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Purdue UniversityPurdue e-Pubs
International Compressor Engineering Conference School of Mechanical Engineering
2018
Geometric Optimization of Scroll CompressorsWieland GelkeIAV GmbH, Germany, [email protected]
Rico BaumgartIAV GmbH, Germany, [email protected]
Joerg AurichIAV GmbH, Germany, [email protected]
Follow this and additional works at: https://docs.lib.purdue.edu/icec
This document has been made available through Purdue e-Pubs, a service of the Purdue University Libraries. Please contact [email protected] foradditional information.Complete proceedings may be acquired in print and on CD-ROM directly from the Ray W. Herrick Laboratories at https://engineering.purdue.edu/Herrick/Events/orderlit.html
Gelke, Wieland; Baumgart, Rico; and Aurich, Joerg, "Geometric Optimization of Scroll Compressors" (2018). InternationalCompressor Engineering Conference. Paper 2609.https://docs.lib.purdue.edu/icec/2609
1435, Page 1
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
Geometric Optimization of Scroll Compressors
Rico BAUMGART1*, Wieland GELKE1, Joerg AURICH1,
Chirra Bhanu Sandeep REDDY1, Andreas MEYER1
1IAV GmbH
Chemnitz, Germany
+49 (0)371 237 34918
* Corresponding Author
ABSTRACT
Electric driven scroll compressors are widely used for the A/C systems in electric vehicles. Besides many other
advantages, scroll compressors offer comparable good efficiencies and can be easily used in both the refrigeration
and heat pump mode.
Scroll compressors were first introduced in the early 20th century. Over the decades they have been further
developed and optimized and became increasingly efficient. Among other things, numerical simulations contributed
to this process. However, most of the studies focus on the optimization of the scroll wraps and their design.
Investigations concerning the design and the positioning of the pre-outlet and main outlet bores are rarely present.
Therefore, this contribution focuses on the optimization of the pre- and main outlet positions of scroll compressors
by means of a geometrical and physical based simulation model. The first part outlines a brief introduction of the
simulation model as well as the applied parameters. Subsequently, different positions of the pre- and main outlet
bores for an exemplary CO2-scroll compressor with a variable wall thickness are investigated. The main focus is on
their influence on the required driving power, the pressure behavior in the chambers and the mass flow rate.
The investigations clarify that with an appropriate positioning of the pre- and main outlet bores the internal over-
compression at low load and partial load operating points can be avoided. Simultaneously, the maximum chamber
pressure at high load operating points is significantly reduced, which leads to a lower oil temperature as well as a
reduction of the required driving power.
1. INTRODUCTION
Currently, the automotive industry is focused on the development of electric vehicles. It is well known, that the
cruising range of these vehicles decreases to half by using the air condition under extreme conditions. This affects
the consumer behavior. Therefore, the development of energy efficient air conditioning systems is one of the major
challenges for the automotive industry.
Electric vehicles are mostly equipped with an electric driven scroll compressor. This component causes by far the
highest power consumption within the air conditioning system. For years, scroll compressors were optimized
considering their efficiencies, durability and manufacturing costs. One of these optimizations is the usage of pre-
outlet bores, which improve the compressor behavior particularly at low load and partial load operating points.
However, publications concerning the positioning of these pre-outlet bores as well as the main outlet bore are rarely
present. Besides cooling the air for the passenger compartment, the air conditioning system in electric vehicles is
increasingly used for further tasks, e.g. cooling the battery. This leads to a higher demand of refrigerating capacity
and consequently to a higher proportion of high load operating points for the scroll compressor.
1435, Page 2
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
That is why IAV investigated the influences of the pre-outlet and main outlet positions on the behavior of an
exemplary CO2-compressor for low load, partial load and especially high load operating points.
For this purpose, a geometrical and physical based simulation model was used, which is briefly introduced in
chapter 2. Chapter 3 describes the operating points as well as the boundary conditions. Subsequently, all simulation
results are explained in chapter 4. The paper closes with a conclusion and an outlook for further investigations.
2. SIMULATION MODEL
The simulation model for the scroll compressor is divided into a geometrical and a thermodynamic part. For the
geometrical modeling of the scroll wraps with variable wall thickness the approach from Liu et. al. (2010) was taken
as a basis, though with modified equations. Thus, by defining the geometric parameters the inner and outer scroll
contours can be calculated (Figure 1). The modelling of the discharge geometry (Figure 1, arc-arc) is based on the
works of Bell (2011) and Bell et al. (2014). Using the contact points between the fixed and the orbiting scroll, the
volume equations of the different scroll chambers as well as their derivatives can be determined. Additionally, the
model calculates the free area of the pre- and main outlet bores and the gap width between the orbiting and the fixed
scroll at the discharge region (Figure 2) depending on the driving angle.
For any chamber, a differential equation for the change of gas mass can be formulated (Aurich, 2018). This equation
is a balance of the inflowing and the outflowing mass flow rate for each chamber and considers the radial and flank
leakage between the scroll wraps (Figure 1) as well. For a mathematical description of the leakage, the approach of
Bell (2011) is used. The modelling of the mass flow rate through the pre-outlets and the main outlet is based on
Böswirth (1994). In addition, the model can also simulate the dynamic valve behavior, which is described in
Baumgart (2010).
With the equations for the chamber volumes, their derivatives, the differential equations for the change of the gas
mass and an equation for the heat flow between the refrigerant and the compressor, a differential equation for the
change of the gas temperature in each chamber can be derived (Aurich, 2018). The properties of the refrigerant, like
the enthalpy and the specific volume, are determined using the software REFPROP (Lemmon et al., 2013).
The compressor simulation starts with assumed refrigerant states in the chambers. The integration of the two above-
mentioned differential equations delivers the gas mass as well as the temperature in the chambers depending on the
driving angle. With these both values, the pressure can be calculated (e.g. using REFPROP). This calculation is
executed for a full revolution of the compressor. Subsequently, the final states (temperature, gas mass, pressure) in
the chambers are compared with the initial states. If the difference is higher than a specified value, a new set of
initial states is defined and the simulation starts again. This iteration loop is repeated until the results are converged.
With these results all other compressor properties like the forces, the efficiencies, the torque and the required driving
power can be determined. The basic structure of the model is shown in Aurich (2018).
This model was validated several times on different typical automotive scroll compressors and the simulation results
showed a good match with the measurement values.
3. SIMULATION PARAMETERS
The following investigations are based on four exemplary operating points defined according to DIN SPEC 74116
(2016). These points listed in Table 1 are representative for typical air conditioning and heat pump refrigerant
circuits in passenger cars using carbon dioxide (CO2) as refrigerant.
Table 1: Operating points
No.
Compressor
Speed
[rpm]
Inlet
Pressure
[bar]
Outlet
Pressure
[bar]
Inlet
Temperature
[°C]
Superheating
[K]
I 1,500 35 50 10
II 6,000 20 125 -15
III 3,000 35 75 25
IV 8,000 35 125 25
1435, Page 3
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
For the simulations an exemplary scroll compressor with a variable wall thickness and an arc-arc discharge
geometry was defined. The geometric parameters are listed in Table 2. The used designation is based on Bell (2011).
The compressor displacement is 6.05 cm³ to meet the specified requirements (DIN SPEC 74116, 2016). The radial
and flank leakage gap is assumed to 15 μm, which is in the value range published e.g. in Bell (2011). Chen, Y. et. al.
(2002) describes an empiric approach to calculate the leakage gap depending on the suction and discharge pressure
(see also Bell, 2011, p. 19). However, this approach is only valid for a small range of pressure ratios and was there-
fore not used here. All simulations were executed with a drive efficiency of 100 % in order to exclude this influen-
cing factor.
Table 2: Geometric parameters of the scroll wrap
Parameter Value Parameter Value
Initial base circle diameter 4.4 mm Start angle outer involute 50°
Initial scroll thickness 4.0 mm End angle inner involute 910°
Thickness increment -0.04 mm/rad Scroll height 7 mm
Initial angle inner involute 47° Radial/Flank leakage gap width 15 μm
Start angle inner involute 215° Displacement 6.05 cm3
The investigation described afterwards includes a detailed analysis of four different pre-outlet positions and three
main outlet positions. Figure 1 illustrates the outlet positions as well as two combinations of them.
The pre-outlet positions were defined in a way to open (Figure 1, center) or close (Figure 1, right) simultaneously at
a defined driving angle φ depending on the current objective. The main outlet was moved in three steps along the
inner contour of the fixed scroll. For any outlet position, the free area of the pre- and main outlets was calculated
depending on the driving angle. This is the free area for the refrigerant to flow to the compressor outlet, limited by
the orbiting scroll as well as the respective outlet bore. In addition, for the pre-outlets it was detected in which
chamber they are located.
Based on Bell (2011) the different chambers are designated as follows (Figure 1):
• s: suction chamber
• c: compression chamber
• ddd: discharge chamber (can be separated in two chambers d and dd, cf. Figure 2)
The compressor is symmetric, which means the chamber pairs (s, c, d) have the same volume at any angle.
Figure 1: Investigated positions of the pre-outlets and the main outlet
main outlet:
a b c dpre-outlet:pre-outlet: a
main outlet: A
j = 180 j = 77
A B Cpre-outlet: b
main outlet: A
orbiting scroll
fixed scroll
s
c
c
ddd
s
ddd
s
s
c
c
inner contour outer contour
discharge geometry
(arc-arc)
flank leakage
radial leakage
1435, Page 4
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
4. RESULTS
Figure 2 illustrates the results for the pre-outlet position a and the main outlet position A. The simulation is based on
the partial load operating point I (see Table 1). Additionally, the results for a compressor without pre-outlets are
shown for comparison purposes.
Figure 2: Simulation results for operating point I (pre-outlet position a and main outlet position A)
At a driving angle of φ = 95° the orbiting scroll covers both pre-outlets completely and the main outlet for a large
part. The refrigerant can leave the discharge chamber ddd only through a slight gap between the orbiting scroll and
the main outlet bore. Consequently, the mass flow rate is very low and the pressure in the discharge chamber ddd
increases significantly at this angle.
At the compressor without pre-outlets the pressure in the compression chamber c exceeds the outlet pressure
(50 bar) at a driving angle of φ = 180° (Figure 2, bottom left diagram). In the following, this effect is named internal
over-compression. The over-compression leads to an increase of the gas temperature as well as an increase of the
required driving power. This effect can be avoided with pre-outlets (Figure 2, right side). As Figure 1 shows, both
0
20
40
60
80
100
0 60 120 180 240 300 360Mass f
low
rate
[k
g/h
]
0
20
40
60
80
100
0 60 120 180 240 300 360
Driving angle [ ]Driving angle [ ]
Pressu
re [bar]
Pressu
re [bar]
Without pre-outlets With pre-outlets
j = 95 j = 240 j = 340
95 180 240 340 95 180 240 340
cs d
s
s
c
c
ddd
ddd
ddd
s
s
c
c
ss
dd
d d
dd
jm = 49.6
jd
pre-outletsmain outlet
Mass f
low
rate
[k
g/h
]
jd
Legend:
Operating point: I
jm = 50
Pre-outlet position: a Main outlet position: A
inner contactorbiting scroll
fixed scroll
dd
d
gap
flow
30
40
50
60
70
0 60 120 180 240 300 360
m 31.91 kg h
30
40
50
60
70
0 60 120 180 240 300 360
55.7 bar
m 33.95 kg h
1435, Page 5
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
pre-outlets are uncovered simultaneously at a driving angle of φ = 180°. Thus, the refrigerant can leave the
compression chamber c through the pre-outlets, which reduces the over-compression significantly.
With the defined geometry, the inner contact between the orbiting and the fixed scroll only exists until a driving
angle of φ = 320°. This angle is called as a discharge angle φd (Bell, 2011). From this angle, between both scrolls a
gap arises connecting the compression chamber c and the discharge chamber ddd (Figure 2, φ = 340°). Thus, a
pressure equilibration process starts between both chambers. According to Bell (2011), the chambers are redefined
after the discharge angle φd as follows:
• chamber c → chamber d (chamber c does not longer exist until φ = 360 )
• chamber ddd → chamber dd (chamber ddd does not longer exist)
The simulation model calculates the gap width between the scrolls depending on the driving angle. Due to the
pressure difference between the chambers, the refrigerant flows from chamber d to chamber dd and afterwards
through the main outlet to the compressor outlet. As a result, the pressure in chamber dd decreases and the pressure
in chamber d increases (Figure 2, bottom left). With an increasing driving angle, the pressures in both chambers are
equalized. However, with the shown operating point the equilibration process is not finished at φ = 360° and is
hence continued at the beginning of the next revolution. At φm ≈ 50° the pressure equilibration process is finally
finished. Starting at this angle, the chambers d and dd are combined to the chamber ddd.
The required driving power of the compressor without pre-outlets is 0.244 kW. With pre-outlets, the driving power
can be reduced to 0.212 kW, which means a saving of 13 %. The refrigerating mass flow rate remains nearly
constant between both cases.
Figure 3 shows the simulation results for the operating points II and IV. The positions of the pre-outlets and the
main outlet are the same as in Figure 2. In operating point II there is no mass flow rate through the pre-outlets
because the pressure in the compression chamber c does not exceed the outlet pressure (125 bar, cf. Table 1).
However, at φ = 95° the pressure in the chamber ddd exceeds the outlet pressure and the discharge process begins.
Since the pre-outlets are completely covered while the main outlet becomes uncovered, there is only a mass flow
through the main outlet bore.
Figure 3: Simulation results for operating points II and IV (pre-outlet position a and main outlet position A)
0
200
400
600
0 60 120 180 240 300 360
0
100
200
300
0 60 120 180 240 300 360
Mass f
low
rate
[k
g/h
]
Driving angle [ ]Driving angle [ ]
Pressu
re [bar]
Pressu
re [bar]
Operating point II Operating point IV
95 95
cs d ddddd pre-outletsmain outlet
Mass f
low
rate
[k
g/h
]
Legend:
Pre-outlet position: a Main outlet position: A
jd jd24
m 103.99 kg h
0
30
60
90
120
150
0 60 120 180 240 300 360
0
50
100
150
200
0 60 120 180 240 300 360
182.0 bar
m 215.07 kg h
1435, Page 6
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
With operating point IV, the pressure in the compression chamber c is always lower than the outlet pressure as well
(125 bar, cf. Table 1). However, at a driving angle of φ = 24° the pressure in the chamber ddd exceeds the outlet
pressure. Since the pre-outlets are located in this chamber, there is a mass flow from the chamber ddd through the
pre-outlets and through the main outlet. At φ = 95° the pre-outlets are completely covered and only a narrow gap
between the main outlet and the orbiting scroll is uncovered (cf. Figure 2). As a result, the discharge of the
refrigerant is throttled and the pressure in the chamber ddd increases. The maximum pressure of 182 bar can cause a
very high thermal load for the oil and leads to a high driving power. With driving angles φ > 95° the uncovered area
of the main outlet and thus the mass flow rate increases which results in a pressure decrease at φ > 120° in chamber
ddd. At φ = φd the pressure equilibration process between the chambers d and dd starts. Compared to operating point
I (Figure 2, bottom left), the pressure difference is reverse (pdd > pd) and consequently a small amount of the
refrigerant flows back from chamber dd to chamber d.
In Figure 4 the results for the operating points I and IV are illustrated. The main outlet position is the same as in the
previous figures. However, the pre-outlets are rotated clockwise to position b. In both operating points, at φ = 77°
the pressure in chamber ddd is higher than the respective outlet pressure (I = 50 bar and IV = 125 bar). At this
driving angle, the orbiting scroll covers the main outlet. However, the pre-outlets are still uncovered and located in
chamber ddd. Consequently, the maximum pressure in chamber ddd can be reduced from 55.7 bar (Figure 2, right
side) to 50.6 bar in operation point I and from 182 bar (Figure 3, right side) to 166 bar in operation point IV,
respectively. This pressure reduction also causes a lower thermal load for the oil.
Figure 4: Simulation results for operating points I and IV (pre-outlet position b and main outlet position A)
0
200
400
600
0 60 120 180 240 300 360
0
10
20
30
40
0 60 120 180 240 300 360 Mass f
low
rate
[k
g/h
]
Driving angle [ ]Driving angle [ ]
Pressu
re [bar]
Pressu
re [bar]
Operating point I Operating point IV
77 77
cs d ddddd pre-outletsmain outlet
Mass f
low
rate
[k
g/h
]
Legend:
j = 77 j = 160 j = 240
s
s
c
c
ddd ddd
s
s
c
c
s
s
jm jm 240 160 240 160 jd jd
ddd
Pre-outlet position: b Main outlet position: A
c
c
orbiting scroll
fixed scroll
30
35
40
45
50
55
0 60 120 180 240 300 360
50.6 bar
m 33.93 kg h
0
50
100
150
200
0 60 120 180 240 300 360
m 215.04 kg h
166.0 bar
1435, Page 7
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
With increasing driving angles φ > 77°, the orbiting scroll more and more covers the pre-outlets and the main outlet
becomes uncovered. At φ = 160° the refrigerant mass flow rate through the main outlet is near the maximum and the
mass flow through the pre-outlets is zero.
With operation point I the over-compression effect (pc > pout) starts at φ = 180°. At this angle the pre-outlets are
uncovered again and located in the chamber c. Figure 4 shows also the situation for a driving angle of φ = 240°.
Compared to pre-outlet position a (Figure 2, right side), the over-compression effect can also be avoided effectively
with the pre-outlet position b. However, the pressure in chamber c is slightly higher compared to the position a.
The change of the pre-outlet position from a to b reduces the maximum pressure and prevents the internal over-
compression effect. This results in a reduction of the required driving power at a constant mass flow rate. The
achievable power savings are listed in table 3. At operating point IV, the pre-outlet position b saves 161 W.
Table 3: Power savings by changing the pre-outlet position from a to b (main outlet position A)
Operating point
Driving Power
pre-outlet position a
main outlet position A
[kW]
Driving Power
pre-outlet position b
main outlet position A
[kW]
Power saving
[%]
I 0.212 kW 0.209 kW 1.4 %
IV 5.247 kW 5.086 kW 3.07 % (161 W)
Figure 5 illustrates the results for the same operating points as in Figure 4 (I and IV). The pre-outlet positions were
rotated clockwise to position c and the main outlet position was changed to position C.
At operating point I, the pressure in the ddd chamber reaches its maximum at φ = 60° because the main outlet is
mostly covered. The refrigerant can leave the chamber ddd only through the pre-outlets and through a small gap at
the main outlet. At φ = 180° the internal over-compression begins. Due to the clockwise rotation of the pre-outlets
the opening process of these outlets starts a few degrees later compared to the pre-outlet positions a and b (cf.
Figure 2 and Figure 4). This results in a higher over-compression in chamber c.
With operating point IV, the pressure in the chamber ddd (144.8 bar) is significantly lower than at all previous
investigated outlet positions. The change of the main outlet position from A to C causes a reduced coverage by the
orbiting scroll and thus a higher flow area. In comparison to the results in Figure 3, the maximum pressure can be
reduced by 37.2 bar and compared with Figure 4 the pressure reduction is 21.2 bar. This results in a power saving of
8.52 % and 447 W, respectively (Table 4).
Table 4: Power savings by changing the pre-outlet positions from a to c and main outlet position from A to C
Operating point
Driving Power
pre-outlet position a
main outlet position A
[kW]
Driving Power
pre-outlet position c
main outlet position C
[kW]
Power saving
[%]
I 0.212 kW 0.209 kW 1.4 %
IV 5.247 kW 4.800 kW 8.52 % (447 W)
1435, Page 8
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
Figure 5: Simulation results for pre-outlet position c and main outlet position C (operating points I and IV)
Figure 6 summarizes the required compressor driving power for the investigated pre-outlet and main outlet
positions. Additionally, the driving power for a compressor without pre-outlets is shown for each operating point. As
already mentioned, the drive efficiency was assumed to 100 % to exclude any drive effect on the compressor
efficiencies. The figure makes clear, that pre-outlets significantly reduce the driving power at all investigated
operating points. Only at operating point II with the pre-outlet position a and the main outlet position A there is no
advantage compared with the compressor without pre-outlets. In this case, there is no mass flow through the pre-
outlets (cf. Figure 3).
Changing the main outlet position (A→B→C) along the inner contour of the fixed scroll leads to a significant
reduction of the driving power with all operating points. The less covering of the main outlet bore by the orbiting
scroll causes this. Hereby, the step from A to B shows a higher power reduction than the step from B to C.
Rotating the pre-outlet bores clockwise lead to a further remarkable reduction of the required driving power at least
at the high load operating points II and IV. At partial load operating point I with the main outlet position A, only a
change from pre-outlet position a to b results in a power reduction. A change from b to c and from c to d leads to a
slight increase in driving power. This can be explained with the delayed opening process of the pre-outlets in
0
200
400
600
0 60 120 180 240 300 360
0
10
20
30
40
0 60 120 180 240 300 360 Mass f
low
rate
[k
g/h
]
Driving angle [ ]Driving angle [ ]
Pressu
re [bar]
Pressu
re [bar]
Operating point I Operating point IV
80 80
cs d ddddd pre-outletsmain outlet
Mass f
low
rate
[k
g/h
]
Legend:
j = 60 j = 80 j = 180
s
c
c
ddd ddd
s
s
c
c
s
s
jm jm 180 180 jd jd
ddd
Pre-outlet position: c Main outlet position: C
c
c
60
s
60
m 33.55 kg h
orbiting scroll
fixed scroll
30
35
40
45
50
55
0 60 120 180 240 300 360
50.8 bar
0
50
100
150
200
0 60 120 180 240 300 360
144.8 bar
m 215.05 kg h
1435, Page 9
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
compression chamber c and the resulting internal over-compression effect, which can be avoided most effectively
with the pre-outlet position a. With the operating point III there is neither a significant power reduction by changing
the pre-outlet position from b to c nor from c to d.
The highest savings of required power can be achieved at operating point IV. With the main outlet position A,
0.27 kW (5.1 %) can be saved by changing the pre-outlet position from a to d. This effect can be explained as
follows:
At driving angles between φ = 70° and φ = 130° the orbiting scroll covers the main outlet bore as well as the pre-
outlets (cf. Figure 2, φ = 95°). Thus, the refrigerant can leave the discharge chamber ddd only through a small gap at
the main outlet bore. This leads to a pressure increase in chamber ddd and as a result to a higher driving power. With
the pre-outlet position d, the flow area can be increased while the main outlet bore (A) is covered. This reduces the
pressure in chamber ddd and consequently the required driving power. Additionally, the reduced pressure causes a
lower thermal load for the oil. A further increase in this effect can be achieved with a change of the main outlet
position from A to C. This reduces the covering of the bore and enables a power saving of 0.46 kW, which is equal
to a reduction of 8.8 %.
A further remarkable fact is that all investigated pre- and main outlet positions showed only a very small influence
on the refrigerant mass flow rate.
Figure 6: Required compressor driving power for different pre- and main outlet positions
0,2
0,21
0,22
0,23
0,24
0,25
3,25
3,26
3,27
3,28
3,29
3,3
3,31
0,94
0,96
0,98
1
1,02
1,04
1,06
4,7
4,8
4,9
5
5,1
5,2
5,3
5,4
Driv
ing
pow
er
[kW
]Operating point I
CA B
0.20a
Operating point II
Pre-outlet position
b c da
Pre-outlet position
b c d
Driv
ing
pow
er
[kW
]
Operating point III Operating point IV
0.21
0.22
0.23
0.24
0.25
3.25
3.26
3.27
3.28
3.29
3.30
3.31
0.96
0.94
0.98
1.00
1.02
1.04
1.06
4.7
4.8
4.9
5.0
5.1
5.2
5.3
5.4
a
Pre-outlet position
b c da
Pre-outlet position
b c d
main outlet position:
max min
m 33.61 kg h
m m2.41%
m
max min
m 104.0 kg h
m m0.02 %
m
max min
m 70.52 kg h
m m0.05 %
m
max min
m 215.05 kg h
m m0.02 %
m
A (without pre-outlets)
Driv
ing
pow
er
[kW
]
Driv
ing
pow
er
[kW
]
0.2
7 kW
0.4
6 kW
1435, Page 10
24th International Compressor Engineering Conference at Purdue, July 9-12, 2018
4. CONCLUSION AND OUTLOOK
Using scroll compressors with pre-outlets, the internal over-compression can be avoided effectively at low load and
partial load operating points. Among others, this results in a reduced driving power requirement. Due to the
increasing refrigerating capacity demand of electric vehicles, e.g. for the battery cooling, the compressor
optimization should not only focus on the low load and partial load operating points.
For that reason, different positions for the pre-outlet and the main outlet bores were investigated regarding their
influence on the properties of an exemplary CO2-compressor, especially with high load operating conditions.
The investigations have clearly shown, that an appropriate positioning of the pre- and main outlet bores reduces the
maximum pressure in the discharge chamber at the high load operating points. This corresponds with a significant
reduction of the required driving power and a lower thermal load for the oil. However, at low load operating points
the clockwise rotation of the pre-outlets leads to a small increase in driving power. The repositioning of the pre- and
main outlet bores showed no significant influence on the mass flow rate.
In a future work these effects will be further analyzed and optimized. Besides other geometric optimizations, also
noncircular outlet bores should be investigated.
NOMENCLATURE
φ driving angle (°) ṁ mass flow rate (kg/h)
Subscript
c compression max maximum
d discharge min minimum
m merged s suction
REFERENCES
Liu, Y., Hung, C., Chang, Y., 2010: Study on involute of circle with variable radii in a scroll compressor,
Mechanism and Machine Theory 45, Nr. 11, 1520-1536.
Bell, I., 2011: Theoretical and Experimental Analysis of Liquid Flooded Compression in Scroll Compressors, Diss.
Purdue University, West Lafayette, Indiana, USA, 622 p.
Bell, I.H., Groll, E.A., Braun, J.E., Horton W.T., Lemort, V., 2014: Comprehensive analytic solutions for the
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