U.S. patent number RE39,625 [Application Number 10/641,999] was granted by the patent office on 2007-05-15 for boosted air source heat pump.
This patent grant is currently assigned to Hallowell International, LLC. Invention is credited to David N. Shaw.
United States Patent |
RE39,625 |
Shaw |
May 15, 2007 |
Boosted air source heat pump
Abstract
A boosted source heat pump (BSHP) is presented having a primary
compressor, a booster compressor, an economizer, and a lubricant
management system. At least the primary compressor is a twin-single
compressor. The lubricant management system includes traps to
prevent migration of lubricant to a non-operating compressor, and
aspiration tubes from the sump to the low pressure intake zone to
each compressor.
Inventors: |
Shaw; David N. (East Falmouth,
MA) |
Assignee: |
Hallowell International, LLC
(Bangor, ME)
|
Family
ID: |
25682524 |
Appl.
No.: |
10/641,999 |
Filed: |
August 15, 2003 |
Related U.S. Patent Documents
|
|
|
|
|
|
|
Application
Number |
Filing Date |
Patent Number |
Issue Date |
|
Reissue of: |
09505584 |
Feb 16, 2000 |
06276148 |
Aug 21, 2001 |
|
|
Current U.S.
Class: |
62/117; 237/2B;
62/160; 62/196.1; 62/228.5 |
Current CPC
Class: |
F04B
9/04 (20130101); F04B 39/0094 (20130101); F25B
1/10 (20130101); F25B 13/00 (20130101); F25B
31/002 (20130101); F25B 49/02 (20130101); F25B
2400/13 (20130101) |
Current International
Class: |
F25B
13/00 (20060101) |
Field of
Search: |
;62/160,117,196.1,228.2,229,510 ;236/1EA ;237/2B |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
|
|
|
|
|
|
|
0 106 414 |
|
Apr 1984 |
|
EP |
|
0 715 077 |
|
Jun 1996 |
|
EP |
|
0 718 568 |
|
Jun 1996 |
|
EP |
|
06213170 |
|
Feb 1994 |
|
JP |
|
WO97/32168 |
|
Sep 1997 |
|
WO |
|
Other References
Theory of Mechanical Refrigeration, N.R. Sparks, McGraw Hill, 1938
(pp. (111-127). cited by other .
J.L. Threlkeld, Thermal Environmental Engineering, Prentice-Hall,
Inc. copyright, 1970, 1962 (pp. 63-69). cited by other .
Survey And Comparison of Interstage Cooling Systems For Two Stage
Compression, Data Sheet, No. 20, May 1979. cited by other .
E.G. Pita, Refrigeration Principles and Systems, Business News
Publishing Company, 1991 (pp. 243-245). cited by other .
B.D. Wood, Applications of Thermodynamics, Waveland Press, Inc.,
copyright 1982 (pp. 218-222). cited by other .
S.M. Elenka and Q.W. Minich, Standard Refrigeration and Air
Conditioning Questions & Answers, McGraw Hill, copyright
1983,1973,1961 (pp. 28-31, 50-53). cited by other .
I. Cerepnalkowski, Modern Refrigerating Machines, Elsevier Science
Publishers, B.V., copyright 1991 (pp. 47,48). cited by other .
A Technical Handbook From SWEP, 1993 (1 page plus cover and back
sheets). cited by other .
"New Technology and Products" Jul. 1999
(www.bristolcompressors.com). cited by examiner.
|
Primary Examiner: Tapolcai; William E.
Attorney, Agent or Firm: Cantor Colburn LLP
Claims
What is claimed is:
1. A compression module for a heating or air conditioning system,
said module including: a primary compressor; a booster compressor;
at least said primary compressor being an unloadable positive
displacement compressor; a sensor for sensing the temperature of
outdoor ambient air; and a controller, said controller being
responsive to signals from said sensor commensurate with the
temperature of outdoor ambient air to operate said primary
compressor and said booster compressor in a predetermined operating
sequence.
2. A compressor mode as in claim 1, wherein said predetermined
operating sequence for heating includes the following: (a) allow
partial capacity operation of said primary compressor when the
outdoor ambient temperature is in the range of about 50.degree.
F.-75.degree. F.; (b) allow full capacity operation of said primary
compressor when the outdoor ambient temperature is in the range of
about 43.degree. F.-50.degree. F.; (c) allow full capacity
operation of said booster compressor and partial capacity operation
of said primary compressor when the outdoor ambient temperature is
in the range of about 33.degree. F.-43.degree. F.; (d) allow full
capacity operation of said booster compressor, and full capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of about 15.degree. F.-33.degree.
F.
3. A compressor module as in claim 1, wherein: said primary
compressor is a multiple-cylinder compressor.
4. A compressor module as in claim 1, wherein: said booster
compressor is a single speed compressor or a two speed
compressor.
5. A compressor module as in claim 1, wherein: the displacement of
said booster compressor is about 30% to about 70% greater than the
displacement of said primary compressor.
6. A compressor module as in claim 1, wherein: the coefficient of
performance (C.O.P.) of the system for heating is at least 1.5 for
the conditions of 0.degree. F. outdoor ambient temperature and
70.degree. F. indoor heated space temperature.
7. A compression module as in claim 1, including: an inlet to said
primary compressor for the supply of working fluid to said primary
compressor; an inlet to said booster compressor for the supply of
working fluid to said booster compressor; a supply line from the
discharge from said booster compressor to the inlet of said primary
compressor; and a lubricant management system, said system having:
(a) an aspiration tube extending from the lubricant sump of said
primary compressor to the intake of said primary compressor, (b) an
aspiration tube extending from the lubricant sump of said booster
compressor to the intake of said booster compressor, (c) a liquid
trap in the supply line to the inlet to said booster compressor,
and (d) a liquid trap in the supply line between the discharge from
said booster compressor to the intake to said primary
compressor.
8. A method of operating in sequence a heating system having a
primary compressor and a booster compressor, the method including
the steps of: (a) sensing the outdoor ambient temperature; (b)
allowing partial capacity operation of said primary compressor when
the outdoor ambient temperature is in the range of about 50.degree.
F.-75.degree. F.; (c) allowing full capacity operation of said
primary compressor when the outdoor ambient temperature is in the
range of about 43.degree. F.-50.degree. F.; (d) allowing full
capacity operation of said booster compressor and partial capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of about 33.degree. F.-43.degree. F.;
(e) allowing full capacity operation of said booster compressor,
and full capacity operation of said primary compressor when the
outdoor temperature is in the range of about 15.degree.
F.-33.degree. F.
9. The method of claim 8, including the step of: allowing operation
of back-up resistance heating when the outdoor ambient temperature
is about 15.degree. F. or lower.
10. A compressor module as in claim 1, wherein said predetermined
operating sequence for cooling includes the following: (a) allow
partial capacity operation of said primary compressor when the
outdoor ambient temperature is in the range of about 60.degree.
F.-85.degree. F.; (b) allow full capacity operation of said primary
compressor when the outdoor ambient temperature is in the range of
above about 85.degree. F.; (c) allow full capacity operation of
said primary compressor and full capacity operation of said booster
compressor when the outdoor ambient temperature is in the range of
about 105.degree. F. or higher.
11. A compression module for a heat pump system, said module
including: a primary compressor; a booster compressor; at least
said primary compressor being a multi-cylinder unloadable
compressor; each of said compressors having an inlet and a supply
line connected to said inlet; a sensor for sensing the temperature
of outdoor ambient air; a thermostat for sensing the temperature of
a volume of air to be .[.healed.]. .Iadd.heated.Iaddend.; and a
controller, said controller being responsive to signals from said
sensor commensurate with the temperature of outdoor ambient air,
and to signals from said thermostat commensurate with the
temperature of the air to be heated to operate said primary
compressor.[.,.]. .Iadd.and .Iaddend.said booster compressor .[.and
said economizer.]. in a predetermined operating sequence.
12. A compressor module as in claim 11, wherein said predetermined
operating sequence for heating includes the following: (f) allow
partial capacity operation of said primary compressor when the
outdoor ambient temperature is in the range of about
50.degree.-75.degree. F.; (g) allow full capacity operation of said
primary compressor when the outdoor ambient temperature is in the
range of about 43.degree. F.-50.degree. F.; (h) allow full capacity
operation of said booster compressor and partial capacity operation
of said primary compressor when the outdoor ambient temperature is
in the range of about 33.degree. F.-43.degree. F.; (i) allow full
capacity operation of said booster compressor, and full capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of about 15.degree. F.-33.degree.
F.
13. A compressor module as in claim 11, wherein: the ratio of the
displacement of said booster compressor to the displacement of said
primary compressor is at least about 1.3 to 1.
14. A compressor module as in claim 11, wherein: the heating
coefficient of performance (C.O.P.) of the heat pump system is at
least about 1.5 for the conditions of 0.degree. F. outdoor ambient
temperature and 70.degree. F. indoor heated space temperature.
15. A compression module as in claim 11, wherein: 100% of the
capacity of said primary compressor is greater than the ordinary
cooling requirements of the heat pump system.
16. A compression module for a heat pump system, said module
including: a primary compressor; a booster compressor; at least
said primary compressor being a multi-cylinder unloadable
compressor; a sensor for sensing the temperature of outdoor ambient
air; a thermostat for sensing the temperature of a volume of air to
be heated; and a controller, said controller being responsive to
signals from said sensor commensurate with the temperature of
outdoor ambient air, and to signals from said thermostat
commensurate with the temperature of the air to be heated to
operate said primary compressor and said booster compressor in a
predetermined heating operating sequence including the following:
(a) allow partial capacity operation of said multi-cylinder primary
compressor when the outdoor ambient temperature is in the range of
about 50.degree.-70.degree. F.; (b) allow full capacity operation
of said multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of about 43.degree. F.-50.degree. F.;
(c) allow full capacity operation of said booster compressor and
partial capacity operation of said multi-cylinder primary
compressor when the outdoor ambient temperature is in the range of
about 33.degree. F.-43.degree. F. (d) allow full capacity operation
of said booster compressor, and full capacity operation of said
multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of about 15.degree. F.-30.degree. F.;
(e) allow operation of back-up resistance heating when the outdoor
ambient temperature is about 15.degree. F. or lower.
17. A compressor module as in claim 16, wherein: the ratio of the
displacement of said booster compressor to the displacement of said
primary compressor is at least about 1.3 to 1.
18. A compressor module as in claim 16, wherein: the heating
coefficient of performance (C.O.P.) Of the heat pump system is at
least about 1.5 for the conditions of 0.degree. F. outdoor ambient
temperature and 70.degree. F. indoor heated space temperature.
19. A compression module as in claim 16, wherein: 100% of the
capacity of said primary compressor is greater than the ordinary
cooling requirements of the heat pump system.
20. A compressor module as in claim 1, wherein: said booster
compressor is a single speed compressor.
21. A compressor module as in claim 16, wherein: said controller is
responsive to signals from said sensor commensurate with the
temperature of outdoor ambient air, and to signals from said
thermostat commensurate with the temperature of the air to be
cooled to operate said primary compressor, and said booster
compressor in a predetermined cooling operating sequence, including
the following: (a) allow partial capacity operation of said primary
compressor when the outdoor ambient temperature is in the range of
about 60.degree. F.-85.degree. F.; (b) allow full capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of above about 85.degree. F.; (c) allow
full capacity operation of said primary compressor and full
capacity operation of said booster compressor when the outdoor
ambient temperature is in the range of about 105.degree. F. or
higher.
22. A method of operating in heating sequence a heat pump system
having a multi-cylinder unloadable primary compressor and a booster
compressor, the method including the steps of: (a) sensing the
outdoor ambient temperature; (b) allowing partial capacity
operation of said multi-cylinder primary compressor when the
outdoor ambient temperature is in the range of about
50.degree.-75.degree. F.; (c) allowing full capacity operation of
said multi-capacity primary compressor when the outdoor ambient
temperature is in the range of about 43.degree. F.-50.degree. F.;
(d) allowing full capacity operation of said booster compressor and
partial capacity operation of said multi-cylinder primary
compressor when the outdoor ambient temperature is in the range of
about 33.degree. F.-43.degree. F.; (e) allowing full capacity
operation of said booster compressor, and full capacity operation
of said multi-cylinder primary compressor, when the outdoor ambient
temperature is in the range of about 15.degree. F.-33.degree. F.;
(f) allowing operation of back-up resistance heating when the
outdoor ambient temperature is about 15.degree. F. or lower.
23. A method of operating in cooling sequence a heat pump system
having a multi-cylinder unloadable primary compressor, and a
booster compressor, that method including the steps of: (a) sensing
the outdoor ambient temperature; (b) allowing partial capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of about 60.degree. F.-85.degree. F.;
(c) allowing full capacity operation of said primary compressor
when the outdoor ambient temperature is in the range of above about
85.degree. F.; (d) allowing full capacity operation of said primary
compressor and full capacity operation of said booster compressor
when the outdoor ambient temperature is in the range of about
105.degree. F. or higher.
24. A compression module for a heat pump system, said module
including: a primary compressor; a booster compressor; at least
said primary compressor being a multi-cylinder unloadable
compressor; each of said compressors having an inlet and a supply
line connected to said inlet; a supply line from the discharge from
said booster compressor to the inlet of said primary compressor; an
economizer; a sensor for sensing the temperature of outdoor ambient
air; a thermostat for sensing the temperature of a volume of air to
be heated; and a controller, said controller being responsive to
signals from said sensor commensurate with the temperature of
outdoor ambient air, and to signals from said thermostat
commensurate with the temperature of the air to be heated to
operate said primary compressor, said booster compressor and said
economizer in the following heating sequence: (a) allowing partial
capacity operation of said multi-cylinder primary compressor when
the outdoor ambient temperature is in the range of about
50.degree.-75.degree. F.; (b) allow full capacity operation of said
multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of about 43.degree.F.-50.degree. F.;
(c) allow full capacity operation of said booster compressor and
partial capacity operation of said multi-cylinder primary
compressor when the outdoor ambient temperature is in the range of
about 33.degree. F.-43.degree. F.; (d) allow full capacity
operation of said booster compressor, and partial capacity
operation of said multi-cylinder primary compressor, and said
economizer when the outdoor ambient temperature is in the range of
about 25.degree. F.-33.degree. F.; (e) allow full capacity
operation of said booster compressor, and full capacity operation
of said multi-cylinder primary compressor, and said economizer when
the outdoor ambient temperature is in the range of about 15.degree.
F.-25.degree. F.; (f) allow operation of back-up resistance heating
when the outdoor ambient temperature is about 15.degree. F. or
lower.
25. A compression module for a heat pump system, said module
including: a primary compressor; a booster compressor; at least
said primary compressor being a multi-cylinder unloadable
compressor; each of said compressors having an inlet and a supply
line connected to said inlet; a supply line from the discharge from
said booster compressor to the inlet to said primary compressor; an
economizer; a sensor for sensing the temperature of outdoor ambient
air; a thermostat for sensing the temperature of a volume of air to
be cooled; and a controller, said controller being responsive to
signals from said sensor commensurate with the temperature of
outdoor ambient air, and to signals from said thermostat
commensurate with the temperature of the air to be heated to
operate said primary compressor, said booster compressor and said
economizer in the following cooling sequence: (a) allow partial
capacity operation of said multi-cylinder primary compressor when
the outdoor ambient temperature is in the range of about 60.degree.
F.-85.degree. F.; (b) allow full capacity operation of said
multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of above about 85.degree. F.; (c) allow
full capacity operation of said multi-cylinder primary compressors
and full capacity operation of said booster compressor when outdoor
ambient temperature is in the range of about 105.degree. F.
26. The compression module of claim 25, including: allow operation
of said economizer and full capacity operation of said primary
compressor and full capacity operation of said booster
compressor.
27. A compressor module as in claim 24, including: a lubricant
management system, said system having: (a) an aspiration tube
extending from the lubricant sump of said primary compressor to the
intake of said primary compressor, (b) an aspiration tube extending
from the lubricant sump of said booster compressor to the intake of
said booster compressor, (c) a liquid trap in the supply line to
the inlet to said booster compressor, and (d) a liquid trap in the
supply line between the discharge from said booster compressor to
the intake to said primary compressor.
28. A compressor module as in claim 24, wherein: the displacement
of said booster compressor is about 10%-about 50% greater than the
displacement of said primary compressor.
29. A compressor module as in claim 24, wherein: the ratio of the
displacement of said booster compressor to the displacement of said
primary compressor is at least about 1.3 to 1.
30. A compressor module as in claim 24, wherein: the coefficient of
performance (C.O.P.) Of the heat pump system is at least about 2
for the conditions of 0.degree. F. outdoor ambient temperature and
70.degree. F. indoor heated space temperature.
31. A compression module as in claim 24, wherein: 100% of the
capacity of said primary compressor is greater than the ordinary
cooling requirements of the heat pump system.
32. A compressor module as in claim 24, wherein: said booster
compressor is a single speed compressor.
33. A compressor module as in claim 24, wherein: each of said
booster compressor and said primary compressor is a twin-single
compressor.
34. A method of operating in heating sequence a heat pump system
having a primary multi-cylinder unloadable compressor, a booster
compressor and an economizer, the method including the steps of:
(a) sensing the outdoor ambient temperature; (b) allowing partial
capacity operation of said multi-cylinder primary compressor when
the outdoor ambient temperature is in the range of about
50.degree.-75.degree. F.; (c) allowing full capacity operation of
aid multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of about 43.degree. F.-50.degree. F.;
(d) allowing full capacity operation of said booster compressor and
partial capacity operation of said multi-cylinder primary
compressor when the outdoor ambient temperature is in the range of
about 33.degree. F.-43.degree. F.; (e) allowing full capacity
operation of said booster compressor, and partial capacity
operation of said multi-cylinder primary compressor, and said
economizer when the outdoor ambient temperature is in the range of
about 25.degree. F.-33.degree. F.; (f) allowing full capacity
operation of said booster compressor, and full capacity operation
of said multi-cylinder compressor, and said economizer when the
outdoor ambient temperature is in the range of about 15.degree.
F.-25.degree. F.; (g) allowing operation of back-up resistance
heating when the outdoor ambient temperature is about 15.degree. F.
or lower.
35. A method of operating in cooling sequence a heat pump system
having a multi-cylinder unloadable primary compressor, a booster
compressor and an economizer, the method including the steps of:
(a) sensing the outdoor ambient temperature; (b) allowing partial
capacity operation of said multi-cylinder primary compressor when
the outdoor ambient temperature is in the range of about 60.degree.
F.-85.degree. F.; (c) allowing full capacity operation of said
multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of above about 85.degree. F.; and (d)
allowing full capacity operation of said multi-cylinder primary
compressor and full capacity operation of said booster compressor
when the outdoor ambient temperature is in the range of about
105.degree. F.
36. The method as in claim 35, including the step of: operating
said economizer along with full capacity operation of said
multi-cylinder primary compressor and full capacity operation of
said booster compressor.
.Iadd.37. A compressor system for a heating or air conditioning
system including: a primary compressor; a booster compressor; at
least said primary compressor being a multi-capacity compressor; a
sensor for sensing the temperature of outdoor ambient air or a
parameter commensurate with the temperature of outdoor ambient air;
and a controller, said controller being responsive to signals from
said sensor commensurate with the temperature of outdoor ambient
air to operate said primary compressor and said booster compressor
in a predetermined operating sequence. .Iaddend.
.Iadd.38. A compressor system as in claim 37, wherein said
predetermined operating sequence for heating includes the
following: (a) allow partial capacity operation of said primary
compressor when the outdoor ambient temperature is in the range of
about 50.degree. F.-75.degree. F.; (b) allow full capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of about 43.degree. F.-50.degree. F.;
(c) allow full capacity operation of said booster compressor and
partial capacity operation of said primary compressor when the
outdoor ambient temperature in the range of about 33.degree.
F.-43.degree. F.; (d) allow full capacity operation of said booster
compressor and full capacity of said primary compressor when the
outdoor ambient temperature is in the range of about 15.degree.
F.-33.degree. F. .Iaddend.
.Iadd.39. A compressor system as in claim 37 wherein: said primary
compressor is a multiple-cylinder compressor. .Iaddend.
.Iadd.40. A compressor system as in claim 37 wherein said booster
compressor is a single speed compressor or a two-speed compressor.
.Iaddend.
.Iadd.41. A compressor system as in claim 37, wherein: the
displacement of said booster compressor is about 30% to about 70%
greater than the displacement of said primary compressor.
.Iaddend.
.Iadd.42. A compressor system as in claim 37, wherein: the
coefficient of performance (C.O.P.) of the system for heating is at
least 1.5 for the conditions of 0.degree. F. outdoor ambient
temperature and 70.degree. F. indoor heated space temperature.
.Iaddend.
.Iadd.43. A compressor system as in claim 37 including: an inlet to
said primary compressor for the supply of working fluid to said
primary compressor; a supply line from the discharge from said
booster compressor to the inlet to said primary compressor; and a
lubricant management system, said system having: (a) an aspiration
tube extending from the lubricant sump of said primary compressor
to the intake of said primary compressor; (b) an aspiration tube
extending from the lubricant sump of said booster compressor to the
intake of said booster compressor; (c) a liquid trap in the supply
line between the discharge from said booster compressor to the
intake to said primary compressor. .Iaddend.
.Iadd.44. A compressor system as in claim 37 wherein said
predetermined operating sequence for cooling includes the
following: (a) allow partial capacity operation of said primary
compressor when the outdoor ambient temperature is in the range of
about 60.degree. F.-85.degree. F.; (b) allow full capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of above about 85.degree. F.; (c) allow
full capacity operation of said primary compressor and full
capacity operation of said booster compressor when the outdoor
ambient temperature is in the range of about 105.degree. F. or
higher. .Iaddend.
.Iadd.45. A heat pump system, said system including: a primary
compressor; a booster compressor; at least said primary compressor
being a multi-capacity compressor; each of said compressors having
an inlet and a supply line connected to said inlet; a sensor for
sensing the temperature of outdoor ambient air or a parameter
commensurate with the temperature of outdoor ambient air; a
thermostat for sensing the temperature of a volume of air to be
heated; and a controller, said controller being responsive to
signals from said sensor commensurate with the temperature of
outdoor ambient air, and to signals from said thermostat
commensurate with the temperature of the air to be heated to
operate said primary compressor, said booster compressor in a
predetermined operating sequence. .Iaddend.
.Iadd.46. A compressor system as in claim 45 wherein: the ratio of
the displacement of said booster compressor to the displacement of
said primary compressor is at least about 1.3 to 1. .Iaddend.
.Iadd.47. A compressor system as in claim 45 wherein: the heating
coefficient of performance (C.O.P.) of the heat pump system is at
least about 1.5 for the conditions of 0.degree. F. outdoor ambient
temperature and 70.degree. F. indoor heated space temperature.
.Iaddend.
.Iadd.48. A compressor system as in claim 45 wherein: 100% of the
capacity of said primary compressor is greater than the ordinary
cooling requirement of the heat pump system. .Iaddend.
.Iadd.49. A compressor system as in claim 37, wherein: said booster
compressor is a single speed compressor. .Iaddend.
.Iadd.50. A method of operating in cooling sequence a heat pump
system having a multi-capacity primary compressor, and a booster
compressor, that method including the steps of: (a) sensing the
temperature of outdoor ambient air or a parameter commensurate with
the outdoor ambient temperature; (b) allowing partial capacity
operation of said primary compressor when the outdoor ambient
temperature is in the range of about 60.degree. F.-85.degree. F.;
(c) allowing full capacity operation of said primary compressor
when the outdoor ambient air temperature is in the range of above
about 85.degree. F.; (d) allowing full capacity operation of said
primary compressor and full capacity operation of said booster
compressor when the outdoor ambient air temperature is in the range
of about 105.degree. F. or higher. .Iaddend.
.Iadd.51. A compressor system for a heat pump system, said system
including: a primary compressor; a booster compressor; at least
said primary compressor being a multi-capacity compressor; each of
said compressors having an inlet and a supply line connected to
said inlet; a supply line from the discharge from said booster
compressor to the inlet to said primary compressor; a sensor for
sensing the temperature of outdoor ambient air or a parameter
commensurate with the temperature of outdoor ambient air; and a
thermostat for sensing the temperature of a volume of air to be
cooled; and a controller, said controller being responsive to
signals from said sensor commensurate with the temperature of
outdoor ambient air, and to signals from said thermostat
commensurate with the temperature of the air to be heated to
operate said primary compressor, and said booster compressor in the
following cooling sequence; (a) allow partial capacity operation of
said multi-capacity primary compressor when the outdoor ambient
temperature is in the range of about 60.degree. F.-85.degree. F.;
(b) allow full capacity operation of said multi-capacity primary
compressor when the outdoor ambient temperature is in the range of
above about 85.degree. F.; (c) allow full capacity operation of
said multi-capacity primary compressors and full capacity operation
of said booster compressor when outdoor ambient temperature is in
the range of about 105.degree. F. .Iaddend.
.Iadd.52. The compressor system of claim 51 including: allow
operation of said economizer and full capacity operation of said
primary compressor and full capacity operation of said booster
compressor. .Iaddend.
.Iadd.53. A method of operating in cooling sequence a heat pump
system having a multi-capacity primary compressor, a booster
compressor and an economizer, the method including the steps of:
(a) sensing the temperature of outdoor ambient air or a parameter
commensurate with the outdoor ambient temperature; (b) allowing
partial capacity operation of said primary compressor when the
outdoor ambient temperature is in the range of about 60.degree.
F.-85.degree. F.; (c) allowing fully capacity operation of said
multi-cylinder primary compressor when the outdoor ambient
temperature is in the range of above about 85.degree. F.; and (d)
allowing full capacity operation of said multi-cylinder primary
compressor and fully capacity operation of said booster compressor
when the outdoor ambient temperatures is in the range of
105.degree. F. .Iaddend.
.Iadd.54. The method as in claim 53 including the step of:
operating an economizer along with full capacity operation of said
multi-capacity primary compressor and full capacity operation of
said booster compressor. .Iaddend.
Description
BACKGROUND AND CROSS REFERENCE
This invention relates to air-source heat pumps. More particularly,
it relates to new and improved air source heat pumps especially
suitable for use in normally colder climates. This invention is an
improvement on my U.S. Pat. No. 5,927,088, the entire contents of
which are incorporated herein by reference.
The following discussion of typical prior art heat pumps refers to
air-source heat pumps other than the heat pumps disclosed in my
U.S. Pat. No. 5,927,088.
The air-source heat pump system is the most prevalent type of heat
pump used in the world today. This is the case whether one is
discussing room units, residential central type, ductless splits,
or rooftop commercial systems.
Although the air-source concept in general has a high application
potential worldwide, its popularity in the United States and
elsewhere has been greatest in mild climate areas. This is because
the compressor-derived heating capacity of typical prior art units
declines rapidly as the outdoor ambient falls, due, in most part,
to the large increase in specific volume (i.e., decrease in
density) of the outdoor coil generated refrigerant vapor as the
ambient (outdoor) temperature falls. This fall in
compressor-derived heating capacity is obviously opposite to the
heating requirement, which increases as the outdoor ambient
temperature falls. When a typical prior art heat pump operates
below its balance point (about 35.degree. F.-40.degree. F.),
supplemental heating is required. The most prevalent form of
supplemental heat used is electric resistance. In other than mild
climates, this use of supplemental electric resistance heat puts
the air-source heat pump at a serious economic disadvantage to a
consumer as compared with other forms of heating (such as natural
gas, oil, and propane), because of the high cost of electric
resistance heating.
Electric utilities are also very concerned because of the
associated large transformers and distribution systems that are
required for any large populations of typical prior art heat pumps
whenever high electric resistant (KW) heat backup is required on a
regular basis for large population.
When operating at low outdoor ambient temperatures such as
0.degree. F., homes heated by typical vapor art heat pumps require
as much as KW input from the utility as does a home heated by
electric resistance KW alone. This is not acceptable to the utility
as they would have to increase their generating capacity to supply
the demand. In other words, a Northern utility that was summer
peaking would now become winter peaking because much more KW output
is required for electrically heated Northern homes than what is
required for cooling those same homes.
As discussed in my U.S. Pat. No. 5,927,088, one of the areas for
capacity and efficiency improvement of air source heat pump systems
lies in the recovery of significant heat energy currently remaining
in the condensed liquid refrigerant leaving the system condenser.
If this remaining energy is recovered and returned directly to the
heating side of the system before being thermally degraded and sent
to the system evaporator as low density vapor (as is now the case
in present day systems), significant increases in compressor
derived heating capacity and C.O.P. can be made at lower outdoor
ambient temperatures.
The basic problem here is that after the refrigerant has been fully
liquefied in the heating condenser, there is still a large amount
of energy left in the leaving warm liquid. This remaining energy
evaporates a large portion of the heating liquid itself during the
normal pressure reduction process that is required to develop the
necessarily low evaporating temperatures. Depending on the
refrigerant utilized, and the degree of temperature existing
between the evaporator and the condenser, as much as one-half of
this liquid can be evaporated during this normal pressure reduction
process across the system expansion device when operating at the
lower outdoor ambient temperatures.
Obviously, if this liquid has already evaporated, it cannot be
again evaporated in the system evaporator, and thus cannot absorb
energy from the outside air. However, the net resulting vapor must
pass through the system evaporator anyway, creating additional
pressure drop along its way, and then must be induced and fully
compressed to the condensing level by the compressor, thus
requiring the necessary power to accomplish the compression. The
C.O.P. of this portion of the heating process is only one (1)
because no energy has been absorbed from the outside air.
Since the compressor must induct this previously evaporated vapor,
the compressor can only induct a correspondingly smaller amount of
vapor that has been derived from the cooling of outside air (by
evaporating the refrigerant liquid that does enter the evaporator
along with the previously mentioned vapor). This is not a
reasonable process for typical prior art air-source heat pumps
operating in other than the milder ambient temperatures because
only under those conditions is the relative amount of liquid to
vapor (by weight) sufficient to result in a good system C.O.P.
The heating energy output of any heat pump system is also closely
proportional to the weight flow of refrigerant vapor entering the
system condenser. Approximately 4 times the amount of heat energy
is required of 0.degree. F. than is required at 50.degree. F. This
means that approximately a 400% increase of entering condenser
refrigerant vapor is required at 0.degree. F. ambient as compared
to 50.degree. F. ambient in order to adequately match the heating
energy requirement. However, the density of the refrigerant vapor
generated in the system evaporator when operating at 0.degree. F.
ambient is only about 32% of that generated when the outdoor
temperature is 50.degree. F. Therefore, when approximately four (4)
times the weight flow is required when only 32% of the vapor
density is generated, it becomes very obvious that significant
changes must be made in order to make an air source pump viable for
colder Northern climates.
In addition, if the entire space heating requirement at 0.degree.
F. outdoor ambient is to be supplied by compressor derived heating
capacity, the air flow across the heating coil of the condenser
must be such that the indoor delivered air temperature will be
around 105.degree. F. in order to provide adequate freedom from a
sensation of cool drafts. This in turn will cause the system
condensing temperature to rise to around 115.degree. F. considering
a reasonably sized indoor coil surface.
The end result of all this (even if the necessary compressor
displacement could be obtained somehow with a present day single
compression stage system) is to cause overall system operating
compression ratios to rise to the point where it becomes
unrealistic to continue the use of typical prior art heat pump
technology in normally colder climates. This is exactly what has
happened in the marketplace of today. Air source heat pumps are no
longer purchased for use in cold climate areas for reasons of both
poor comfort as well as the very high cost of the electric energy
requirement.
SUMMARY OF THE INVENTION
In this invention, a system and method are presented that achieve a
great increase in refrigerant pumping capacity as related to large
fall in the outdoor ambient temperature, combined with a method of
extracting energy from warm liquid leaving the system condenser
where this is needed as well. The heat pump system itself is made
to be sufficiently flexing in order to accomplish the necessary
goal of both sufficient and efficient heating for a wide range of
outdoor ambient temperatures.
In order for anti-source heat pumps to become serious contenders
for use in colder climates, significant changes must be made for
them to realize their true potential. Fundamental Carnot Theory
thermodynamic principles unquestionably show that electric powered
air source heat pumps indeed to have significant potential in cold
climates. In fact, the theoretical Carnot C.O.P. (coefficient of
performance) limit for a sink (room) temperature of 70.degree. F.
and a source (outside air) temperature of 0.degree. F. is 7.57
units of energy delivered to the sink for every 1 net unit of
energy supplied to the compression process.
Carnot C.O.P.=(T.sub.2.DELTA.S)+(T.sub.2-T.sub.1).DELTA.S where
T.sub.2 is the delivered energy sink temperature (room temperature
in absolute degrees) and T.sub.1 is the supplied energy source
temperature (outside air temperature in absolute degrees) and
.DELTA.S is the constant change in entropy for this theoretical
cycle. Therefore, executing the equation, the .DELTA.S's cancel out
and the final equation is simply Carnot
C.O.P.=T.sub.2+(T.sub.2-T.sub.1) where only T.sub.2 need be in
absolute degrees.
It is expected that with this present invention, an actual heating
C.O.P. of at least 2.0 will be reached at this 0.degree. F. outdoor
ambient condition. This represents a Carnot efficiency level of
only [(2.0+7.5).times.100]=26%, which clearly is within the bound
of rational achievability. Accordingly, the system and method of
this invention allows the basic electric utility service size to be
only about 50% of that which is required for today's heat pumps
selected for cold climates.
In typical prior art systems, the actual delivered C.O.P. is only
about 1.0 at the condition of 70.degree. F. room temperature and
0.degree. F. outside air temperature, because most of their
delivered energy comes from electric resistance coils which
obviously operate with a C.O.P. of 1.0 (1 unit of delivered energy
for every 1 unit of supplied energy). Also, with these typical
prior art systems, the delivered energy that does come from the
refrigeration circuit may even come with a C.O.P. of less than 1 at
these low outdoor ambient temperatures because of a significant
percentage energy loss between the outdoor unit and the indoor
condenser when operating under low refrigerant flow rate condition.
Further to this problem, most residential type compressors also
operate rather inefficiency at this 0.degree. F. condition. Some
system manufacturers even shut down their compressors at
temperatures of 0.degree. F. and rely entirely on electric
resistance for a variety of different reasons as well.
Typical Embodiment of a New System
An embodiment of the present invention is directed to a
refrigeration circuit which comprises one first stage booster
compressor, one second stage primary compressor (always the first
to run), a condenser, an economizer, an evaporator, and conduit
means bearing a compressible refrigerant working fluid and
connecting the first stage compressor, the second stage compressor,
the condenser, the economizer, and the evaporator, in series and in
that order, in a closed loop. An important feature of the present
invention is a novel apparatus and method for oil equalization
between the primary and booster compressors.
The closed loop further comprises a bleed line for bleeding a
portion of the condensed refrigerant from the closed loop
downstream of the heating condenser and expanding it within the
economizer for highly subcooling the liquid refrigerant within the
closed loop being fed to the evaporator. The expanded refrigerant
from the economizer is then delivered to a point between the outlet
of the first stage compressor and the inlet to the second stage
compressor. The subcooling of the liquid refrigerant in the
economizer greatly increases the ability of the refrigerant to
absorb heat energy in the evaporator. Also, the vapor created by
this subcooling process significantly increases the refrigerant
weight flow into the heating condenser as it is directly added to
the flow coming form the first stage compressor.
In accordance with this invention, at least the primary compressor
is a multi-cylinder unloadable compressor, such as a Bristol
Twin-Single, or similar type compressor (TS), preferably a 40/100
TS or a 50/100 TS compressor. A Bristol twin-single compressor has
two cylinders and pistons. In one direction of rotation of the
drive shaft, both cylinders/piston are operating (full capacity
operation). In the reverse direction of rotation of the drive
shaft, only one piston/cylinder is operative, and the other
piston/cylinder is idle (partial capacity operation). A TS
compressor is a preferred type of unloadable positive displacement
compressor for use in this invention. In a preferred embodiment,
the booster compressor is a one-speed compressor (although a two
speed or variable speed compressor could also be used). In another
embodiment, both the primary compressor and the booster compressor
are Bristol BS compressors.
While the invention will be described as using the Bristol
twin-single compressor, it will be understood that other steps of
unloadable positive displacement compressors or other unloadable
multi-cylinder positive displacement compressors could be used. For
example, using the designation of 2/1 for a Bristol twin-single
compressor, a 4/2 multi-cylinder compressor (i.e., four cylinders
for fully capacity/two cylinders for partial capacity) or a 6/3
multi-cylinder compressor (six cylinders for full capacity/three
cylinders for partial capacity), etc. could be used.
An important point related to maximum system pumping capacity is
that the first stage (booster) compressor has a larger displacement
(by about 10% to about 50%) than the second stage (primar)
compressor.
The discharge pressure of the booster will rise to the point where
the density (pounds per cubic foot) of the vapor entering the
primary compressor times the primary compressor pumping capacity
(in cubic feet per minute) exactly equals the pounds per minute of
vapor existing the booster compressor plus the pounds per minute of
vapor exiting the economizer. The increased displacement of the
booster (compared to the primary) along with a very high volumetric
efficiency of (because of the low booster discharge pressure)
results in a very high booster flow rate. This very high
refrigerant flow rate multiplied by the increased energy pickup per
pound of refrigerant flowing through the evaporator (because of the
low liquid refrigerant temperature entering the evaporator due to
the economizer), results in a very large increase in the total
amount of energy per minute absorbed from the outside air into the
operating system. This increase also comes about when it is most
needed, i.e., at the lower outdoor ambient air temperatures.
A typical control system for the present invention includes a
transducer for directly sensing the outdoor ambient temperature,
preventing excess system capacity (through utilization of a
micro-processor) until the outdoor temperature reaches a
predetermined low enough value to allow or enable more system
capacity, if called for by the indoor thermoset. The control system
(on heating) also responds to a preferred, three step indoor
thermoset which will step the system heating capacity to various
levels (upon indoor temperature demand) that, in turn, are allowed
or enabled by the various outdoor ambient temperature ranges that
are encountered. The control system, on cooling, also responds to
the same indoor thermoset, which has two cooling steps, which will
step the system cooling capacity to various levels (upon indoor
temperature demand) that, in turn, are allowed or enabled by the
various outdoor ambient temperature ranges that are
encountered.
The preferred displacement ratio of the booster to the primary (at
100% primary compressor flow) is only about 1.3 to 1. This keeps
the pressure ratio across the booster relatively low (whenever it
needs to be low), thus resulting in higher booster volumetric
efficiencies. This displacement ratio thus keeps the economizer
boiling temperature low (again, whenever needed), thus allowing the
extraction of as much energy as possible from the warm liquid
leaving the system condenser. These facts combined with the
utilization of Refrigerant R-410A keep the total displacement
required very low considering the heating capacity level that is
obtained during low ambient heating. It also allows utilization of
relatively low cost smaller reciprocating or other type
compressors.
The above-discussed and other features and advantages of the
present invention will be appreciated and understood by those
skilled in the art from the following detailed description and
drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic diagram of a heating mode operation of a
closed loop boosted air source heating pump of the present
invention.
FIG. 2 is a schematic diagram of a cooling mode operation of a
closed loop boosted air source heat pump of the present
invention.
FIG. 3 is a schematic showing of the shaft, lobe and pistons of a
Bristol Twin-Single compressor.
FIG. 4 is a schematic diagram of lubricant management in accordance
with the present invention when only the primary compressor is
operational.
FIG. 5 is a schematic diagram of lubricant management in accordance
with the present invention when both the primary and booster
compressors are operational.
FIG. 6 is a heating capacity chart for a typical system of the
present invention.
FIG. 7 is a chart showing a typical operating sequence in the
heating mode for a system of the present invention.
FIG. 8 is a cooling capacity chart for a typical system of the
present invention.
FIG. 9 is a chart showing a typical operating sequence in the
cooling mode for a system of the present invention.
DESCRIPTION OF THE PREFERRED EMBODIMENT
Referring to FIGS. 1 and 2, there is shown a closed loop heat pump
system forming an embodiment of the present invention. Referring
first to FIG. 1, the closed loop system includes a first or booster
stage compressor 22, a second or high stage primarily compressor
24, an indoor coil or condenser 26 which delivers heated air to a
space to be heated, an economizer 28, and an outdoor coil or
evaporator 30 which, together with conduit means interconnecting
these elements in a closed loop circuit, are basic components of
the closed loop heat pump system. High stage or primary compressor
24 is normally operating whenever the heat pump system is
delivering energy, but booster compressor 22 and economizer 28 are
operated only when operation is allowed by the control system
depending on outdoor ambient temperature. Warm output vapor of the
primary compressor 24 is fed to the inlet of indoor coil 26 via 4
way valve 80 and conduit segment 32, thus heating air flowing over
indoor coil 26 for delivery to the indoor space to be heated. A
variable speed fan 27 normally controls the flow of air over indoor
coil 20. The warm refrigerant vapor is, of course, cooled and
condensed in indoor coil 26. The outlet of indoor coil 26 delivers
the condensed refrigerant to flow via conduit segment 34 and check
valve 35 to the economizer 28. A bypass or bleed line 38 may permit
a portion of the liquid refrigerant to be bled from the primary
closed loop circuit and to expand via an expansion valve 40 within
economizer 28. However, expansion valve 40 is normally closed, and
it is opened upon receipt of an operating signal from a
microprocessor 54 to allow operation of economizer 28. With
expansion valve 40 in its normally closed position, the refrigerant
passes directly through economizer 28, but without any economizer
action or effect, to conduit segment 42, and then through expansion
valve 76 to outdoor coil or evaporator 30. A fixed or variable
speed fan 31 delivers the air flow over outdoor coil 30. The
refrigerant then flows from evaporator 30 through conduit segment
46a, and through four-way valve 80 to conduit segments 46b and 46c,
and then to the inlet to primary compressor 24.
When operation of the booster compressor 22 is allowed by
microprocessor 54, the refrigerant flows from conduit section 46b
to the inlet to booster compressor 22, and then via conduit 48 from
the discharge from booster 22 to the inlet to primary compressor
24.
Each of components 22 and 24 has its own internal motor, indicated
at 23 and 25, respectively. Each motor is connected to
microprocessor 54, and the operation of each compressor is allowed
only by the presence of an activating signal from microprocessor 54
to the compressor motor to operate the compressor.
The system also includes a temperature transducer 56, such as a
thermistor, at outdoor cell 30 to sense the temperature of the
outdoor air flowing over outdoor coil 30, a temperature transducer
58, such as thermistor, at indoor coil 26 to sense the temperature
of air leaving indoor coil 26, and an indoor thermostat 62 which
senses the temperature of the air in the space to be heated and
sends signals to microprocessor 54 when heat is required or when
the desired temperature has been attained. Transducers 56 and 58
and thermostat 62 are connected to deliver signals to
microprocessor 54. The signals received at microprocessors 54 from
outdoor ambient temperature transducer 54 are used to allow various
combinations of operation of the primarily compressor 24, booster
compressor 22 and economizer 28 as a function of outdoor ambient
temperature to meet heating requirements; and the signals received
at microprocessor 54 from transducer 58 are used to reduce the
speed of fan 27 to reduce or eliminate the "cold blow" problem
common to heat pump systems (by reducing air flow at higher ambient
temperatures).
In the present invention, thermostat 62 has three operating points
or stage for heat operation and two operating points or stages for
cooling operation. A thermostat of this type may, be e.g., a
Minneapolis Honeywell thermostat type T8611M2005 or T8511M1002,
available from Minneapolis Honeywell.
In the embodiment now being discussed of the present invention,
both primary compressor 24 and booster compressor 22 are
Twin-Single compressors available from Bristol Compressors of
Bristol, Va. Referring to FIG. 3, the Bristol Twin-Single is a
compressor reciprocating compressor having two pistons 202 and 204,
mounted on a shaft 206. Shaft 206 can be rotated either clockwise
or counterclockwise. Rotatable eccentric lobe 208 is also mounted
on shaft 206. When shaft 206 is rotating, one of the pistons, e.g.,
piston 202, is always reciprocating, regardless of the direction of
rotation of shaft 106. When shaft 206 is rotating in the direction
shown in FIG. 3A, lobe 208 is positioned off center of the axis of
shaft 206 so that piston 204 also reciprocates (along with piston
202). However, when shaft 206 is rotated in the opposite direction,
as indicated in FIG. 3B, lobe 208 is repositioned on the center
axis of shaft 206, whereby piston 204 is idle, i.e., does not
reciprocate, and only piston 202 reciprocates.
It will be understood that the cylinders in which the pistons 202
and 204 reciprocate are not shown.
The fluid flow capacity of a Bristol TS compressor can be split,
i.e., allocated, as desired between the two pistons/cylinders. For
example, the capacity can be split between 40%/100% to 50%/100% or
somewhat larger ratios, where 100% is the flow capacity when both
cylinders are reciprocating, and where the lower number is the
percentage of total flow capacity when only one piston is
reciprocating. For the present invention, a 50/100 split or a
40/100 split is preferred. A split of 40/100 provides adequate
heating capacity at higher ambient temperatures where operation of
only one cylinder of the primary compressor is required.
In a second embodiment of the present invention, only primary
compressor 24 is a Bristol TS of similar type compressor. In this
second embodiment, booster compressor 22 is either a two speed
compressor, or a single speed (fixed displacement) compressor, the
latter resulting in the lowest manufacturing cost for the system.
In this case, the booster compressor can be any type of positive
displacement compressor.
Also, is preferred that the compressors be sized so that (1) 100%
of the capacity of primary compressor 24 be equal to the rated
capacity normally required for cooling by Air Conditioning and
Refrigerant Institute (ARI) standards, and (2) the displacement
ratio of booster compressor 22 to primary compressor 24 be in the
range of 1.1:1-1.7:1, preferably about 1.3:1.
Operating sequences for heating and cooling will be set forth and
discussed below. At various points in the operating sequences,
valve 40 will be operated by a signal from microprocessor 54 to
bleed and expand refrigerant fluid from point 36 in line 34 to
economizer 28. The expansion of the refrigerant in economizer 28
results in significant subcooling of the main body of liquid
refrigerant which flows in a closed conduit through economizer 28.
This subcooled liquid refrigerant then passes directly to
evaporator 30 via conduit segment 42. This highly subcooled liquid
refrigerant expands via expansion valve 44 into and within the
evaporator 30 to perform the function of absorbing energy from the
outside air flowing over outdoor coil 30 and vaporizing in
evaporator 30. The amount of energy absorbed within evaporator 30
is greatly increased because of the highly subcooled refrigerant
delivered from economizer 28 to the evaporator. The refrigerant
vapor from evaporator 30 then flows via conduit segment 46a, 46b
and 46c and check valve 47 to point 52 and via conduit segment 48
to the suction or low side of primary compressor 24 to complete the
closed loop circulation in effect when only the primary compressor
24 is operating, or to the suction or low side of booster
compressor 22 if both compressors are operating.
Meanwhile, the refrigerant bled via line 38 which vaporizes within
the economizer to perform the cooling effect in the economizer,
passes via conduit segment 50 to point 52 in conduit 48 connected
to the inlet of the primary compressor 24.
The heating mode of operating of the heat pump system is shown in
FIG. 1, and the cooling mode of operation is shown in FIG. 2. In
the heat mode operation of FIG. 1, the refrigerant flows through
the closed loop conduit in the direction shown by the arrows in the
conduit. To change from the heating mode of FIG. 1 to the cooling
mode of FIG. 2, four way valve 80 is operated, as by a mode
selection signal from thermoset 62 or microprocessor 54, whereby
the direction of refrigerant flow in the closed loop conduit is
reversed, as indicated by the arrows in FIG. 2. In the cooling
mode, indoor coil 26 functions as an evaporator, and outdoor coil
30 functions as a condenser.
Lubricant, e.g., oil, management is an important aspect of the
present invention. With two compressors connected to series, a
potential exists for most or all of the lubricant in the system to
accumulate in the sump of one of the compressors, and for the other
compressor to become starved for lubricant. That, of course, can
lead to failure of the lubricant-starved compressor. The present
invention addresses and solves this problem.
Reference is made to FIG. 4 and 5, which are side schematic views
of a compressor module housing the booster and primary compressors
22, 24. Parts in FIGS. 4 and 5 are numbered as in FIG. 1.
As shown in FIG. 4, only the primary compressor is operational.
Each of compressors 22 and 24 has a reservoir of oil, respectively
104 and 106 in the sump of each compressor. The compressors also
have aspiration tubes 108, 110, respectively, from the sump to the
cylinder intake. The tubes 108, 110 operate to prevent accumulation
of lubricant above the lower level of the tubes when each
compressor is operating. When the lubricant level rises above the
lower level of a tube, the tube sucks lubricant from the sump into
the cylinder intake when a compressor is operating. The lubricant
is then entrained as liquid droplets in the circulating refrigerant
for circulation through the system, and the lubricant droplets then
return and drop into the compressor sump when the refrigerant
enters the compressor intake.
The refrigerant and entrained lubricant, as indicated by the
arrows, flows into the inlet to the interior of compressor shell or
can 150. The lubricant droplets fall into sump 106, and the
refrigerant gas flows through the holes 150 of the compressor motor
to cool the motor, and the refrigerant gas then pass flows to the
intake manifold to the cylinders 154 of primary compressor 24. The
lubricant in the sump 106 is at the pressure of the refrigerant
entering can 150, which is slightly higher than the pressure of the
refrigerant entering the intake 154 to the cylinders. Aspiration
tube 110 extends from the lubricant sump to the point of lowest
pressure in the intake to the cylinders of the primary
compressor.
In the system as shown in FIG. 4, since booster 22 is not
operating, it is important to prevent the lubricant from entering
into and accumulating in the sump of the booster. If the lubricant
were permitted to enter into booster 22, the lubricant would merely
accumulate in the sump of booster 22, since compressor 22 is
inoperative and, therefore, no aspiration occurs through tube 108.
This would eventually result in adequate lubrication for primary
compressor 24.
Entry of the lubricant into the booster 22 when the booster is
inoperative is prevented by a pair of traps 112 and 114. Trap 112
in conduit section 100 prevents entry of lubricant from conduit
segment 46b, and trap 114 in conduit section 48 prevents entry of
lubricant from conduit segment 48. Accordingly, all circulating
lubricant is directed to the operating primary compressor 24.
If the level of lubricant 106 in the sump of primary compressor 24
rises above the bottom of siphon tube 110, the excess accumulation
will be aspirated into the intake of the primary compressor and
then circulates with the refrigerant vapor leaving the
compressor.
Referring to FIG. 5, which shows both compressors operational, the
refrigerant and entrained oil flow via conduit 100 to the inlet to
booster can or shell 110. The lubricant droplets fall into the
lubricant sump 104, and the refrigerant gas flows through the holes
162 of the compressor motor to cool the motor, and the refrigerant
gas then flows to the intake manifold to the cylinders 164 of
booster compressor 22. The lubricant in sump 104 is at the pressure
of the refrigerant entering can 160, which is slightly higher than
the pressure of the refrigerant entering the intake 164 to the
cylinders. Aspiration tube 108 extends from the lubricant sump to
the point of lowest pressure in the intake to the cylinders of the
booster compressor.
When both compressors 22 and 24 are operating (see FIG. 5), the
circulating refrigerants and any entrained lubricant, are drawn
from conduit segment 46b through trap 112 and conduit segment 100
into the shell of booster 22. The refrigerant discharge from
compressor 22 then flows through conduit segment 48 and trap 114 to
the intake manifold of compressor 24. If the level of lubricant 104
in the sump of compressor 22 rises above the bottom of tube 108,
the excess lubricant is aspirated into the cylinder of compressor
22 and then entrained in the refrigerant fluid delivered through
conduit 48 and trap 114 to the intake to compressor 24. Similarly,
if the level of sump lubricant 106 rises above the bottom of the
tube 110, the lubricant is aspirated into the cylinder of
compressor 24 and is then circulated as before. In this way, the
sump lubricant levels in both compressors are maintained at desired
levels, and both compressors are lubricated.
While the lubricant management has been described for a pair of
compressors connected in series, the same system of traps and
siphon tubes can be used with compressors connected in parallel.
The tubes maintain desired levels of lubricant when in each
operating compressor, and the traps prevent delivery and build-up
of lubricant in compressor 22 when it is not operating.
EXAMPLE 1
Referring now to FIGS. 6 and 7, a heating capacity chart and an
exemplary operating sequence are shown. It will be recalled that
thermostat 62 preferably has three stages in the heat mode. The
thermostat stages signal microprocessor 54, which, in turn, sends
signals to allow (i.e., control) the operation of the compressors,
and/or one or both cylinders of the compressors, and/or the
economizer.
Typically, the heating cycle starts when the first stage or step of
indoor thermostat 62 calls for heat. When this occurs somewhere
between 75.degree. F. and above 50.degree. F. outdoor ambient
temperature, as sensed by sensor 56, one piston of primary
compressor 24 is allowed to operate, i.e., is activated by a signal
from microprocessor 54. Depending on the configuration, this
provides 40% or 50% of the displacement of primary compressor 24.
This mode of operation is identified in FIGS. 6 and 7 as 010,
signifying 0 cylinder operation of the booster, 1 cylinder
operation of the primary and no operation of the economizer.
When the ambient temperature drops to about 50.degree. F.,
operation of the second piston of primary compressor 24 is allowed
by microprocessor 54, but only if called for by the second stage of
thermostat 62. This mode of operation is indicated by the 020 lines
on FIGS. 6 and 7.
No additional heating capacity can be brought on line until the
outdoor ambient further drops to about 43.degree. F. or so, even if
the third step of the indoor thermostat calls for more heat. This
is designed to prevent the system from supplying more capacity than
is really needed, as, if it were to be supplied, it would come
about at a low efficiency level because the condenser would operate
at an unnecessarily high pressure and the evaporator would operate
at an unnecessary low pressure.
When the outdoor ambient temperature reaches about 43.degree. F.,
microprocessor 54 allows operation of both cylinders of booster
compressor 22 (100% booster operation), but with operation of only
one cylinder (40% or 50% displacement) of the primary compressor,
and without operation of the economizer. This mode is indicated at
the 210 lines in FIGS. 6 and 7. This mode becomes the maximum
capacity heat capacity allowed until the outdoor ambient
temperature drops to about 33.degree. F. Then, microprocessor 54
also allows operation of the economizer 28 by sending a signal to
open valve 40, but this signal is sent only if the third stage or
step of indoor thermostat 62 calls for more heat. This mode of
operation is indicated at the 211 line in FIGS. 6 and 7.
When outdoor ambient temperature reaches about 25.degree. F. a
signal from microprocessor 54 allows operation of both cylinders of
primarily compressors 24, along with both booster cylinders and the
economizer. This mode of operation is indicated at the 221 lines in
FIGS. 6 and 7. This is the maximum capacity heat pump mode, and it
continues in operation until the outdoor ambient temperature
reaches about 15.degree. F.
At the 15.degree. F. outdoor ambient temperature level, and if the
third step of the indoor thermostat is calling for more heat,
microprocessor 54 sends a signal to allow operation of back-up
electric resistance heating.
As seen from FIG. 6, the BSHP (boosted source heat pump) of this
invention meets the heating requirement without the need for
back-up resistance heating all the way down to an outdoor ambient
temperature of about 10.degree. F. This is far superior to a
typical prior art heat pump, the capacity line of which is labeled
"HP-TODAY" in FIG. 6, where back-up resistance heat is required at
about 30.degree. F. outdoor ambient temperature. Bearing in mind
that the high cost of resistance back-up heat is one of the main
disadvantages of typical prior art heat pumps, the significant
advantages of the present invention are apparent.
EXAMPLE 2
While incorporation of economizer 28 in the system is preferred,
the economizer can be omitted. In that case, the displacement ratio
of the booster compressor 22 to primary compressor 24 would be
increased sufficiently to realize a system capacity about that of
the system with the economizer, understanding that a system
efficiency loss would occur due to the absence of the economizer.
In this case, at conditions of 0.degree. F. outdoor ambient and
70.degree. F. indoor heated space temperature, the heating
coefficient of performance (C.O.P.) will be at least 1.5 and may
approach 2. By way of example, the displacement ratio of booster
compressor 22 to primary compressor 24 could be increased to about
1.4:1 to about 1.7:1. Referring to FIGS. 6 and 7, with the
economizer eliminated, the 211 line would be eliminated, and the
211 line is replaced by a 220 line, with the allowance point being
the outdoor ambient temperature at which the 221 line was
previously allowed, i.e., between about 15.degree. F.-25.degree. F.
in Example 1. The 220 line is shown as a dashed line in FIG. 6.
FIG. 8 shows cooling performance for the heat pump of the present
invention, and FIG. 9 shows a typical cooling operating
sequence.
With outdoor ambient temperature of about 80.degree. F., and with
the first stage of thermostat 62 calling for cooling,
microprocessor 62 allows operation of only one piston (40%-50%
capacity) of the primary compressor 24. It is expected that this
will handle most of normal cooling requirements. This is the 010
line in FIG. 8.
At about 85.degree. F. outside ambient temperature, microprocessor
54 allows operation of only both pistons of primary compressor 24
(100% primary capacity) if called for by the thermostat. This is
indicated at the 020 line in FIGS. 8 and 9.
At about 105.degree. F. outdoor ambient, microprocessor 54 allows
operation of both pistons of booster compressor 22 and both pistons
of primary compressor 24. (100% capacity for both compressors).
This is indicated at the 220 line of FIGS. 8 and 9. This will be
effective to meet cooling needs up to about 115.degree. F. outdoor
ambient.
In addition to the foregoing, operation of both primary pistons,
both booster pistons, and the economizer can be manually selected
for special requirements, e.g., quick cool down, or to handle large
numbers of people in a room, or high humidity conditions, etc. This
is indicated at the 221 line in FIGS. 8 and 9.
It will be noted in FIGS. 6, 7, 8 and 9, that whenever operation of
the booster compressor is allowed, both cylinders (100% capacity)
are utilized. This means that variable capacity is not required for
booster 22. Accordingly, booster 22 can be a single speed
compressor (or any type). This will reduce the manufacturing cost
of the system, since single speed compressors can be obtained less
expensively than the Bristol or similar type TS compressor.
It will be understood that the operating modes and sequences
illustrated in FIGS. 6-8 are only be way of example. Other
operating modes and sequences can be employed within the scope and
interest of this invention.
While this invention has been described in terms of a system for
both heating and cooling, the invention can be applied for a
heating system alone or a cooling system alone. In that event, the
four-way valve 80 would be eliminated, and the refrigerant would
always flow in one direction only.
If the booster compressor fails to operate when called for (for
example because of an electrical contractor problem) the
microprocessor is programmed to sense the non-operation of the
booster and to proceed to a single TS mode of operation (for the
primary compressor). The heating operating sequence of this single
TS mode is as follows, with reference to FIG. 7: 1. From 50.degree.
F.-75.degree. F. outdoor ambient temperature, only one cylinder of
the primary compressor is allows (010 operation). 2. From
43.degree. F.-50.degree. F., operation of both cylinders of the
primary compressor only are allowed (020 operation). 3. From any
temperature below and up to 43.degree. F., backup heat will be
allowed, along with operation of both cylinders of the primary
compressor. It will be noted that in this mode of operation, the
210, 211 and 221 steps of operation are eliminated, because the
booster is inoperative.
In this mode of operation, where only the TS primary compressor is
greater, the cooling sequence is also varied by the
microprocessor.
Since the system is sized to deliver its rated cooling capacity at
normal AR1 operating conditions with a 0-2-0 combination, it would
just operate as a typical system of today does with the exception
that 0-2-0 would not be allowed until 85.degree. F. outdoor ambient
or thereabout whereas today, it is allowed whenever the indoor
thermostat would call for it. In this case as mentioned above
(booster failing to operate for some reason), the second step of
the cooling thermostat could call for 0-2-0 at any outdoor ambient
about 65.degree. F. or so. The first step of the cooling thermostat
would still call for 0-1-0 as long as the outdoor ambient
temperature is above 60.degree. F. or so.
Accordingly, the one TS sequence of operation for cooling is as
follows: 1. When the indoor thermostat calls for step 1 of cooling,
operation of only one cylinder of the primary compressor will be
allowed (the 010 mode) by the microprocessor as long as the sensed
outdoor ambient temperature is in the range of about 60.degree.
F.-85.degree. F. 2. When the indoor thermostat calls for second
stage cooling, and the outdoor ambient temperature rises to about
85.degree. F., operation of both stages of the TS primary
compressor will occur (the 020 mode). 3. The operation of both
cylinders of the primary TS compressor will be allowed on manual
selection of the second step of the indoor thermostat and as long
as outdoor ambient temperature is above about 60.degree. F. (the
020 mode).
While preferred embodiments have been shown and described, various
modifications and substitutions may be made thereto without
departing from the spirit and scope of the invention. Accordingly,
it is to be understood that the present invention has been
described by way of illustrations and not limitations.
* * * * *