HYBRID COMPRESSOR SYSTEM AND METHODS
CROSS-REFERENCE TO RELATED APPLICATION
[0001] Benefit is claimed of US Patent Application Ser. No. 61/491,515, filed May 31,
201 1, and entitled "Hybrid Compressor System and Methods", the disclosure of which is
incorporated by reference herein in its entirety as if set forth at length.
BACKGROUND
[0002] The disclosure relates to refrigeration. More particularly, the disclosure relates to
chiller systems.
[0003] Large water-cooled chillers (e.g., 300-1500 ton capacity (1055-5275W)) typically
use a single centrifugal compressor for cost reasons. However, such compressors are subject
to surge (particularly at partial load).
[0004] One way of addressing this has been to provide multiple centrifugal compressors
in parallel. This allows individual compressors to be taken off line to better match
compressor capacity to required load.
SUMMARY
[0005] One aspect of the disclosure involves an apparatus having a centrifugal
compressor, a positive displacement compressor, a first heat exchanger, and a second heat
exchanger. A plurality of valves are positioned to provide operation in at least two modes. In
a first mode, refrigerant is compressed in the positive displacement compressor and the
centrifugal compressor at least partially in parallel. In a second mode, refrigerant is
compressed in the positive displacement compressor and the centrifugal compressor is
offline. In a third mode, refrigerant is compressed in the positive displacement compressor
and the centrifugal compressor at least partially in series.
[0006] In various implementations, the positive displacement compressor may be a screw
compressor.
[0007] Further aspects of the disclosure involve operating in such at least two modes.
[0008] The details of one or more embodiments are set forth in the accompanying
drawings and the description below. Other features, objects, and advantages will be apparent
from the description and drawings, and from the claims.
BRIEF DESCRIPTION OF THE DRAWINGS
[0009] FIG. 1 is a schematic view of a chiller system.
[0010] FIG. 2 is a longitudinal vertical schematic view of a condenser of the system of
FIG. 1.
[0011] FIG. 3 is a longitudinal vertical schematic view of a cooler of the system of FIG.
1.
[0012] FIG. 4 is a plot of entering condenser water temperature against percent capacity.
[0013] FIG. 5 is a control flowchart for the system of FIG. 1.
[0014] Like reference numbers and designations in the various drawings indicate like
elements.
DETAILED DESCRIPTION
[0015] In addition to suffering from surge problems, centrifugal compressors are
typically less effective than positive displacement compressors at providing high head. This
renders centrifugal compressors as poor candidates for heat reclaim operation. As is
discussed below, a hybrid system features a centrifugal compressor and a positive
displacement compressor. The exemplary positive displacement compressors are two- or
three-rotor screw compressors powered by variable frequency drives. Alternative positive
displacement compressors include reciprocating compressors and scroll compressors.
[0016] FIG. 1 shows a vapor compression system 20 having a compressor subsystem 22.
The compressor subsystem 22 includes a first compressor 24 (centrifugal) and a second
compressor 26 (positive displacement). In the exemplary implementation, both compressors
have capacity control features which may be of any well known type (e.g., variable inlet
guide vanes for the centrifugal compressor and a slide valve for a screw compressor used as
the positive displacement compressor and variable speed drives for both compressors).
[0017] In a first mode of operation, the compressor subsystem drives refrigerant in a
downstream direction 500 along a refrigerant flowpath 30. The flowpath 30 passes,
sequentially, through a first heat exchanger 32, an expansion device 34, and a second heat
exchanger 36. In the first mode, the first heat exchanger 32 is a heat rejection heat exchanger
and the second heat exchanger 36 is a heat absorption heat exchanger.
[0018] An exemplary system 20 is a chiller system wherein the first heat exchanger 32 is
a liquid-cooled condenser or gas cooler and the second heat exchanger 36 is the cooler. An
exemplary expansion device 34 is an electronic expansion valve (EV) which may be
controlled by the chiller's controller 40 (e.g., a computer or microcontroller). An alternative
expansion device 34 is a float valve within the condenser 32. The exemplary first heat
exchanger has at least one inlet 50 and at least one outlet 52 along the refrigerant flowpath
30. Similarly, the second heat exchanger 36 has at least one inlet 54 and at least one outlet 56
along the refrigerant flowpath. The compressor 24 has an inlet port 60 and an outlet port 62.
Similarly, the second compressor 26 has an inlet port 64 and an outlet port 66.
[0019] As is discussed further below, the compressor subsystem includes one or more
valves coupled to the compressors to allow switching of the compressors between two or
more compression modes. The exemplary system includes three valves 70, 72, and 74.
[0020] In a first mode of operation, the compressors are operated at least partially in
parallel. In an exemplary illustrated fully parallel situation, the respective suction ports and
discharge ports of the compressors are at essentially identical conditions. To provide such
parallel operation, the exemplary flowpath has two parallel branches 80 and 82 diverging at a
junction 84 downstream of the second heat exchanger outlet 56 and re-merging at a location
86 at or upstream of the first heat exchanger inlet 50. In partially parallel situations, the
separation and/or rejoinder may be at different locations.
[0021] A bypass branch or line 90 extends between the branches 80 and 82. The
exemplary bypass branch 90 extends between upstream of one of the compressors to
downstream of the other. In the exemplary implementation, the branch extends from a
location 92 downstream of the first compressor to a location 94 upstream of the second
compressor. For controlling flow in the branches 80 and 82, the exemplary valves 70 and 74
are respectively along such branches. In an exemplary implementation, the valve 70 is
downstream of the first end 92 of the bypass line 90 and the valve 74 is upstream of the end
94.
[0022] In the exemplary at least partially parallel operation, the valve 72 is closed
whereas the valves 74 and 70 are open. In a second mode, only the second compressor 26 is
in operation. The valves 70 and 72 are closed whereas the valve 74 is open. A third mode is a
series mode wherein the compressors are operated in series. In the exemplary series mode,
the valves 70 and 74 are closed whereas the valve 72 is open. Refrigerant passes without
diversion from the second heat exchanger outlet 56 through the first compressor, the valve
72, and the second compressor before entering the first heat exchanger inlet 50. A fourth
possible mode involves having only the first compressor 24 in operation. In this mode, the
valves 72 and 74 are closed and the valve 70 is open.
[0023] FIG. 1 shows further exemplary details of the condenser 32 and cooler 36. The
exemplary condenser 32 includes an upper condenser tube bundle 120 and a lower subcooler
tube bundle 122. FIG. 1 also shows a liquid refrigerant accumulation 124 within the
condenser. The tube bundles 120 and 122 are connected to one or more sources of heat
transfer fluid to withdraw heat from the refrigerant. The sub-cooler tube bundle 122 is
contained within a chamber 126. One or more inlet orifices 128 are along the bottom of the
chamber 126. A float valve 130 feeds the outlet 52. A pressure sensor 132 may be located in
the headspace of the condenser near the inlet 50.
[0024] In an exemplary implementation, the heat transfer fluid (e.g., water) passes along
a water loop 138 (FIG. 2) and is received via an inlet 140 and discharged from an outlet 142.
Respective temperature sensors 144 and 146 measure inlet temperature TicoND and outlet
temperature T2COND of the water. An exemplary flow meter 147 along the water loop 138
measures a flow rate FMCOND of the water.
[0025] The cooler 36 also includes a lower tube bundle 160 and an upper tube bundle
162. FIG. 1 further shows a refrigerant accumulation 164 in the cooler. In an exemplary
implementation, a heat transfer fluid (e.g., water) passes along a water loop 168 (FIG. 3) and
is received via an inlet 170 and discharged from an outlet 172. Respective temperature
sensors 174 and 176 measure inlet temperature T1COOL and outlet temperature T2COOL of the
water. An exemplary flow meter 177 along the water loop 168 measures a flow rate FMCOOL
of the water. FIG. 1 further shows a distributor 180 in the lower portion of the cooler
approximately fed by the inlet 54. A pressure sensor 182 is shown in the headspace near the
outlet 56.
[0026] FIG. 4 shows a plot of the entering condenser water temperature T1COND against
capacity. Line 200 represents the American Refrigeration Institute (ARI) load line. In the
United States, chillers are subject to ARI Standard 550. This standard identifies four
reference conditions characterized by a percentage of the chiller's rated load (in tons of
cooling) and an associated condenser water inlet/entering temperature. Operation is to
achieve a chilled water outlet/leaving temperature of 44F(6.67C). The four conditions are:
100%, 85F (29.44C); 75%, 75F (23.89C), 50%, 65F (18.33C); and 25%, 65F (18.33C also).
These conditions (or similar conditions along a curve of connecting them) may provide
relevant conditions for measuring efficiency. In API testing, the water flow rate through the
cooler is 2.4 gallons per minute per ton of cooling (gpm/ton) (0.043 liters per second per
kilowatt (1/s/kW)) and condenser water flow rate is 3gpm/ton (0.054 1/s/kW).
[0027] With typical heat exchangers, the water temperature rise across the condenser is
approximately 8F (4.4C) times the percentage load or 8F at 100% load, 6F (3.3C) at 75%
load, 4F (2.2C) at 50% load and 2F (1.1C) at 25% load. The cooler saturation temperature is
IF (0.6C) or 2F (1.1C) below the leaving chilled water temperature (e.g., 43F in the ARI
test). Similarly, the condenser saturation temperature is IF or 2F above the sum of entering
condenser water temperature and water temperature rise. For entering condenser water
temperature of 85F, temperature rise is 8F, hence leaving condenser water temperature is 93F
and condenser saturation temperature is 93+2=95F.
[0028] Line 202 represents a constant temperature of 85F (29.44C). In tropical regions,
the ambient temperature changes very little from day to night. In such regions, the condenser
water temperature remains constant. It' s an industry standard, to consider the entering
condenser water temperature constant at 85F between 25% and 100% load. Table I shows lift
for the ARI conditions and corresponding tropical conditions. Centigrade temperatures are
conversions from the listed Fahrenheit values and thus do not add and present false precision.
Other SI parentheticals herein similarly represent conversions from the original US or
English values.
Table I
[0029] In a first exemplary implementation, the at least partially parallel first mode is
utilized at high loads and the second mode (positive displacement compressor-only) is used at
low loads. For example, the second mode may be used from essentially zero load to an
intermediate load value. Between the intermediate load value and the maximum load, the at
least partially parallel mode is used. It is noted that the intermediate load value may,
however, be subject to appropriate hysteresis control to avoid excessive cycling when
operating near changeover conditions. For example, from zero to changeover, the second
compressor may be operated at increasing speed and/or power. At changeover, the first
compressor may be brought online at full or near full capacity and the second compressor
reset to zero or other low capacity value. Thereafter, with increasing load, the speed and/or
power of the second compressor may be increased. Thus, in the at least partially parallel
mode system load/capacity variations are principally accommodated by the positive
displacement compressor. For example, at a minimum, the positive displacement compressor
may address more of the variation than the centrifugal compressor does. More narrowly, the
positive displacement compressor may address at least 75% or at least 90% of the load
variation. The load variation may represent at least an exemplary 30% of a peak load of the
system, more narrowly, at least 40%.
[0030] In the first, simple, exemplary implementation, the rated capacities of the two
compressors are essentially the same (e.g., the same or appropriately differently sized to
address any hysteresis issues). The changeover point is, therefore, at essentially half load.
[0031] More broadly, the changeover point may be between 45% and 55% or 40% and
60% of the total rated system load.
[0032] By using the centrifugal compressor only at or near its own rated load (or, more
broadly, not at a low load) issues of surge may largely be avoided.
[0033] For example, an exemplary rated maximum capacity of the positive displacement
compressor is 50-200% of the rated maximum capacity of the centrifugal compressor, more
narrowly, 100 to 150%.
[0034] In one variation, the fourth (series) mode may be added and used at high
condenser water temperatures such as a water heating or a heat reclaim mode.
[0035] In a further variation, the centrifugal compressor may be used alone when very
low lift is needed (e.g., less than 25F (13. 9C)).
[0036] In a second, more complex, exemplary implementation, a control process 300
(FIG. 5) starts by measuring or otherwise determining 302 the saturation temperatures of the
condenser (TCOND) and the cooler (TCOOL) - TCOND and TCOOL may respectively be determined
by measuring the pressures via the pressure sensors 132 and 182 and then calculating the
saturation temperatures (either via a lookup table or programmed function). The lift is
calculated 304 as TCOND minus TCOOL- I the lift is greater than a given threshold (e.g., 50F
(28C)) the system may be operated 306 in the fourth (series) mode. In the series mode, the
capacity of the centrifugal compressor is controlled by compressor speed and by inlet guide
vane orientation. If the volume of discharge gas from the centrifugal compressor is higher
than the capacity of the screw compressor, then the pressure between the centrifugal
compressor and the screw compressor will rise resulting in surge of the centrifugal
compressor. At this point, the centrifugal compressor speed is incrementally increased and its
guide vanes are incrementally closed until the centrifugal compressor comes out of surge. A
similar logic is applied for the screw compressor (i.e., first speed followed by slide valve).
Reducing the speed always results in reduced power consumption or increased efficiency.
[0037] For lesser lift, a non-series operation may be performed. Measurements 308 are
made of the flow rate FMCOOL and the temperatures TicooL and T2COOL- Capacity may also be
calculated 310.
[0038] At low capacity (e.g. less than a first value such as 50% of a maximum) operation
is then refined based upon the head. For low head, the compressors may be run 320 in the
first mode at equal loads. This may involve controlling capacity via the speed when variable
speed drive is present and by the centrifugal compressor inlet guide vanes and the screw
compressor slide valve for fixed speed case. Head is proportional to the temperature lift. In
the example, low head corresponds to temperature lift less than 35F (19C) and high head
between 35F and 50F (19C and 28C).
[0039] For high head, the system is operated 322 in the second mode (screw
compressor-only). Capacity is controlled via speed when variable speed drive is present and
by slide valve for fixed speed case.
[0040] At intermediate capacity (e.g., 50-75%), operation may also be in the first mode.
Balance between the compressors may be refined based upon the head. For low head, the
system is run in the parallel mode with the screw compressor operating at a fixed capacity
and the centrifugal compressor operating at variable capacity to provide the required overall
capacity. For example, with centrifugal and screw compressors of maximum equal capacity,
the screw compressor may be operated at 50% of its maximum capacity and the centrifugal
compressor being operated at between 50 and 100% of its maximum capacity (thereby
combining to provide the exemplary 50-75% of maximum system capacity operation). Such
an operation is chosen so as to avoid surge of the centrifugal compressor.
[0041] At high head, the system may be run 332 in the first mode with the centrifugal
compressor at essentially fixed capacity and the screw compressor providing capacity control.
The exemplary set points of the constant capacity compressor may differ relative to the
condition 330. The centrifugal compressor may be run at a relatively high capacity. In the
foregoing example, this may be at 80% at its maximum capacity thereby providing 40% of
total system capacity. The screw compressor may be run between 20 and 70% of its
maximum capacity (thereby providing 10-35% of maximum system capacity and combining
with the centrifugal compressor to provide the 50-75% of maximum system capacity). Such
operating condition may be selected because the centrifugal compressor is susceptible to
surge at low loads and high head.
[0042] At high capacity, (e.g., 75-100%), operation may also be in the first mode, with
compressors running 340 at equal loads. Thus, depending upon the necessary capacity, each
may be run at between 75 and 100% of its own maximum capacity to satisfy the required
capacity.
[0043] One or more embodiments have been described. Nevertheless, it will be
understood that various modifications may be made. For example, in reengineering an
existing system, details of the existing system or its intended use may influence details of any
particular implementation. Accordingly, other embodiments are within the scope of the
following claims.

CLAIMS
What is claimed is:
1. An apparatus (20) comprising:
a centrifugal compressor (24);
a positive displacement compressor (26);
a first heat exchanger (32);
a second heat exchanger (36); and
a plurality of valves (70, 72, 74) positioned to provide at least two of:
operation in a first mode wherein:
refrigerant is compressed in the positive displacement compressor and
the centrifugal compressor at least partially in parallel;
operation in a second mode wherein:
refrigerant is compressed in the positive displacement compressor; and
the centrifugal compressor is offline; and
operation in a third mode wherein:
refrigerant is compressed in the positive displacement compressor and
the centrifugal compressor at least partially in series.
2. The apparatus of claim 1 wherein:
the positive displacement compressor is a screw compressor.
3. The apparatus of claim 1 wherein:
the centrifugal compressor has a rated maximum capacity; and
the positive displacement compressor has a rated maximum capacity of 100-150% of
the rated maximum capacity of the centrifugal compressor.
4. The apparatus of claim 1 wherein:
the plurality of valves are positioned to provide all three of said modes.
5. The apparatus of claim 1 wherein:
the plurality of valves are positioned to provide at least the first mode and second
mode.
The apparatus of claim 1 being a chiller wherein:
the first heat exchanger is part of a condenser unit;
the second heat exchanger is part of a cooler; and
an expansion device (34) is positioned between the condenser and the cooler.
7. The apparatus of claim 1 wherein the plurality of valves comprises:
a first valve (70) between the centrifugal compressor and the first heat exchanger;
a second valve (72) along a bypass (90) extending from a location (92) between the
centrifugal compressor and the first valve to a location (94) between the second valve and the
screw compressor; and
a third valve (74) between the second heat exchanger and the positive displacement
compressor.
8. The apparatus of claim 1 further comprising:
a controller (40) programmed to automatically switch between said at least two
modes.
9. The apparatus of claim 8 wherein:
the controller is programmed to automatically switch between said first mode and said
second mode.
10. The apparatus of claim 9 wherein:
said controller is programmed to switch from said first mode to said second mode
responsive to a decrease in load and from said second mode to said first mode responsive to
an increase in load.
11. The apparatus of claim 10 wherein:
the controller is programmed to switch to said third mode responsive to calculating a
high requirement for lift.
12. An apparatus comprising:
a centrifugal compressor (24);
a positive displacement compressor (26);
a first heat exchanger (32);
a second heat exchanger (36); and
means (70, 72, 74) for providing at least two of:
operation in a first mode wherein:
refrigerant is compressed in the positive displacement compressor and
the centrifugal compressor at least partially in parallel;
operation in a second mode wherein:
refrigerant is compressed in the positive displacement compressor; and
the centrifugal compressor is offline; and
operation in a third mode wherein:
refrigerant is compressed in the positive displacement compressor and
the centrifugal compressor at least partially in series.
13. A method for operating a vapor compression system (20) having a centrifugal
compressor (24) and a positive displacement compressor (26), the method comprising at least
two of:
operating in a first mode wherein:
refrigerant is compressed in the positive displacement compressor and the
centrifugal compressor at least partially in parallel;
operating in a second mode wherein:
refrigerant is compressed in the positive displacement compressor; and
the centrifugal compressor is offline; and
operating in a third mode wherein:
refrigerant is compressed in the positive displacement compressor and the
centrifugal compressor at least partially in series.
14. The method of claim 13 wherein:
the system is operated in the second mode from a minimum load condition to an
intermediate load condition; and
the system is operated in said first mode at loads above said intermediate load
condition.
15. The method of claim 14 wherein:
in the first mode, load variation is principally accommodated by the positive
displacement compressor.
16. The method of claim 13 wherein:
operation is based on a combination of required capacity and required lift and wherein
some-to-all of:
in a high required lift, the system is operated (306) in the third mode; and
at low required lift, some-to-all of:
at low capacity:
the system is run (320) in the first mode; and
at high head the system is run (322) in the third mode;
at intermediate capacity:
at low head, the system is run (330) in the third mode with
essentially constant positive displacement compressor capacity and the
centrifugal compressor at variable capacity; and
at high head, the system is run (322) in the first mode with the
centrifugal compressor at essentially constant capacity and the positive
displacement compressor at variable capacity; and
at high capacity, the system is run (340) in the first mode with both
compressors providing variable capacity.