Q9, for selected days from the Washington, D.C. simulation.

Figures III-2a and b are for a 40 gallon tank on different days, and Figures

III-2b and c are for 40 and 52 gallon tanks on the same day. Status is shown

after all gain and loss transactions for an hour have been completed.

Therefore, Q is never shown below the reserve status, Ql, and Q is never allowed

to rise above the heat input limit, Q9. Except for extraordinary cases, backup

heat is needed whenever the ending status is less than Ql, and seme of the

offered heat is not used when the ending status is Q9* It is seen that the tank

heat status can be quite dynamic when the HP is active and that larger tanks are

able to accept more of the offered heat than smaller capacity tanks.

Further understanding of the general interactions involving the house

system and local weather may be gained from Figure III-3 , in which several

heat recovery factors are presented in schematic form as functions of the

primary independent variable, the outdoor dry-bulb temperature. The HP

operating regimes of heating and cooling are separated by a zone of zero house

load, in which the HP does not operate.

The water heater factor, WHF, is shown in the top graph, III-3a. WHF is the

fraction of the HP output that is offered to the water tank. WHF rises with

outdoor temperature because the discharge temperature of the refrigerant coming

from the HP compressor increases with the temperature of the outdoor coil

whereas the minimum temperature to which the refrigerant may be cooled in the

heat exchanger is determined by the required output water temperature.

House load and HP capacity vary with outdoor temperature, as indicated in

Figure Ill-Jb. The two capacity curves are for a nominal three-ton HP. House

leads are shown here and in Figare I I 1-4 solely as functions of outdoor

temperature. The general effect of a ventilation option during the cooling

season is shown. The load curve slopes and intercepts on the temperature axis

are determined mainly by house size, construction and operation. Geographic

location may affect the load through wind, air quality, cloud cover and outdoor

humidity characteristics of the locality. Latitude affects the amount of

daylight through the seasons and the incident angle of solar radiation. Houses

have thermal masses that cause loads to lag the forcing functions. Even though

Information from several sources was used to develop Figure III-3, including

Bibliography items 6 and 14, test data, engineering calculations and externally

supplied simulation results.

(2) v nourly house load data and corresponding outdoor dry-bulb temperature were obtained

from another NBS project. These data were sorted according to five-degree temperatui

bins, averaged within bin, and plotted.

15
the load can vary considerably at ary given temperature, the outdoor temperature

can serve alone as an effective indicator of house load for present purposes.

(Estimating methods using outdoor temperatures are discussed in Section E and

numerical results are compared in Table III-2 with those from simulations using

hourly heating, cooling loads and temperature.)

Supplemental resistance heat is needed when the combined house heating and

HRU loads exceed HP capacity. The effect of offering thermal energy to the WH

during heating mode operation is shown as a cross-hatched area, reducing the HP

heating capacity available to the house. This capacity reduction does not occur

during cooling mode operation. If the house load exceeds the HP capacity in the

cooling mode, the indoor temperature will rise appropriately.

There is a ten-to-one scale expansion from III-Jb to III-3c. The domeshaped curve in III-3c labeled "From Heat Pump" corresponds exactly to the heat

flow segment in Ill-Jb labeled "To Water Heater". Figure III-3c shows the

average hourly heat requirement for the WH without regard for time of day, and

the thermal energy sources in the integrated water heating system. Thermal

energy offerings from the HP (the curves labeled "From Heat Pump") are found by

multiplying the WHF by the lesser of the HP load or the HP capacity. The HP

load is the same as the house load in the cooling mode, but includes the thermal

energy given to the water by the HRU and the house load in the heating mode.

Some of the offered thermal energy at the higher outdoor temperatures is more

than can be accepted on a steady-state basis. However, some of this excess

thermal energy can be used if the tank status before the offering is below the

maximum level that is allowed. (High outdoor temperatures accounted for all

cases in the simulations in which some of the offered heat could not be used,

thus reducing the decoupling factor.) The backup heater must be capable by

itself of meeting the full water-heating load when the HP will not be operating.

A sample climate profile in terms of hours per year as a function of

outdoor temperatures is shown in Figure III-3d. This profile is for Washington,

D. C. and the single year of 1957. Profiles of this type for other locations in

the United States are given in Figures A-3 through A-13 of Appendix A. These

latter profiles, representing data over at least ten years, provide a useful set

of weighting factors for judging system attractiveness in various regions. Data

for these curves came from Item 6 in the Bibliography, in which plots for many

more sites are available. The attractiveness of HP-WH systems is greatest at

those locations having many hours per year in the temperature ranges where the

HP offerings (lll-3c) are high.

House load data for the simulations were generated for the same

hypothetical house at four locations, using a particular -weather year at each

location. These load data were sorted into five-degree temperature bins,

averaged within bin and plotted against bin temperature in Figure III-4 for

Washington, D.C. and Atlanta, G-A. Sensible and latent loads were added to

obtain the total load for each hour. Differences in climate and sun elevation

may account for the differences between sites.

Some general inferences that are borne out by the simulations may be

deduced directly from Figure III-3.

16
1 . The WHF declines as the outdoor temperature goes down, for both

heating and cooling applications, reaching zero where the temperature

of the refrigerant ccming out of the compressor is too low to

transfer thermal energy to water in the HRU at the required water

output temperature. Hours at temperatures below these zero points

could be dropped from consideration for purposes of heat recovery

analysis.

2. There is some temperature below which supplemental resistance heat

will be needed, given reasonable sizing of the HP with respect to the

house load profile. The supplemental heater must make up any

shortfall in the heat that the HP supplies to the house. Therefore,

no advantage is seen in giving HP thermal energy to the WH if it must

be replaced by the supplemental heater. The simulation does not

include any allowances for thermal inefficiencies between the HRU and

the WH, and thus treats the energy used by the tank's backup heater

or the supplemental heater as equivalent in the energy totals. Any

such inefficiencies mean that the heat recovery system should be

turned off when the supplemental heater is needed.

3. The backup heater will have to carry much of the water heating load

during the heating season, with the further implication that the tank

heat status will be at or near the reserve level, Ql, during this

season. Referring to the Water Heater Operation graphs in Figure

III-3, the dashed line labeled Average Hourly Requirement is one

twenty-fourth of (64.3 gallons per day x 8.25 pounds per gallon x

' 90°F temperature rise plus 1 percent per hour for heat loss during

standby), or a total of about 2100 Btu per hour. The highest

offerings during the heating season occur in the 25 to 35°F

temperature range and are about half of these average demands. Some

combination of major reductions in water use and increased thermal

energy offerings (by increasing both house heating loads and heat

pump capacity) would be needed to close the gap in this limited

temperature range. Also, where all the thermal energy offered during

the heating season is accepted, the total annual electricity

consumption by the HP and the supplemental heater is independent of

tank size. However, WH jacket losses do vary with tank size,

affecting annual energy use by the backup heater and hence the total

annual consumption of electricity.

4. The practice of ventilation during the cooling season is economically

sound as long as satisfactory indoor temperature and humidity

conditions can be maintained. Some "free” thermal energy to the WH

would be missed, but this less would be outweighed by the savings in

HP operation.

C. Simulation Methodology

There are many variables affecting the amount of thermal energy that is

available from a HP-HRU-WH system and the operating cost of obtaining this heat.

A computer simulation model was developed to study the interactions of several

of these variables on an hourly basis for selected situations. The hourly steps

17
in this simulation are outlined in Figure III-5 and the simulation is described

further in Appendix A. The purpose of the simulation is to generate numerical

information bearing on the economics of HP-HRU-WH systems in terms of major

system variables. Several of these variables and their interactions have

already been introduced.

The thermal performance of houses and other buildings is a significant

study topic at KBS for which a highly detailed computer program, called the KBS

Load Determination, or KBSLD (Bibl. 8), program, has been written. NBSLD was

used on another KBS project (a retrofitting study using 1800 square foot houses

and four-person families) to generate hour-by-hour house load values for

selected locations and weather years. Thermostats inside the house were set at

68°F during the heating season and 78°F during the cooling season. Ventilation

was used when a 68 to 78°F indoor temperature and an indoor relative humidity

below 60 percent could be maintained by this method. Neither ventilation nor

the HP were used when satisfactory conditions could be maintained without than.

The present simulation makes use of these house load data and the corresponding

outdoor temperatures. An averagg hot water usage rate of 64-3 gallons per day

(450 gallons per week) with a 90 F temperature rise was used. The key

parameters of HP coefficient of performance (COP) and output capacity (CAP) and

the water heater factor (WHF) were expressed as functions of outdoor

temperature. (See Appendix A, Table A2.) The COP and CAP values in the

simulation are consistent with KBS laboratory data. An hourly hot water draw

schedule developed by KBS was used for each day of the simulation year. (See

Appendix A, Table Al.) The heat status of the WH at the end of the previous

hour is the beginning status in the current hour.

The computational steps for each hour are:

1. Subtract from the initial heat status, 0, the standby heat loss and

the thermal energy in any hot water that was drawn during the hour to

produce the pre-offer heat status. A standby thermal energy loss

rate of one percent per hour of the initial heat status of the tank

was used.

2. If there is any load on the HP, determine which alternative will set

the limit on the amount of thermal energy that the tank will accept.

a) The tank can only accept thermal energy up to its heat

capacity, Q9.

b) - The maximum thermal energy offering is the product of the HP

capacity and the WHF at the temperature for the hour.

c) If the HP load is at or below the HP capacity, then the thermal

energy offering is determined by the product of HP load and the

WHF. The HP load is solely the house load in the cooling mode,

but includes thermal energy for the WH in the heating mode.

5. Compute the new tank status and the following items for accumulation

and eventual printout:

18
a. Heat recovery unit offering

b. Heat recovery unit output (= water tank acceptance)

c. Eackup heater load

d. Supplemental heater load

e. Heat pump electricity used.

The backup heater is used only when necessary to achieve the minimum (or

reserve) heat status, Ql, at the end of the hour. Calculations of

electricity use by the HP include adjustments for efficiency degradation

during cyclic operation.

4. Go on to the next hour.

D. Results

Results from ten simulation runs are summarized in Table III-l. Two runs

were conducted for each city using systems with 52 gallon tanks, one without

heat recovery (the base case) and the other with heat recovery. These pairs

indicate regional variations in energy use and savings potential. Tank size

effects were also tested with runs for Washington and Atlanta using heat

recovery systems with 40 gallon and 80 gallon tanks. Both the 52 and 80 gallon

tank systems accepted all offered thermal energy. Jacket losses for the 80

gallon tank system were higher than for the 52 gallon tank system, so the 80

gallon results are not presented. The operating rules with and without heat

recovery are identical except that no thermal energy is accepted from the HRU in

the base case. All water thermal energy in the base case comes from the backup

heater and the tank is restored to the reserve status, Ql, at the end of each

hour.

Two system performance measures are provided. The one of greatest interest

to a homeowner is annual electric energy saved , measured relative to a HP and WH

system without heat recovery and in which the lower coil in the TWH is similarlydisconnected. The second will be called the decoupling factor and is the ratio

of thermal energy accepted by the taTT to the thermal energy offered by the HRU.

This factor is an index of the system's ability to decouple the two time series

of thermal energy offering^ and thermal energy uses by storing offered thermal

energy for later use.

Electric energy savings are shown on line M and decoupling factors are

shown on line IT. Savings are affected by tank size to a minor degree through

differences in the energy lost during standby and the ability to accept offered

thermal energy. Larger tanks have larger hot water reserves and therefore have

more hot water exposed to standby losses. Hot water reserves of 10 and 15

gallons were used in the simulation respectively for the 40 and 52 gallon

systems. Line 0, labeled "Ql Hours," shows hours for the year ending in the

reserve status. At least for these hours in common, the 52 gallon system has

19
three more gallons of hot water losing heat during standby than the 40 gallon

system. Eased upon the savings indicated on line M and energy costs of 3*5 to 5

cents per kilowatt-hour, use of the smaller tank would save $2 - ^3 per year

over the larger tank in Washington and Atlanta.

The obvious next question is why not go to still analler tanks? It is seen

that any tank will spend much of its time in the reserve status, 01, and that

major hot water draws such as for washing clothes and dishes must be anticipated

when in this status. The largest single draws used in the simulation are one 15

gallon draw in the morning and another in the evening. Both exceed the 15 or 10

gallon reserves, and could be a basis for insisting on a tank of 60 gallons or larger. However, the tanks under consideration have recovery rates of 17 to 25

gallons per hour, and recovery heat will be available during the actual period

of the draw to reduce or eliminate the possible shortfall. As a practical

matter, the 52 gallon system offers some flexibility in scheduling hot water

use, whereas the 40 gallon system requires more scheduling care. The general

conclusion is that the most economical system uses the smallest tank that has

enough reserve hot water as long as the decoupling factor is high.

The decoupling factors on line N are high for both tank sizes and at all

locations, indicating that, at least for the particulars of these simulations,

there is little problem in making full use of the thermal energy offered by the

HEU. The simulation for Washington indicates a minor inability of the 40 gallon

tank system to take all of the offered thermal energy. Thirteen kBtu that were

accumulated during eight hours (line P) during the year could not be used (line

E-line D). These hours occurred during the hottest part of the summer. The 52

gallon systems were able to accept all of the offered heat in Atlanta and

Seattle, and all but one kBtu in Washington and Chicago.

Unused thermal energy and the Q9 hours increase as tank size decreases and

also as the hours at very high outdoor temperatures increase. The decoupling

factor decreases under these circumstances, making the time phasing between

thermal energy offerings and the water draws more important.

Combined electric energy used by the HP (line I) and the supplemental

heater (line K) is independent of tank size as long as all the unusable thermal

energy is offered during cooling operations. Unusable thermal energy, if any,

offered during heating operations leads to a decrease in the year’s combined HP

and supplemental heater energy use and to increased energy use by the backup

heater.

E. Alternative Estimation Methods

One objective in detailed simulation is to preserve the effects that events

in one time period ( hour-to-hour or day-to-day, for example) have on events in

another period. Another is to provide a baseline from which the acceptability

of alternative analytical methods may be judged. • As will be shown, certain

aspects of the simulation procedure may be reshaped to good effect and without

changing the numerical results enough to affect most consumer decisions.

Hourly data on house loads and corresponding weather data are available for

only a few locations and quite specific house characteristics and operating

20
strategies. This places severe limits on the ability to extend results to other

locations and situations. Therefore, the numerical significance of making two

adjustments to the more rigorous hour-by-hour simulation was examined. The

first of these was to sort the load data by five-degree temperature bins,

compute bin averages and then compute the various heat recovery output results.

Average house loads by bin are shown in Figure III-4 for Washington and Atlanta.

The issues being examined are the importance of specific sequences of hourly

loads and the effects of nonlinear relationships on the results. The second

change was to replace the bin averages of the house loads with loads derived

from linear regressions of the bin averages. If the first adjustment is

acceptable, then annual summaries of temperature data could be used. If the

second adjustment is acceptable, the characterization of houses and their

operations can be much more flexible and parametric studies become more

practical.

Comparative results are summarized in Table III-2 for four locations and

three computation methods. Systems with 52 gallon tanks were used throughout.

The reference case for computing savings is a 52 gallon tank maintained at a 15

gallon reserve level, Ql, by a backup heater. Both the Bin Average and the Bin

Linear methods are based on steady state conditions of heat recovery within each

tanperature bin and do not allow for thermal energy acceptances within the

dynamic range of the tank heat capacity when the average offering rate exceeds

the acceptance rate, accounting for most of the difference in savings among the

methods. Nonlinearities and approximations of constants account for the

remaining differences in savings estimates.

The difference in numerical results among the methods do not appear to be

important with respect to the most important consumer decision on 'whether a heat

recovery system of this kind should be installed or not. The differences are

well within common year-to-year weather variations. Although these few examples

hardly constitute proof, the strong inference is that the simpler computing

methods are adequate. Furthermore, they offer opportunities for parameteric

examination of several house and recovery system variables without expensive and

sometimes unavailable weather data and preparatory lead calculations. The

various curves in Figure III-3 and COP values could be hypothesized, for

example, in order to explore the circumstances of economic feasibility for such

systems. Bibliographic item 6 or its equivalent provides a rich compilation of

weather data, house load curves could be hypothesized, and so on. All of the

reported approaches use discrete weighting and summation methods to produce the

main results. There is an untested possibility that the various curves could be

expressed in continuous algebraic form and dealt with at that level for

sensitivity analysis purposes.

21
Table

II-l:

Time-Averaged

Conditions

and

Thermodynamic

Performance

Data

for

Heat

Pump

1

By

Itself

and

as

Part

of

Integrated

HP-HRU-WH

Systems

1

and

2

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23

nominal

temperature,

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data.
Table 1 1 - 3 : Mean Heat Recovery Rate, Thermal Energy Rate,

Energy Consumption Rate, and Heat Recovery

Efficiency for Integrated Systems 1 and 2

System

Nominal Outdoor Dry-Bulb Temperature

20°F 75°F 85°F 95°F

Mean Heat \ Recovery Rate, kBtu/h ' 112

2.26 7.20

6.80 6.80

8.50

7.40 |

Thermal Energy

Rate, kBtu/h

123.26 7.49

7.34 7.77

9.59 i 7.79 j

Energy Consumption

Rate, kWh/h

123.64 3.88

4.00 4.10

4.28 1 4.40

Heat Recovery /^\

Efficiency, Btu/Wh'

120.62 1.86

1.70 1 .66

1.96

1 .68 —Heating ^ Cooling

Notes :

(1) The Mean Heat Recovery Rate was determined from the difference in total

heat recovery by the water between 1 and 3 hours following the start of

system operation.

(2) The Thermal Energy Rate was determined from the refrigerant mass flow rate

and the difference in the inlet and outlet enthalpy of the HRU 2 hours

following the start of system operation.

(3) The Heat Recovery Efficiency is the ratio of the Mean Heat Recovery Rate

to the Energy Consumption Rate.

24
Table

III-l:

Tank

Size

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25
Table 1 1 1 - 2 : Alternative Calculation Comparisons

Item Method

Hourly Bin Average Bin Linear

Heat Recovered No Yes No Yes No Yes

Location: Washington

,

D.C. kWh for Year: Heat Pump 6554 6979 664/ 7083 6779 7245

Backup Heater 5355 3710 5356 3710 5356 3673

Supplemental Heater 89 104 58 80 36 40

Total 11998 10793 12061 10873 12171 10958

Savi ng NA 1205 NA 1188 NA 1213

kBtu Accepted 0 5772 0 5616 0 5740

kBtu Offered 5773 5775 5777 5777 5906 5906

Decoupling Factor 0 1.0 0 0.97 0 0.97

Location: Atlanta, GA kWh for Year: Heat Pump 4973 5255 5053 5345 4728 5033

Backup Heater 5355 4099 5356 4093 5356 4149

Supplemental Heater 10 14 0 4 0 0 Total 10338 9368 10409 9442 10084 9182

Saving NA 970 NA 967 NA 902

kBtu Accepted 10 4399 0 4309 0 4116

kBtu Offered 4399 4399 4401 4401 4208 4208

Decoupling Factor 0 1.0 0 0.98 0 0.98

Location: Chicago, IL kWh for Year: Heat Pump 8232 8759 8363 8917 8344 8910

Backup Heater 5355 3808 5356 3787 5356 3784

Supplemental Heater 688 731 594 609 451 466

Total 14275 13298 14313 13313 14151 13160

Saving NA 977 NA 1000 NA 991

kBtu Accepted 0 5418 0 5352 0 5361

kBtu Offered 5419 5419 5432 5432 5439 5439

Decoupling Factor 0 1.0 0 0.99 0 0.99

Location: Seattle, WA

(b)

kWh for Year: Heat Pump 5612 6263 5680 6349 5636 6304

Backup Heater 5370 3833 5370 3804 5370 3814

Supplemental Heater 0 0 0 0 0 0 Total 10982 10096 11050 T0T53 11006 T0 TT8 Saving NA 886 NA 897 NA 888

kBtu Accepted 0 537 5 0 5344 0 5309

kBtu Offered 5375 5375' •5374 5374 5338 5338

Decoupling Factor 0 1.0 0 0.99 0 0.99

(a) Results are for 52 gallon systems. Hourly data are from the hour-by-hour

simulation. "Bin Average" data are based on hour-by-hour house loads averaged

for each five degree outdoor temperature bin. "3in Linear" refers to house

load inputs generated by linear regression of "Bin Average" data.

(b) Seattle data are for a leap year of 8784 hours.

26
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O c G 1 c L. JO JO 3 G Li CL u a CC o Li > o *H G LI o G cO 3 3 Li C3 G Li 00 > 3 4-1 Li •H P3 O 3 o G o , s G so SI 3 CL G !-i o CL r- o c C L> G V >> 4-1 L-l 3 TO H U u o *H X u 3 L4 4—1 G PP c •L G

CL

G JO CJ 3 o G

G Li LiG 00

G LU Li s Li D > rH rH rH L Li 3 o o G G 3 'Ll rH 3 LI 3 G TO

•r—> a > aG Li *H

G Cl

G 3 G pp 3 G G J3 O LI a- jO H <—*

CL Li

C

G

v; Li

L 3 3 c G

H

G 5 3 G cr.

G o o - aj G pp •H SO -H j-i

•1 ) (!) rH CO

G

n •U •H TO q.

G [ r“ L-l rH TO

G

G • C • GU

rH s D *H

O 00

G !4 CN 4_>H Li 4-1 1 Li o £ *H Li > i —1 G o G M G c o Li G r—1 • H J-a G pp

rH *H

G 00 L-i -3 TG

A

Gu

a c Li Li 3 3 4-1 > > G PL, 3 G

Li o G

GG •H Li Li 3P PP Li 3 3 G > 00 G rHO G G 3 3

CO 3 u

C >H •H -C JO Li G O JO G *H 3 *H X! O o G H <a '-H TO CJ LI •H G G fa Li u CJ >

27
ABSOLUTE

PRESSURE

(lbs

/in 2)

ENTHALPY (BTU/!b above saturated liquid at -40°F)

-30 -20 5000

scale! CHANGE

-30 -20 0 20 40 60 80 S)0 100 110 120 130 140

Figure II- 1 : Detailed Sample Pressure vs. Enthalpy Diagram

28
29

Figure 1

1—2:

Schematic

of

Integrated

Systems 1

and
SCHEMATIC REPRESENTATIONS

OF HEAT RECOVERY UNITS 1-3

HOT COOLED

Figure 1 1-3: HRU No.l. COLD WATER IN

i

i
C3 C3

31

Figure

II—

5:

Schematic

for

Integrated

System
TC = Thermocouple

HRU = Heat Recovery Unit

WH = Water Heater

HP

c = Heat Pump in Cooling Mode

HP^ = Heat Pump in Heating Mode

T

q = Nominal Outdoor Temperature

TC 1 is 3-7/8” from the Bottom of the WH Tank

TC 2 is 11-5/8” from the Bottom of the WH Tank

TC 5 is 19-3/8” frcm the Eottom of the WH Tank

TC 4 is 27-1/8” from the Bottom of the WH Tank

TC 5 is 34-7/8” from the Bottom of the WH Tank

TC 6 is 42-5/8” from the Bottom of the WH Tank

TC 7 measures the water temperature entering the HRU

TC 8 measures the water Cut-Out Temperature leaving the HRU

TC 9 measures the water Cut-In Temperature leaving the HRU

Addendum to Figure II-7A-C

32
7a HPu T •- 20°F

n U Jx "Pc T0

• 7S°F

T0 95°F

Figures II-7a-c; Water Te.-pgra t'jrss at Various Locations in the 'ffl versus

T

t

me of Operation or the HP - HftLi - 'wH 1 utegra ted APPLIANCE/ L'S IDG HRU ilo . 1

33
HEAT

RECOYERY

(kBtu)

8c HP

C T

q = 95°F

Figures 1I-3a-c: Heat Recover y v ersus T ime of Operation of the

I IP - HP. Li - nil I ,m i: t p a r t D Appliance , ilsifio n'RU lie, 1

34
TIME (h) TIME (h)

1 2 3 4 5 6 7 TIME (h)

Figures II- 3a-c: Refrigerant Temperatures versus T r

r

e of Oration for the HP - HRU - V/H I ‘ITEGRATED APPLIANCE, USING HRU No . 1 Notes: Tcdg = Temperature of Compressor Discharge Gas

I RGE = Temperature of Refrigerant Gas Entering HRU

Trgl = Temperature of Refrigerant Gas Leaving HRU

T cg£ = Temperature of Suction Gas Entering Compressor

35
]Qa HP. T

0 = 20°F

300

250

200

150

100 -

50

2 3 4 TIME (fi)

1 1 1 1 1—1 300

- 250

P CD 200

04

<»a.

prg

Pr?C £ 150 - HPj,—HRU —V/H CO

CO

uu

a: - 100

Pcs

50

1 1 1 1 1 1

5 6 7

PCS

J 1 I I I L

1 2 3 4 5 S TIME (b)

10c HP

C T

q = S5°F

Figures II-IQa-c: Refp? i g erant. Pressures, versus, Time of Opeqat ro r i of the

_rIP - HR'J - V/H Integrated Appliance, Using HP.U 'lo. 1 Motes

P„

r = Compressor Discharge Pressure Preceding Gas Line to HRU

Prg

= Refrigerant Gas Pressure Following Gas Line from HRU

Pcs

= Compressor Suction Pressure

P RFC

= Refrigerant Pressure Following Condenser

36
11a H R 11b

h T

0 = 20°F

TIME (h)

HP, 75°F

lie HP

C T

0 = 95°F

Figures II-11a-c: Energy Consumed ver s us Time of Operation ce the

HP - HRU - WH fllTEGR A TED App L I A M CE, U 3 I f! G HRH No 1

37
TC = Thermocouple

HRU = Heat Recovery Unit

WH = Water Heater

HP = Heat Pump in Cooling Mode

c T

q

= Nominal Outdoor Temperature

TC 1

TC 2

TC 3

TC 4

TC 5

TC 6

TC 7

TC 8

TC 9

is 3-7/8

is 11-5/8

is 19-3/8

is 27-1/8

is 34-7/8

is 42-5/8

measures

measures

measures

" from the Bottom of the WH Tank

" from the Bottom of the WH Tank

" from the Bottom of the WH Tank

" from the Bottom of the WH Tank

" from the Bottom of the WH Tank

" from the Bottom of the WH Tank

the water temperature entering the HRU

the water Cut-Out Temperature leaving the HRU

the water Cut-In Temperature leaving the HRU

Addendum to Figure II-12A-C

38
12a HP

r: r

0

= 75°F

TIME [h]

12b HP

c Tq = 35°F

12c HP

C T

0 = 95°F

Figures 1I-12a-c: Water Temperatures at Various Locations in the WH versus

Tipe of Operation of ~he HP - HRU - WH Integrated

Appliance, Using hRIJ ?!q

.

2

39
T5a WP

c F - 75"p" HP

C T

0 - o5

5 F

!, Figures I1-13a-c: Meat Pfcoveey versus Time Operation of the

.HP - MRIJ - VH Integrvted Appliance, Us ins HRU No, 2» 40
Figures II-14a-c: Refrigerant Temperatures versus Time of Operation for the

HP - HRU - WH [ ute^rated Appliance, Using HP-U No. 2 Notes: Tcdg = Temperature of Compressor Discharge Gas

T rge

= Temperature of Refrigerant Gas Entering HRU

Trcl

= Temperature of Refrigerant Gas Leaving HRU

T soc

= Temperature or Suction Gas Entering Compressor

41
CO<>

TIME (hj

Figures II-15a-c: Refrigerant Pressures versus Time of Operation of the HP - IIRU - WH I utegratsd Appliaiice, Using HRU 'Jo. 2 Motes: P = Compressor Discharge Pressure Preceding Gas Line to HRU

PRG

= Refrigerant Gas Pressure Following Gas Line from HRU

Pcs

= Compressor Suction Pressure

Prfc

= Refrigerant Pressure Following Condenser

42
ENERGY

CONSUMED

(kWh)

26a UP

c

T0

- 75°F ]6d HP

C

T q

= 85°F

16c HP

C T

o = 95°f

Figures 1I-16a-c: Energy Consumed versus Time of Operation of the

HP - HRU - WH Integrated Appliance. Using HRU Ho. 2

43
COOLED

REFRIGERANT

GAS

LINE

44

Figure

III-ls

MODIFIED

AC/HP

-

HRU

-

WH

INTEGRATED

APPLIANCE

SYSTEM
a. Sample hourly water heater status using a 40 gallon tank

b. Sample hourly water heater status using a 40 gallon tank

c. Sample hourly water heater status using a 52 gallon tank

Figure III-2 a, b, and c : Heat Status vs. Time of Day, for Washington, D.C.

a and b: different days, b and c: same day

45
0.2

0

40

*

20

0

4

2

0

150

100

i

50

r

a. WATER HEATING

FACTOR

b. HOUSE LOAD AND

HEAT PUMP CAPACITY

/////

//SUPP.

// HEAT REQ

OUTDOOR DRY-BULB TEMPERATURE (°F)

Figure III-3: Heat Recovery Factors vs. Outdoor Temperature

46
a>u

3

UJ

hLU

OUJ

hCDO

fi3

I—

O

(M/nWJ!) QVOTHSnOH

03Vj

1)

47

Figure

III-4:

NBSLD

House

Heating

and

Cooling

Loads

vs.

Outdoor

Dry-Bulb

Temp
I

NEXT HOUR'S INPUT

COP = COEFFICIENT OF PERFORMANCE

CAP = CAPACITY (HEATING OR COOLING)

WHF = WATER HEATER FACTOR

-ft# HEAT PUMP CAPACITY,

HOUSE LOAD, OR

W.H. TANK STATUS

Figure III-5: Simulation Outline Diagram
Bibliography

1. Ambrose, E. R., Heat Pumps and Electric Heating; Wil<?y and Sons.

2. American Society of Heating, Refrigerating and Air Conditioning Engineers

(ASHRAE), ASHRAE Handbook and Product Directory, 1977 Fundamentals,

ASHRAE, 1977^ :

3. , ASHRAE Psychrometric Chart No. 1 , ASHRAE, 1963*

4. E. I. DuPont de Nemours & Company, Refrigerant - 22 Pressure-Enthalpy

Diagram ; duPont, 1964.

5. Kelly, G. , Potential of Air-to-Air Heat Pumps for Energy Conservation in

Residential Buildings , Proceedings of Heat Pumps and Their Contribution to

Energy Conservation, floor dhoff-Leyden, 1976.

6. U.S. Department(s) of Air Force, Army, and Navy, Engineering Weather Data,

AM 88-8, chapter 6, TM 5-785, NAVFAC P-89 , June 15, 19767

7. , National Bureau of Standards, Kelly, G. , Parken. W., Method of

Testing, Rating and Estimating the Seasonal Performance of Central

Air-Conditioners and Heat Pumps Operating in the Cooling Mode,

NBS1H 77-12i n:7"I97B"

5

8. > Kusuda, T., NBSID, The Computer Program for Heating and Cooling

Loads in Buildings, NBS Building Science Series 69 , July 1976.

9* , Energy, Argonne National Laboratory, Ashbury, J. and Others,

Assessment of Energy Storage Technologies and Systems, Phase II:

Heat Pump and Solar Energy Applications , ANT75PG-5 , 1978.

10. Division of Buildings and Community Systems, Concept Paper for

Air Conditioner/Heat Pump - Water Heater , 1979.

11. , Oak Ridge National Laboratory, Fischer, H- , Ice-Maker Heat

Pump-Part II, RSC February 1977.

12. , O'Neal, D. , Carney, J. , Hirst, E. , Regional Analysis of

Residential Water Heating Options: Energy Use and Economics,

frnltco^5I7T97b: :

13* Housing and Urban Development, Intermediate Minimum Property

Standards Supplement, 1977 Edition, Solar Heating and Domestic Hot

Water Systems, HUD 4930.2 ,~1977.

14. Weather King, Heat Pumps, Models QH, AH , Addison Products, 1978.

15* Ritchie, D. , Water Heating by Refrigeration, Weather King, Inc.,

March 1974. :

'

16. Wetherington, T., Heat Recovery Water Heating, Energy Conservation

Unlimited, (undated)

.

49
Appendix A. Simulation Background and Related Data

A-I. House Heating, Cooling and Hot Water Loads

Hourly data developed on another project at KBS were used as the source of

outdoor temperatures and loads on the heating and cooling system. Detailed

weather data and a suitable description of a house and its operation were

analyzed using portions of the NBSLD (National Bureau of Standards Load

Determination) program. Weather data for several years have been examined for

several locations in the United States, and for each location a reasonably

typical year, called a Test Reference Year (TRY) has been chosen for use in

weather dependent studies. The TRY's are 1957 for Washington, DC, 1975 for

Atlanta, GA, 1974 for Chicago, IL, and I960 for Seattle, WA. Dry- and wet-bulb

temperatures, wind speeds, cloud cover and solar radiation are among the

important environmental input data.

The same house was used in all locations for which the load data are

available and has the following features:

o Single family, one story detached dwelling

o 1800 square feet of floor over a crawl space (49' x 36' 9”)

o Insulation includes R-38 for the ceiling, R-13 for the walls, R-19

for the floor and double glazing for windows and glass doors.

o Infiltration was calculated separately for each hour, depending on

wind speed and the indoor/ outdoor temperature differential.

o Thermostat settings were 68°? for heating and 78°? for cooling.

o Indoor relative humidity was maintained between 20 percent and 60

percent.

o Natural ventilation was used only when it could keep the house

temperature at 68 to 78°F, while maintaining the indoor relative

humidity in the specified range. The amount of ventilation wa

s

variable and vents were closed if the humidity requirement could not

be met with ventilation or the ventilated indoor dry-bulb temperature

would be either above or below 78

U

F.

o Occupants were two adults and two children.

o 'The use of lights and equipment was given on an hourly schedule.

Data used in the present simulation include the hourly streams of outdoor

temperature and the heating and cooling loads on the environmental control

equipment. A daily cycle of hourly draws of hot water that was developed at NBS

was added to complete the description of the environment in which the heat

recovery system operates. This schedule is given in Table A-l, in which the

fractional draws were first translated info gallons on the basis of 64. 3 gallons

of hot-water use per day, aggregated by hour of the day and translated into kBtu

i cn
on the basis of a 90°F temperature rise and 8.25 pounds per gallon. Heat losses

from the WH were computed on the basis of one percent per hour of the thermal

energy in the water (i.e., the tank heat status, Q) at the beginning of each

hour. NBS tests indicate that this one percent factor is reasonable for well

insulated WH.

The input data from the NBS house load simulation studies consist of the

temperature for each hour and the corresponding sensible and latent heat

movement loads on the heat pump. Sensible loads are the amounts of thermal

energy that must be added or removed to control the indoor diy-bulb temperature.

Latent loads refer to humidity control. Both loads were computed using the

NBSLD program and are the resultants remaining after all other thermal energy

gains and losses have been considered. These loads were added for each hour to

yield the total load on the heat pump. Total loads for the year are the sums of

the absolute values of these hourly loads.

A-II. Heat Pump and Heat Recovery System

Three parameters were used to describe the essential characteristics of the

HP and HRU. These are the HP capacity, CAP, the HP coefficient of performance,

COP, and the water heating factor, WHF, for the HRU. WHF is the fraction of the

HP output that is available for heating water. CAP and COP depend on both the

indoor and outdoor temperatures and WHF depends on the temperature of the

refrigerant leaving the compressor. CAP, COP and WHF were evaluated at the midpoints of five degree temperature bins from 110°F down. to -40°F to cover the

main climate range in the U.S., taking into consideration the indoor

temperatures associated with heating and cooling situations. The reference HP

for these estimates is one of the two used in the experimental phase of this

study. It has a cooling capacity of 35,000 Btu per hour under standard ARE

onditions of 95°F outdoor dry-bulb temperature and indoor temperatures

dry-bulb and 67°F wet-bulb. Rated heating capacities for 7Q°F indoors

and outdoor dry- and wet-bulb combinations of 47°F/43 F and 17 F/15 F are 37,000

and 24,000 Btu per hour, respectively. The values for CAP, COP and WHF that

were used in the simulation and in the alternate calculation procedures are

given in Table A-2 and are compatible with the results of laboratory tests of a

sample HP-HHJ-WH system. The values for a given bin were used for all hours

with outdoor temperatures falling within the range of the bin. Since it would

be more economical to rely on supplemental resistance heat when the COP is less

than 1.0, CAP and COP are set to zero for these cold-weather conditions.

The COPs apply to steady state operation at the corresponding CAPs.

However, when the load, L, is less than the CAP, the HP will cycle on and off

and this part load COP will be less than the full load COP. (If L exceeds CAP,

the HP output is limited to CAP. ) This means that actual seasonal operating

costs will be somewhat higher than would be estimated using full load COP's

throughout the year. A part load COP adjustment factor developed at NBS was

used in the calculation of energy consumption for each hour or each bin,

depending on the computing model in use.

rating i of 80®F

COP (part load)

. 0.75 + 0.25 x -A- (l)

COP (steady state) CAP

51
A-III. Simulation Processing

The main loop in the hourly simulation was described in Section III, along

with two bin procedures. Figures A-l and A-2 are more detailed flow diagrams of

the computational logic of the hourly simulation and the alternate methods,

respectively. Details of the initializing print array generation and printout

are not shown. A page of symbol definitions accompanies each flow diagram. The

bin procedures do not consider the hourly variations of the heat status, Q, of

the water heater tank. Annual results are derived by computing results for one

hour under stable conditions at the selected location for each bin and then

weighting the results by the number of hours per year at the location.

Otherwise, the engineering concepts are identical for the hourly and bin

methods

.

Briefly, the steps in the hourly simulation are:

1. Specify the COP, CAP and WHF (described above) together with the

initial value of the tank heat status, Q, and the maximum and reserve

heat statuses, Q9 and Ql, respectively, for the tank being modeled.

2. Read the outdoor temperature and the house load for each successive

hour, carrying out the steps and making the accumulations described

in Section III.

A-IV. Additional Temperature Profiles

The outdoor dry-bulb temperature correlates quite well with the average

total sensible and latent house loads. Data from Bibliography Item 1 were used

to generate annual temperature profiles for several locations in the country.

The data were compiled in each case from hourly readings covering a period of

ten years or more. These profiles (Figures A-3—A-13) correspond in their

application to the climate profile at the bottom of Figure III-3.

52
Table A1 : Water Draw Schedule

Hour

Time Gallons Starting kBtu

7: 1: 0

7: 1:24

7: 5: 0

7:10: 0

7:25: 0

7:25:24

0.6

15

0.6 • 16.2

1

2

3

4

0

0

0

0

8: 0: 0

8: 0:24

0.6 0.6

5

6

7

8

9

10

11

12

0

0

12.029

0.446

0

9.356

0

0.891

10: 0: 0

10: 2: 0

10:22: 0

10:23: 0

10:30: 0

10:30:24

8

4

0.6 12,6

12: 0: 0

12: 0:24

12:45: 0

12:45:24

0.6

0.6 1,2

16: 0: 0

16: 0:24

0.6 0.6

13

14

15

0

n 17:20: 0

17:20:24 . 0.7 0.7

0

18: 0: 0

18: 0:24

18:20: 0

18:22: 0

18:28: 0

18:30: 0

0.6

3

3

16

17

18

19

0.446

0.520

11.583

0

18:36: 0

18:38: 0

18:44: 0

18:46: 0

18:52: 0

18:54: 0

3

3

3 15.6

20

21

22

23

0

0

12.029

0.446

22: 0: 0

22: 0:24

0.6 24 0

22: 5: 0

22:10: 0

' 15 Total 47.746

22:25: 0

22:25:24

0.6 16,2

23:30: 0

23:30:24 0.6 0.6

Total 64.3 64.3
Table A2: Heat Pump Capacities (CAP), Heat Pump

Coefficients of Performance (COP) and

Water Heating Factors (WHF)

BIN TDO^ CAP^ COP^ WHF^

C

1 107.5 30.15 2.05 0.33

2 102.5 32.25 2.17 0.28

3 97.5 34.15 2.29 0.24

4 92.5 35.65 2.40 0.20

5 87.5 36.75 2.51 0.16

6 82.5 37.^5 2.59 0.13

7 77.5 37.9 2.68 0.10

8 72.5 38.3 2.77 0.07

9 67.5 38.6 2.86 0.05

10 62.5 38.8 2.94 0.03

11 57.5 42.1 2.80 0.18

12 52.5 40.2 2.73 0.16

13 ^7-5 38.0 2.66 0.14

14 42.5 35.3 2.57 0.12

15 37.5 32.7 2.47 0.11

16 32.5 30.0 2.36 0.09

17 27.5 27.8 2.25 0.07

18 22.5 25.9 2.14 0.05

19 17.5 24.1 2.02 0.04

20 12.5 22.3 1.91 0.02

21 7-5 20.7 1.79 0 1 22 2.5 19.4 1 . 66 0

23 -2.5 18.1 1.54 0

24 -7.5 16.8 1.41 0

25 -12.5 15.5 1.28 0

26 -17.5 14.3 1.15 0

27 -22.5 13.1 1.01 0

28 -27.5 0 (d) 0 (d) 0

29 -32.5 0 0 0

30 -37.5 0 0 0

(a) Outdoor dry-bulb temperatures at the midpoints of the bins.

(b) _Full-load capacity, kBtu/h.

(c) Dimensionless.

(d) CAP and COP were set to zero for COP less than 1.0 because

the use of resistance heat alone would be more economical

than using the heat pump at all under’ these conditions.

WHF is zero when water at the required temperature cannot

be delivered to the tank.
Notes for Figure A-l

INT(xx) 3 Find the integer component of (xx).

TDO 3 Outdoor dry- bulb temperature for bin, deg. F.

HLS 3 Sensible house load, kBtu/h.

HLL 3 Latent house load, kBtu/h.

HLN 3 HLS + HLL 3 Net house load, kBtu/h.

WDR 3 Hot water draw for hour, kBtu.

BIN 3 Five-degree temperature bin containing TDO.

Q

3 Heat status of WH at start of hour, kBtu.

Q4 3 Heat status of WH after jacket losses and WDR for the hour, kBtu.

Q9 3 WH maximum heat status, kBtu.

Q1 = WH reserve heat status 3 Q9/4, kBtu.

BHLT 3 Cumulative backup heater load, kBtu.

(Accumulation formulas take the form; New Total = Old Total + Increment.)

Q6 3 Q9-Q4 3 Heat that the WH can accept, kBtu.

HPL 3 Heat pump (HP) load, kBtu/h.

HPK 3 HP electricity use, kWh/h.

HPKT 3 Cumulative HPK, kWh.

HOF 3 Heat offered to WH, kBtu/h.

HOFT = Cumulative HOF, kBtu.

WHF = Water heater factor, dimensionless.

WHR 3 WH heat refusal, kBtu/h.

WHRT 3 Cumulative WHR, kBtu.

SHL 3 Supplemental heater load, kBtu/h.

SHLT 3 Cumulative SHL, kBtu.

SHKT = Cumulative SHK, kWh.

BHKT = Cumulative BHK, kWh.
Figure A-l. Computational Flow Diagram for Hourly Simulation Method
Notes for Figure A-2

There are 30 five-degree temperature bins covering the range from 110°F to

-40°F. Data from curves are evaluated at the mid-point of each bin. There is

one value per bin for the following data items:

COP s HP coefficient of performance, dimensionless.

CAP = HP capacity, kBtu/h.

WHF = Water heater factor, dimensionless.

HRS = Site bin-hours per year.

HL1 s House load, bin average from detailed simulation input tape, kBtu/h.

HL2 s House load, from linear regressions of HL1 , kBtu/h.

HLX = HL1 for Bin Average method = HL2 for Bin Linear method, kBtu/h.

Bin variables derived in the course of the analysis include:

HPL - HP load, kBtu/h. (HPL is not allowed to exceed CAP).

HPK = HP kWh for the bin, corrected for part load effects on the HP, kWh/yr.

_ unc * HPL

y

4 x CAP-

- Mto x 3>412 x COp

x

3 x CAp + HpL

SHK = Supplemental heater load for bin, kWh/yr.

HOF = Heat offered to WH, kBtu/h.

WHO = Average hourly heat demand by WH, covering water draws and jacket losses.

For a 52 gallon tank,

_ ( 52-gal

x

q*| -] oss p er h our + 355 x 54 3 ga -| per day/8760 hrs per yr.)

x 8.25 lb per gal x 90°F rise/1000 = 2.086 kBtu/h.

WHA = Water heater acceptance, kBtu per hour.

WHK = Electricity demand on backup heater = HRS x WHA /3.412, kWh/yr.

Comparable bin results without heat recovery are also computed;

SHKO = Supplemental heater electrical load, kWh/yr.

HPLO = Heat pump load, kBtu/h.

HPKO = Heat pump electrical load, kWh/yr.

WHKQ = Backup heater electrical load, kWh/yr.
Start Read site and system data.

Initialize arrays and constants.

> f t Bin / . \

fov-—K

Main l0° p

>

VI/ \ Bins 1-30 /

Next Method

® 2

/

1 r

J Set HLX

No

HPL - lesser of

HOF = WHF x HPL

(HLX and CAP)

Method 1 = Bin Average; HLX = HL1

Method 2 = Bin Linear ; HLX = HL2

HOP = WHF x lesser of and CAP)

WHA = lesser of (HOF and WHD)

SHK = greater of

(0 and x (HLX + WHA " CA?)

}

HPL = lesser of ((HLX + WHA) and CAP)

A A

©*<£>

HPK = HRS x 4- x HPL x CAP

3.4-12 x COP x (3- x CAP + HPL)

HRS WHK = ^ x greater of

(6 and (WHO - WHF x HPL))

Compute kWh loads without heat recovery

(Suffix =0):

SHKO ='^77TT X greater of (0 and HLX - CAP)

3.4-12 = HPLO = lesser of (HLX and CAP)

HRS x 4 x HPLO x CAP =

3.4-12 x CO? x (3 x CAP t HPLO)

WHKO =—

y

Accumulate item totals

\ to printout

Figure A-2. Flow Diagram for 3in Average and Bin Linear Methods

58
CM

I

O

CM LO LTD

mk and shroh nib jo seuBUiH

59

Figure

A3:

Bin

Hours

per

Year

versus

Dry-Bulb

Bin

Temperature

for

Andrews

AFB
DRY-BULB

BIN

TEMPERATURE

(°F)
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NBS-114A (Rev. 9-70

U.S. DEPT. OF COMM.

BIBLIOGRAPHIC DATA

SHEET

1. PUBLICATION OR REPORT NO.

MBS'"- SO- 2 145

Z, Gov't. Accession No.}

zm>mb&*ph.

m tMWWwwfflt

4. TITLE AND SUBTITLE

A Study of an Integrated Appliance, the Air

Conditioner/Heat Pump-Heat Recovery Unit-Water Heater

5. Publication Date February 1981

S. Pefforswng Organization Code

r.;

'

7. AUTHOR(S) 8. Performing Organ. Report No. King-Mon Tu, Alan Davies, Sidney Fischler

9. PERFORMING ORGANIZATION NAME AND ADDRESS

NATIONAL BUREAU OF STANDARDS

DEPARTMENT OF COMMERCE

WASHINGTON, DC 20234

12. SPONSORING ORGANIZATION NAME AND COMPLETE ADDRESS (S treat. City, state, ZIP)

11. Contract/Grant No.

13. Type of Report & Period Covered United States Department of Energy

Oak Ridge National Laboratory

Oak Ridge, Tennessee 37830 14. Sponsoring Agency Code

1 '

: ' .v:; •

15. SUPPLEMENTARY NOTES

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

16. ABSTRACT (A 200~word or leas factual summary of most significant information . If document includes a significant bibliography or

literature survey, mention it here,)

Three integrated heat pump-heat recovery unit-water heater appliances were tested

under various environmental conditions to measure the functional parameters and

study the operating characteristics of these systems. It was found that the heat

recovery, heat recovery rate and heat recovery efficiency were dependent on the heat

recovery unit's characteristics. The use of the heat recovery unit in the system

resulted in a reduced work load for the heat pump's compressor and slightly improved

the heat pump's performance.

A computer simulation model of the integrated system was developed to study the

interactions between several of the pertinent system variables on an hourly basis

for selected situations and to estimate energy savings. Two alternative estimation

methods that utilize five-degree temperature bin data were also developed. The

estimated savings determined by using the alternative methods were about the same

as those estimated using the hourly data. Conclusions were also reached concerning

the use of water heaters with different tank capacities and on methods of increasing

potential energy savings.

17. KEY WORDS (six to twelve entries; alphabetical order; capitalize only the tirat letter of the first key word unless a proper name;

separated by semicolons)

Air conditioner retrofit; energy recovery system; heat recovery device; heat recovery

unit; heat pump retrofit; integrated energy recovery system; retrofit energy saver;

water heater energy saver

18. AVAILABILITY nUnlimited 19. SECURITY CLASS

(THIS REPORT)

21. NO. OF

PRINTED PAGES

X For Official Distribution. Do Not Release to NTIS

| ; Order From Sup. of Doc., U.S. Government Printing Office, Washington, DC

20402, SD Stock No. SN003-003-

UNCLASSIFIED

20. SECURITY CLASS

(THIS PAGE)

22. Price

| | Order From National Technical Information Service (NTIS), Springfield,

VA. 22161 UNCLASSIFIED

USCOMM-DC