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A Novel Defrosting Method in Gasoline Vapor Recovery Application
Liang, Jierong; Sun, Li; Li, Tingxun
Published in:Energy
Link to article, DOI:10.1016/j.energy.2018.08.172
Publication date:2018
Document VersionPeer reviewed version
Link back to DTU Orbit
Citation (APA):Liang, J., Sun, L., & Li, T. (2018). A Novel Defrosting Method in Gasoline Vapor Recovery Application. Energy,163, 751-765. https://doi.org/10.1016/j.energy.2018.08.172
Accepted Manuscript
A Novel Defrosting Method in Gasoline Vapor Recovery Application
Jierong Liang, Li Sun, Tingxun Li
PII: S0360-5442(18)31708-0
DOI: 10.1016/j.energy.2018.08.172
Reference: EGY 13648
To appear in: Energy
Received Date: 08 February 2018
Accepted Date: 22 August 2018
Please cite this article as: Jierong Liang, Li Sun, Tingxun Li, A Novel Defrosting Method in Gasoline Vapor Recovery Application, (2018), doi: 10.1016/j.energy.2018.08.172Energy
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A Novel Defrosting Method in Gasoline Vapor Recovery ApplicationJierong Lianga,b , Li Sunc, Tingxun Lia,* ,
a School of Engineering, Sun Yat-sen University, West XINGANG Road 135,Guangzhou, China, 510275b Guangdong Shenling Environmental System Co., Ltd., XINGLONG 10th Road 8, Foshan, China, 528313c Department of Chemical and Biochemical Engineering, Technical University of Denmark, 2800 Kongens Lyngby, Denmark
*Corresponding author. Tel.: +86 020 3933 1115.E-mail address: Litx@hotmail.com (Tingxun Li), Litx@mail.sysu.edu.cn.Postal address: No. 135, Xingang West Rd., Guangzhou City 510725, Guangdong PR China.
Abstract: Condensation method is comprehensively applied for gasoline vapor recovery (GVR), of which frosts in
the heat exchanger is the greatest challenge, especially for the continuous long running cases. A novel dual channel
GVR cascade refrigeration system with shell-tube heat exchanger was presented and tested in this paper. With one-
work-one-standby evaporator settings, combined with refrigerant evacuation and delay switching strategies, the
defrosting of low temperature shell-tube heat exchanger was analyzed and solved. Also multi-stage cycle was
introduced to supply three cooling stage, which cooled the gasoline vapor from ordinary temperature to about -
70°C. By the means of industrial application validation and process calculation, the ability of the non-stop cooling
during defrosting was verified. The refrigerant evacuation was proposed to prevent high pressure drop caused by
frost accumulation, which also improved the cooling capacity by 28.2% and approached the defrost efficiency of
55.4%. In addition, it was found that delay switching can effectively reduce the capacity fluctuation. Based on
sensitivity studies, 20 minutes delay was identified as the best switching timing for this device. The capacity of this
system performed lower reduction, higher duty ratio and defrost efficiency.
Keywords:Gasoline vapor, Dual channel, Defrost, Cascade cycle, Shell-tube
1. Introduction
Ministry of Environmental Protection of China had stepped up efforts to control gaseous non-
methane hydrocarbons (NMHC) emitted from petrochemical industry, chemical terminals, and oil
depots. Gasoline vapor containing volatile organic compounds (VOCs) was rooted in loading,
unloading and storage operations (Fig.1), which not only resulted in air pollution and potential fire
risks, but also brought about huge economic losses [1]. Many of separation methods have been
developed for GVR, which include adsorption [2-3], absorption [4], condensation [5], and membrane
[6]. Each of these methods was used in a certain range of flow rate and NMHC concentrations for
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economical consideration. The high concentration gasoline vapor was normally recovered by the
condensation separation method for the advantage of convenient installation and direct resources
recovery [5].
Fig.1 Schematic of VOCs emitting and its recovery
Currently low temperature gas cooling process mainly focused on liquefied natural gas (LNG, [7])
and cryogenic air separation process [8]. A limited number of studies have provided insight into the
special GVR cooling process. Shie et al. [9] discussed the relationship of recovery efficiency and
refrigeration temperature and concluded that treated gas should be refrigerated to about -73°C to meet
the vapor recovery efficiency of 90%. Shi and Huang [5] presented a three-stage condensation system
for industrial GVR, and compared the total cooling duties between single-stage and three-stage
cooling processes. These studies prepared the ground for cooling temperature and stage settings of
GVR. But the detailed understanding of the corresponding refrigeration cycle performance efficiency
and stability is important for GVR applications. A three-stage refrigeration cycle for pure propane
liquefaction was analyzed by Kalantar-Neyestanaki et al. [10], but not related to defrosting issues. To
the best of our knowledge, no publication presented experimental or industrial investigation on the
cooling cycle behaviors and especially defrosting control strategy for GVR systems.
A three-stage cooling process of GVR (Fig.2) was designed as pre-cooling (1st stage), middle-
cooling (2nd stage) and super-cooling (3rd stage) with multi-temperature cycles. Cascade cycle was
adopted to meet the cooling requirement of lower than -70°C, which was classified as classic and
auto cascade. A classical cascade included two or more separate cycles with different refrigerant,
while auto-cascade system was driven by a single compressor, using zeotropic mixture as refrigerant.
Since auto-cascade was more sensitive against the classical cascade system [11], classical cascade
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system [12-13] was chosen in this work. Shell-tube heat exchanger was chosen as the evaporator for
the highly required safety in explosive environment.
Fig.2 Temperature profile of cooling process
In contrast to LNG process, GVR cooling process has no pre-process such as dehydration [14-15],
the vapor might contain multiple heavy ingredients such as benzene and water. These heavy
components might freeze or frost during cooling process. According to demand of long continuous
time working in shipping or tank farm GVR application [1], there is no interruption opportunity to
remove the frost regularly. Frosting was the greatest obstacle which seriously penalized the device
efficiency and functionality [16]. Therefore, defrosting method is essential for uninterrupted cooling
of GVR. All defrosting methods have the common disadvantage of reducing the overall cooling
efficiency because the energy is used for defrosting but not cooling. For the purpose of uninterrupted
cooling, Jang et al. [17] designed a non-stop defrosting cycle with dual hot gas spray to dual outdoor
heat exchanger. But different from normal heat pump systems that adopt commonly tube-fin heat
exchangers, the GVR cooling system needs to solve the defrosting problems in shell-tube heat
exchanger. In addition, GVR process might have much shorter frosting period and more frequently
defrosting intervals due to its lower cooling temperature. A novel dual, parallel channel defrosting
method was proposed and researched in this study. Distinct from conventional multi-evaporator
system relating to variable refrigerant flow (VRF) systems [18-20], this method designed the
switchable one-work-one-standby evaporator mechanism.
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The defrosting techniques used in the HVAC&R industry may be adopted in GVR cooling system.
Amer and Wang [21] and Hu et al. [22] summarized defrosting methods as follows: (1) on-off cycling
[23], (2) hot gas bypass [24-26], (3) reverse cycle [27-28], (4) electric resistive heating [29-30], (5)
desiccants as dehumidifiers or (6) ultrasonic wave [31-32] and their combined methods [33-35]. Hot
gas bypass (HGB) was employed here as the key candidate defrosting method instead of reverse cycle
for simple structure [36]. Comparing with the defrosting of microchannel heat exchangers [37-38]
and fin-tube exchangers [39], the defrosting in low temperature shell-tube heat exchanger shell side
was seldom mentioned in previous studies, and was more complex (Fig.3). In conventional HGB
mechanism, the residual refrigerant inside evaporator was removed by evaporation during defrosting
[17, 30], which consumed the energy of hot gas [23]. Especially in low evaporating temperature
conditions, this additional energy consumption will become more and even excessive. Thus, a new
defrosting energy recovery strategy was provided, which contained evacuating the refrigerant inside
the defrosting evaporators and the channel switching delay. Most cooling potential of the frost and
cold refrigerant have been recovered. From the existing study [40-42], the defrosting process could
be divided into: (ⅰ) defrosting initiation (pre-heating) stage, (ⅱ) melting and draining (vaporizing)
stage and (ⅲ) defrost termination and recovery (dry heating) stage. In this defrosting strategy, on-off
cycling and refrigerant evacuation methods were set in motion for stage (ⅰ). And different HGB
mode and switching delay was adopted for stage (ii) and (iii). Also the electric distributed heater with
conductive heat transfer was frequently in use for heating the shell and preventing re-frosting at the
internal wall of the shell.
Fig.3 Schematic of shell-tube heat exchanger defrosting
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Table 1 - Technologies summation in this system
Aspects References Descriptions Differentiations in this system
Multi-evaporator Cho et al. [33]
Arpagaus et al. [42]
Evaporators for parallel use or
different modes
Multiple temperature levels
Switchable one-work-one-standby
evaporator for continuous cooling
Defrost Amer and Wang [21]
Song et al. [36]
Jang et al. [17]
Defrosting methods for
microchannel and fin-tube HX
Combined defrosting method for
shell-tube heat exchanger
Evacuating refrigerant before
defrosting
Channel switching delay strategy
Multi-stage
cooling and
cascades
Granwehr and Bertsch [44]
Mathison et al. [45]
Shen et al. [46]
Two-stage cycle with closed
economizer and second heat source
Cascade cycle with dual mode
Adding dual, parallel evaporators for
each stage
Capacity control Qureshi and Tassou [47]
Chen et al. [48]
Wang et al. [49]
Xu et al. [50]
Dutta et al. [51]
HGB control
On/off control
Slider control
Liquid injection
Slider for coarse modulation
Dual HGB for fine modulation
Liquid injection and on/off control
for extreme conditions
Oil cooling Lin and Hrnjak, [52]
He et al. [53]
Abu-Mulaweh, [54]
Two-phase thermosiphon loop for
refrigerator and gas turbine
Refrigerant thermosiphon loop
between liquid line and oil cooler
In this paper, a novel dual channel defrosting method for GVR was firstly studied to meet the long
continues time working requirements, especially in shipping and tank farm GVR application. The
cooling system consisted of switchable one-work-one-standby evaporators, multi-stage cooling and
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cascade cycle. The defrosting method was efficient for shell-tube heat exchanger and firstly adopted
the refrigerant evacuation and switching delay. Other relevant technologies of this cooling system
such as capacity control were summarized in Table 1. The ability of continuous cooling was validated
by the industrial application. The method is also applicable to the cases of non-stop long-running gas
cooling process, which contains heavy components.
NomenclatureA1 regression coefficients hx heat exchangerh specific enthalpy, (kJ·kg-1) i initiateK Loss coefficient, , 2∆P/ρv2 in inletL liquid level, (m) j stagem mass, (kg·s-1) k componentm mass rate, (kg) min minimumP pressure, (Pa); Power, (kw) out outletQ cooling capacity, (kw) period operation periodQ capacity rate, (kw·s-1) ref refrigerantRe Reynolds number s solidT temperature, (K) sf solid to fluid statet time, (s) steady steady state𝑉 volumetric flow rate, (m3·s-1) AbbreviationGreek CDR Capacity duty ratioρ density, (kg· m-3) DSP Delay switching phaseΔ difference, (-) GVR Gasoline vapor recoverySubscripts HGB Hot gas bypass0 initial state HGBP Hot gas bypass phasedis discharge HTC High-temperature cycledry dry condition IHX Internal heat exchangere end LNG Liquefied natural gasel electric LTC Low-temperature cycleevap evaporator NMHC gaseous non-methane hydrocarbonsf frost REP Refrigerant evacuating phasegl gasoline liquid TCHGB Temperature control HGBgv gasoline vapor VRF Variable refrigerant flowhot hot gas
2. Process description
As shown in Fig.4, the whole system can be divided into the follow processes:
Gasoline vapor recovery: The gasoline vapor of ordinary temperature passed through the
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recuperator (REC1) for cooling down some degrees and recovering cooling energy of the outlet
gasoline vapor (path 001-002). Then, the gasoline vapor followed three stage cooling with
temperature lower than -70°C, each cooling stage consisted of dual switchable one-work-one-standby
evaporators (path 013-014-015-016 for channel A and path 023-024-025-026 for channel B). After
flowed out the recuperator, its temperature came to 10~20°C and emitted (path 007-008).
Refrigeration cycle: two classical cycles were employed with different refrigerants. As a
preliminary system, the criteria of refrigerant selection was first the safety, nonflammable refrigerant
was preferred. And combined with the operation temperature, R404A was the refrigerant in HTC,
while R23 acted as the low temperature cycle (LTC) refrigerant. The three groups of evaporators
(Evap.1A~Evap.3A, Evap.1B~Evap.3B) operated at different evaporating pressures matching with
the three cooling stages. Two separate heat exchangers in HTC was designed as the economizer,
where the first one was for economizing function (ECO1), and the second one was for pre-cooling of
gasoline vapor (path 115-116A-117A-118 for channel A and path 115-116B-117B-118 for channel
B). Two-stage screw compressor with port (COMP1 and COMP2) for refrigerant injection was
chosen in this study, which was manufactured by Fusheng Corporation and explosion-proof
certificated. An internal heat exchanger (IHX) was installed in LTC to extend operation range and
improve energy efficiency. To regulate the intermediate heat exchanger pressure of HTC during
refrigerant evacuating phase (refer to section 2.2), the automatic control valve (ACK1) was installed.
Drainage accumulation: gasoline condensate from all evaporators and recuperator is draining to
the gasoline storage tank (V5), and condensing heat was recovered by liquid line (path 104-105).
Oil cooling loop: an external oil separator was arranged for each compressor to remove oil from
the high pressure discharged refrigerant gas. The separated oil was cooled in the thermosiphon oil-
cooler and then returned into the compressor (path 301-302 and path 303-304).
Other bypasses: they included hot gas bypass (red lines) and liquid injection (carmine lines). The
HGB provoked through both working and standby evaporator respectively in terms of defrosting and
capacity control, which constituted the dual HGB mechanism and discussed in section 2.2. Liquid
injection was added in each cycle (path 401, 402) to prevent excessively high discharge temperature
when operating at extreme high compression ratios.
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Fig.4 Schematic of cooling system
2.1 Refrigerant circuits
P-H plots for the refrigerant circuits were depicted in Fig. 5. In the diagrams corresponding to the
position numbers in Fig.4, phases 101-102-103-104 indicated that the HTC refrigerant was cooled by
the airflow and gasoline condensate and heated by the oil circuits. Phases 104-115-116-117
accomplished the economizing and pre-cooling functions, while phases 125-126-127 indicated the
middle-cooling process. The cascade process within the intermediate heat exchanger was approaching
the middle-cooling phases, which was however not depicted in the diagrams. For LTC in normal
refrigerating operations, phases 202-203 and 205-206 represented the heat transferring in IHX. Phases
204-205 depicted the super-cooling process.
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(1) (2)
Fig.5 P-H diagrams of (1) HTC and (2) LTC
2.2 Channel switching and defrosting methodology
The defrosting control methodology was illustrated in Fig.6. The two evaporators of each cooling
stage were operating in one-work-one-standby procedure. The existing techniques to initiate and
terminate defrost cycles are focused on time-temperature controlled, pressure differential,
temperature difference [55] and heat transfer balance [56]. In this device, pressure drop of gasoline
vapor was observed more notable response than temperature difference. The frosting condition was
monitored by the differential pressure transmitter and flowmeter. The criterion for switching was that
the pressure drop was twice of one in dry conditions ( ). adopted the loss coefficient ∆𝑃𝑑𝑟𝑦 ∆𝑃𝑑𝑟𝑦
equation of Reuter and Anderson [57] (Eq.[1]), which was suitable for bare tube bundle in dry mode.
The calculation of was fitted by the tested airflow in this device according to Eq.[2]. ∆𝑃𝑑𝑟𝑦
[1] [2]𝐾 = 14.5049𝑅𝑒 ‒ 0.04678 ∆𝑃𝑑𝑟𝑦 = 𝐴1𝑉1.95322
Where K was Loss coefficient, Re was Reynolds number, A1 was regression coefficient and was 𝑉
volumetric flow rate.
The defrost process is initiated when the pressure drop has reached a calculated value correspond to the
current flow rate (i.e. 2× ), and terminated when the evaporator temperature has reached a predetermined ∆𝑃𝑑𝑟𝑦
value. The procedure was divided into the following phases:
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Fig.6 Switching control conditions
Refrigerant evacuating phase (REP): this phase was to remove the cold liquid refrigerant [58]
from the defrosting evaporator before the hot gas supply was initiated.
Step1.1: closed all the liquid line solenoid valves (SV1-SV6), the refrigerant was evacuated to the
reservoir.
Step1.2: control valve (ACK1) was regulated for minor flow to enhance the vacuum of the suction
line for refrigerant evacuation.
Step1.3: the gasoline vapor still flowed through the defrosting channel for a period as a manner of
on-off cycling defrosting.
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Step1.4: the outlet of Evap.1A and Evap.1B was reconnected to the suction line by electrical tee
valve (TV1) for more thorough evacuation of refrigerant.
Delay switching phase (DSP): this phase distributed the gasoline vapor between the dual channels
for balancing the cooling and defrosting effect.
Step2.1: opened the liquid line solenoid valves (SV1/3/5 or SV2/4/6) in the cooling channel. The
refrigerant was migrated to the working evaporators, which were prepared to be placed back into low
temperature service.
Step2.2: constant frequency hot gas bypass (CFHGB) mechanism was invoked in defrosting
evaporators to enhance the defrosting effect. The definition of CFHGB was described that HGB
operated with fixed period and pulse width. As a lot of heat was needed for defrosting in this period,
HGB spraying for too long would result potential risks to the system performance and the compressor
safety.
Step2.3: opened all the channel valves (MOV1-MOV4); both working and standby evaporators
shared responsibility for the gasoline vapor cooling.
Hot gas bypass phase (HGBP): the defrosting was dominated by HGB in this phase.
Step3.1: closed the defrosting channel valves (MOV1/3 or MOV2/4). The gasoline vapor was
switched to the cooling channel completely.
Step3.2: the electrical heat-tracing system in defrosting evaporators started to heat the shell.
Step3.3: the HGB mechanism was changed from CFHGB to temperature control hot gas bypass
(TCHGB). TCHGB mode reduced the HGB frequency for energy saving, the PLC judged whether
opened or stopped the HGB by the evaporator temperature signal.
3. Results and discussions
The industrial application for the system proposed by this study was shown in Fig.7, which is located
at Port of Lianyungang, China. Since there were no test standards or regulations for this type system
of until now, the test condition was set as that the inlet gasoline vapor temperature was 25°C. The
system performance was tested as the following four cases for 72 hours.
(1) Case #1: twenty minutes delay switching. Refer to section 2.2, the switching delay was set to
20 minutes. The timing was approximately that the frosts at the defrosting evaporator of 1st
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stage have just been melted.
(2) Case #2: forty five minutes delay switching. The procedure was the same as case #1, except
that the delay was set to 45 minutes. The timing was approximately that frosts at the defrosting
evaporator of 2nd stage have just been melted.
(3) Case #3: no refrigerant evacuation. Refer to section 2.2, the process went through only DSP
and HGBP.
(4) Case #4: no delay switching. Refer to section 2.2, the process went through only REP and
HGBP.
Fig.7 100~600 m3/h gasoline vapor cooling device for ship loading, serving 4 marine loading arms
3.1 Operation characteristics
Stable pressure drop of the channel was an essential prerequisite of a successful continuous running,
because the ship tanker should work in the design pressure range. Fig.8 compared the pressure drop
of the gasoline vapor channel between case #1 and case #3. For case #3, the pressure difference
increased continuously and could not return to normal range. The main reason was inefficient
defrosting, which due to hot gas energy self-consumption by cold residual refrigerant in the defrosting
evaporator. Therefore, the frosts accumulated all the time and resulted in the high pressure difference
and operation failure after about 24 hours running. For case #1, despite the maximum and minimum
values of the pressure difference in each period were not the same completely because of unstable
working conditions, the value of the pressure difference was still able to return to normal range. This
device was believed to be suitable for the long time non-stop operations.
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Fig.8 Pressure difference in case #1 and case #3
3.2 Capacity comparison
With the components concentrations of inlet gasoline vapor were measured in advance, the cooling
capacity was calculated by the enthalpy difference method on the gasoline vapor side in Eq.[3].
[3]𝑄𝑔𝑣 = ∑3𝑗 = 1{(𝜌𝑔𝑣,𝑗 ‒ 1𝑉𝑔𝑣,𝑗 ‒ 1ℎ𝑔𝑣,𝑗 ‒ 1 ‒ 𝜌𝑔𝑣,𝑗𝑉𝑔𝑣,𝑗ℎ𝑔𝑣,𝑗) ‒ ∑
𝑘[(𝑚𝑔𝑣,𝑘,𝑗 ‒ 1 ‒ 𝑚𝑔𝑣,𝑘,𝑗)ℎ𝑔𝑙,𝑘]}The subscripts gv, gl and j represented gasoline vapor, gasoline liquid (condensate) and the state of
after jth cooling stage (0 stage meant inlet state), respectively. represented the inlet volumetric 𝑉𝑔𝑣,0
flow rate and was measured by inlet flowmeter; represented the mass content of k component 𝑚𝑔𝑣,𝑘,𝑗
in gasoline vapor after jth cooling stage; represented the specific enthalpy of k component in ℎ𝑔𝑙,𝑘
gasoline liquid. , , and were calculated from Table. A.1.𝜌𝑔𝑣,𝑗 ℎ𝑔𝑣,𝑗 𝑉𝑔𝑣,𝑗 𝑚𝑔𝑣,𝑘,𝑗
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Fig.9 Cooling capacity comparison at one period
Fig.9 illustrated the cooling capacity values in different cases at one period. The corresponding
temperature differences (ΔT) between the evaporator (T01-T06) and gasoline vapor channel (T11,
T12, T21, T22 and T30), as well as the pressure drop (ΔP) were also recorded. In case #1, the cooling
capacity was increased by 28.2% than case #3 without refrigerant evacuation. There are two main
reasons for the capacity reduction of case #3:
(1) Frost accumulation cause by defrosting failure (section 3.1) which decreased the heat transfer.
It’s also checked by the temperature difference shown in Fig. 9, the ΔT of case #3 was 13°C on
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average, while ΔT of case #1 was only 5°C.
(2) Residual refrigerant in defrosting evaporator was difficult to pump out only by the way of hot
gas heating, which resulted in insufficiency of cycling refrigerant. Observing the region at initial
relative time in Fig.9, larger ΔT and lower ΔP was measured. It indicated high superheat and lack of
cycling refrigerant.
The performance characteristics of case #2 was approaching to case #1. The only difference was the
capacity decline in DSP, as the heat sink of defrosting channel was not enough to cool down the
gasoline vapor after long switching delay. In case #4 without delay switching, the cooling capacity
was worse than case #1 and 2#, which represented the delay switching cases. It also declined
dramatically during channel switching, because the working evaporator did not reach steady state
when just switching.
3.3 Stability and defrost efficiency analysis
Cooling duty ratio (CDR) and cooling reduction ratio (CRR) were defined artificially as Eq.[4] and
[5]. These two parameters were used to indicate the fluctuating period and range during defrosting.
The lower CDR value was, the larger fluctuating period existed. While the higher CRR value meant
the larger fluctuating range.
[4] [5] 𝐶𝐷𝑅 = t𝑠𝑡𝑒𝑎𝑑𝑦 t𝑝𝑒𝑟𝑖𝑜𝑑 𝐶R𝑅 = (𝑄𝑠𝑡𝑒𝑎𝑑𝑦 ‒ 𝑄𝑚𝑖𝑛) 𝑄𝑠𝑡𝑒𝑎𝑑𝑦
and represented the times of the steady state and the whole operation period, t𝑠𝑡𝑒𝑎𝑑𝑦 t𝑝𝑒𝑟𝑖𝑜𝑑
respectively. represented the minimum cooling performance value during defrosting. 𝑄𝑚𝑖𝑛
Furthermore, these parameters were identified in Fig. 10.
The cooling capacity in case #1 and case #3 was shown in Fig.10; as mentioned in section 3.1, the
capacity measurement of case #3 was terminated in near 24 hours because of the high pressure drop.
The values of CDR and CRR were summarized in Table. 2, which also compared to published heat
pump defrosting studies. The fluctuating period and range in case #1 of this study were all smaller
than the references’ studies. The situation of case #2 was similar to case #1 except about 10%
reduction of both steady state time and minimum cooling performance values when channel switching.
In case #3 without refrigerant evacuation, the value of CDR was much lower. That meant fluctuating
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period prolonged significantly, because the cycling refrigerant was insufficient for much time.
Furthermore, the residual refrigerant inside the defrosting evaporator consumed the heat energy of
hot gas, which increased the defrosting load. In case #4 without switching delays, the cooling capacity
reduction became larger. Because the cooling potential of frost was wasted, and there is not enough
time for the working evaporator to pre-cool. It was drawn that refrigerant evacuation and switching
delay are effective methods for improving the energy conversion efficiency. Based on this, the system
performed successfully in the non-stop cooling application with little fluctuations and reductions.
Fig.10 Cooling capacity in case #1 and case #3
Defrost efficiency [58-59] was defined as the ratio of the minimum energy required to melt the frost
to the energy applied to actually melt the frost. Base on Hoffenbecker [59], the distribution of defrost
energy was tracked in Appendix B. Since the defrost energy contribution of gasoline vapor was Q𝑔𝑣
the necessary heat source in cooling process, it was treated as no power consumption to defrost and
not including in the applied energy term of defrost efficiency.
[6]η𝑑 =Q𝑚𝑒𝑙𝑡
Qℎ𝑔 + Q𝑒𝑙
Cycle-average defrost efficiency and energy contribution of case #1 and case #3 in different stages
was evaluated in Fig.11. It’s found that most heat load appeared at the 2nd stage, because of large
defrost temperature span and melting over half the frost at this stage. The phenomenon of frost
accumulating at 2nd stage was also captured in Fig.9: the fast increment of ΔT at 2nd stage with the
ascending of ΔP. On the other hand, gasoline vapor energy made an important positive effect. As a
free heating factor, it even led to over 100% defrost efficiency at the 1st stage. Heat exchanger term
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dominated the parasitic heat load, because the heavy shell-tube should met the requirements of
Chinese standard GB150 in such inflammable and explosive condition. Another inefficient factor was
the refrigerant heating, which also played major role in case #3, and was eliminated by refrigerant
evacuating in case #1. The values of overall defrost efficiency were also summarized in Table. 2.
Compared to other published studies, the heavy evaporator structure and low temperature made this
system a lower defrost efficiency, but the free heating of gasoline vapor compensated the short slab.
Fig.11 Defrost efficiency and energy consumption in case #1 and case #3
Table 2 - Performance and defrost efficiency comparison of some defrosting systems
References Key descriptions CDR CRR η𝑑
Case #1 Switchable dual evaporator; non-stop defrosting; switching delay;
refrigerant evacuation; cascade; evaporating temperature of 0/-35/-70°C75% 10% 55.4%
Case #2 Over delay switching, evaporating temperature of 0/-35/-70°C 70% 20% 53.0%
Case #3 No refrigerant evacuation, evaporating temperature of 0/-35/-70°C 21% 27% 40.6%
Case #4 No switching delay, evaporating temperature of 0/-35/-70°C 55% 47% 46.6%
Kim et al. [34], Fig.7 Periodic reverse cycle defrosting - 100% -
Kim et al. [34], Fig.8 Dual hot gas bypass defrosting - 53.8% -
Kim et al. [34], Fig.9 Dual hot gas bypass with inverter heater - 57.1% -
Zhang et al. [62], Fig.12Reverse cycle defrosting with using heat energy dissipated by the
compressor25% 35.3% -
Jang et al. [17], Fig.9 Periodic reverse cycle defrosting 50% 100% -
Jang et al. [17], Fig.9 Dual spray hot gas defrosting 75% 57.1% -
Qu et al. [63], Fig.11Cascade air source heat pump, thermal energy storage based reverse
cycle defrosting- 73% -
Cho et al. [33], Fig.4 Multi-evaporator refrigeration; on-off cycle 60% 100% -
Cho et al. [33], Fig.7 Multi-evaporator refrigeration; hot gas bypass - 25.5% -
Dong [28] Section 4 Reverse cycle defrost experiment, evaporating temperature of 0°C - - 60.1%
Melo et al. [30] Table 1 Defrost experiments of electric heaters, freezer temperature of -23°C 48.0%
Hoffenbecker [64] Table 3 Coil defrost model of hot gas, freezer temperature of -23°C - - 43.7%
Song [65] Table 6 Hot gas defrost experiment, evaporating temperature of 0°C - - 47.1%/58.8%
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From pressure control standpoint in case #1, Fig.12 illustrated the gauge pressure values at different
positions of low pressure side (P01-P04). Each curves appeared periodic spike and sustained
fluctuation. The pressure spike was synchronizing with the refrigerant switching to the warm
evaporator after defrosting (step 2.1 of section 2.2); while the sustained fluctuation was synchronizing
with the hot gas spraying solenoid valve action. Violent pressure spike would result in the mechanical
damage to the compressor and sustained pressure fluctuation would affect the evaporating pressure
of the simultaneous cooling. The pressure spike ranges were less than 0.2 MPa and safe for the
compressor. The sustained fluctuations were less than 0.05 MPa, the corresponding effects on the
evaporating temperature of the cooling side were less than ±2.5°C, which was important for the
stable operation.
Fig.12 Low pressure in different positions for case #1
3.4 Sensitivity of the switching delay
Sensitivity studies of different switching timing were carried out by analyzing the gasoline vapor
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temperatures (T30) in Fig.13. From the perspectives of the cooling temperature at steady state, case
#1 and case #2 performed at the same level, and both better than case 4. As a failure case, the
temperatures of case #3 cannot maintain a stable level and ascend dramatically with the existing
switching strategy. Then we can summarize that inappropriate switching strategy will lead to
performance reduction and even failure operation. The minimum temperature fluctuations were
located in case #1, which were within 10°C. For case #2 of over delay switching, the fluctuations
were about 5oC higher than case #1. The temperature fluctuations in case #4 appeared significantly
higher than other normal cases, which was up to about 45oC. Therefore, the timing of switching was
essential for the temperature fluctuations. If under-delay switching, the precooling at working
evaporators before switched was not enough, the working evaporators were not ready for cooling just
after defrosting. If over-delay switching, heavy components in the gasoline vapor can lead to frost at
2nd stage and even at 3rd stage when they flowed through the defrosting channel, which resulted in
redundant defrosting and energy consumptions. The optimal switching delay was about 20 minutes
for this device. At this time, the gasoline vapor temperature for defrosting at the outflow of 1st stage
began to increase, and it’s supposed that the frosts at the defrosting evaporator of 1st stage have just
been melted. With this prerequisite, the frost cooling potential was best recovered by gasoline vapor.
Fig.13 Comparison of the gasoline vapor cooling temperatures
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4. Conclusions
A novel dual channel cooling system for GVR was developed in this study. Compared to
conventional VOCs condensation method, the switchable one-work-one-standby evaporators and its
control strategies were adaptive to non-stop running and high energy efficiency. According to
industrial testing results, the following conclusions can be summarized:
(1) Refrigerant evacuation ensured the amount of cycling refrigerant and therefore improved the
cooling capacity by 28.2%. Otherwise, it might fail to defrost completely after long time running.
(2) No switching delay didn’t result in notable capacity reduction of steady state, but affected its
fluctuations.
(3) Switching delay can bring free energy of gasoline vapor heating, refrigerant evacuation can
eliminate the parasitic load of refrigerant heating, which were all profitable to defrost efficiency.
(4) The best case of defrost in this study was 10% the performance reduction ratio, 75% the duty
ratio, and 55.4% the defrost efficiency, which was competitive compared to the references’
studies.
(5) Appropriate switching delay can optimize the performance and fluctuation by recovery of the
frost cooling potential, which was an effective method of energy saving. The best timing of
switching delay was about 20 minutes for this device.
The main achievement of this study was the discovery of the dual switchable channel method to
supply continuous long-running cooling in the multi-stage and cascade cycles. By adjusting the
defrosting control strategy, solved the problem of frost interruption, which has been the greatest
shortcoming of non-stop cooling process with heavy components. But the channel switching timing
and criteria still needs to be optimized. We suggest developing system-level modeling integrated with
a detailed frost growth model for future exploration, which is beneficial to energy performance
analysis and model-based control design.
AcknowledgementThis work was supported by Guangdong Shenling Environmental System Co., Ltd. (China)
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Appendix A
Table A.1 - Data sheet of gasoline vapor condensation process calculation results
T V h ρ Components mass flow rate of gasoline vapor (kg/hr)
oCm3· h-
1
kJ·kg-
1
kg· m-
3 C2H6 C3H6 C4H102-C4H10
1-C4H8
2-C4H8
C5H12 C6H14 Air etc
25 600.0 -
1020.7 1.859 3.042 4.679 55.460 79.064 37.305 28.918 286.904 175.864 444.342
5 499.9 -854.1 1.824 2.977 4.586 51.971 73.122 33.479 26.162 206.029 71.333 442.135
-28 347.6 -402.1 1.688 2.630 4.074 34.003 43.594 16.452 13.419 37.132 3.017 432.479
-55 279.8 -224.2 1.734 2.246 3.535 18.807 20.724 5.778 4.991 3.259 0.042 425.717
-56 277.5 -218.9 1.736 2.216 3.494 18.006 19.668 5.405 4.678 2.942 0.037 425.437
-57 275.3 -213.8 1.739 2.185 3.451 17.213 18.638 5.049 4.380 2.656 0.033 425.166
-58 273.1 -208.9 1.742 2.153 3.407 16.431 17.637 4.712 4.095 2.398 0.029 424.905
-59 270.9 -204.1 1.745 2.120 3.361 15.662 16.667 4.392 3.824 2.167 0.026 424.655
-60 268.8 -199.4 1.749 2.085 3.313 14.908 15.729 4.090 3.567 1.958 0.023 424.416
-61 266.7 -194.9 1.752 2.050 3.264 14.171 14.825 3.805 3.324 1.769 0.020 424.188
-62 266.7 -194.9 1.752 2.050 3.264 14.171 14.825 3.805 3.324 1.769 0.020 424.188
-63 262.7 -186.3 1.760 1.975 3.160 12.753 13.120 3.284 2.878 1.446 0.016 423.763
-64 260.8 -182.3 1.764 1.937 3.106 12.075 12.321 3.047 2.674 1.308 0.014 423.567
-65 258.8 -178.5 1.769 1.897 3.050 11.418 11.557 2.825 2.482 1.182 0.013 423.380
-66 257.0 -174.8 1.773 1.856 2.992 10.783 10.829 2.617 2.302 1.069 0.011 423.204
-67 255.1 -171.3 1.778 1.815 2.934 10.171 10.135 2.422 2.133 0.967 0.010 423.037
-68 253.3 -167.9 1.783 1.772 2.873 9.582 9.476 2.240 1.975 0.874 0.009 422.880
-69 251.5 -164.7 1.789 1.729 2.811 9.016 8.850 2.070 1.827 0.790 0.008 422.731
-70 249.7 -161.7 1.794 1.685 2.748 8.474 8.256 1.912 1.689 0.714 0.007 422.592
-71 248.0 -158.9 1.800 1.640 2.683 7.955 7.695 1.764 1.560 0.645 0.006 422.460
-72 246.3 -156.2 1.806 1.595 2.618 7.459 7.164 1.627 1.439 0.583 0.006 422.336
-73 244.7 -153.7 1.812 1.549 2.551 6.986 6.663 1.499 1.327 0.526 0.005 422.220
-74 243.0 -151.3 1.818 1.502 2.483 6.535 6.191 1.380 1.223 0.475 0.004 422.110
-75 241.4 -149.1 1.825 1.455 2.414 6.106 5.747 1.269 1.125 0.429 0.004 422.008
Appendix B
The defrosting energy flows was quantified as follow:
(1) - the amount of energy supplied to defrosting by the hot gas, which was calculated Qℎ𝑔
according to:
[B.1]Qℎ𝑔 = ∫𝑡𝑒𝑡𝑖
𝑚ℎ𝑔(ℎℎ𝑔,𝑑𝑖𝑠 ‒ ℎℎ𝑔,𝑒𝑣𝑎𝑝,o𝑢𝑡) ∙ 𝑑𝑡𝑠𝑝𝑟𝑎𝑦
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Where was the hot gas flowrate, was hot gas enthalpy at discharge point; 𝑚ℎ𝑔 ℎ𝑑𝑖𝑠 ℎℎ𝑔,𝑒𝑣𝑎𝑝,o𝑢𝑡
was hot gas enthalpy at the outlet of defrosting evaporator, was the hot gas spraying time, 𝑡𝑠𝑝𝑟𝑎𝑦
which is captured by the pulsing signal of solenoid valve; and were the timers which 𝑡𝑒 𝑡𝑖
initiated and terminated the defrost period, respectively. In addition, was determined using 𝑚ℎ𝑔
the collected experimental data and the Fusheng Selection Software from the compressor
manufacturer.
(2) - the amount of energy supplied to defrosting by the electric heater, which was calculated Q𝑒𝑙
according to:
[B.2]Q𝑒𝑙 = ∫𝑡𝑒𝑡𝑖
𝑃𝑒𝑙𝑑𝑡𝑒𝑙
Where and were the power and heating time of the electric heater, the was held 𝑃𝑒𝑙 𝑡𝑒𝑙 𝑃𝑒𝑙
constant at 1.0 kw per evaporator.
(3) - the amount of energy supplied to defrosting by the gasoline vapor during the step 1.3 refer Q𝑔𝑣
to section 2.2, which was derived by:
[B.3]Q𝑔𝑣 = ∫𝑡𝑒𝑡𝑖
𝑄𝑔𝑣 ∙ 𝑑𝑡𝑔𝑣
Where was calculated by Eq[4] using the local temperatures of defrosting channel. Since 𝑄𝑔𝑣
gasoline vapor duct is small-size radial and circular, its radial temperature distribution is
assumed as uniform. is the time of gasoline vapor flowing through the defrosting channel. 𝑡𝑔𝑣
To prevent the confusion of melt and condensate drainage, dry nitrogen was chosen instead of
gasoline vapor when estimating the frost mass. Therefore, was certain underestimated as Q𝑔𝑣
neglecting the latent heat of gasoline vapor condensation.
(4) - the amount of energy it took for the mass of frost to melting point and change phase Q𝑚𝑒𝑙𝑡
from solid to liquid. Since the gasoline vapor was from refined oil, the frost component was
assumed as only water. Then, we had:
[B.4]Q𝑚𝑒𝑙𝑡 = ∑3𝑗 = 1𝑚𝑓,𝑗 ∙ [∆ℎ𝑠𝑓 + 𝑐𝑝,𝑓,𝑠(𝑇𝑚𝑒𝑙𝑡 ‒ 𝑇𝑓,𝑖,𝑗)]
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Where was the enthalpy change from solid to liquid, was the specific heat of frost in ∆ℎ𝑠𝑓 𝑐𝑝,𝑓,𝑠
solid phase, was the melting temperature, and were the frost temperature and 𝑇𝑚𝑒𝑙𝑡 𝑇𝑓,𝑖 𝑚𝑓,𝑖
mass at jth stage. was evaluated by the liquid level of gasoline storage tank, and was 𝑚𝑓,𝑗
estimated by:
[B.5]𝑚𝑓,𝑗 = 𝑚𝑓(𝐿𝑗) ‒ 𝑚𝑓(𝐿𝑗 ‒ 1), 𝑗 = 1,2,3
Where was the liquid level of gasoline storage tank when the jth stage condensate began to 𝐿𝑖
drain and was the liquid level before defrosting, as it’s observed that the melt drainage of 𝐿0
different stage evaporator occurred in different time.
(5) - the energy provided beyond that needed to heat given frost past the melting point. Q𝑒𝑥𝑐𝑒𝑠𝑠
[B.6]Q𝑒𝑥𝑐𝑒𝑠𝑠 = ∑3𝑗 = 1𝑐𝑝,𝑓,𝑙 ∙ 𝑚𝑓,𝑗 ∙ (𝑇𝑒 ‒ 𝑇𝑚𝑒𝑙𝑡)
Where was the specific heat of frost in liquid phase, was the predetermined evaporator 𝑐𝑝,𝑓,𝑙 𝑇𝑒
temperature when the defrost period was terminated. was overestimated by using Q𝑒𝑥𝑐𝑒𝑠𝑠 𝑇𝑒
instead of the drainage temperature.
(6) - the energy stored in the refrigerant inside evaporator during a defrost period, which was Q𝑟𝑒𝑓
calculated according to:
[B.7]Q𝑟𝑒𝑓 = ∑3𝑗 = 1m𝑟𝑒𝑓,i,𝑗(ℎ𝑟𝑒𝑓,𝑒,𝑗 ‒ ℎ𝑟𝑒𝑓,𝑖,𝑗)
Where and are the enthalpies of the refrigerant at jth stage evaporator when the ℎ𝑟𝑒𝑓,𝑖,𝑗 ℎ𝑟𝑒𝑓,𝑒,𝑗
defrosting initiating and terminating; is the mass of the refrigerant inside jth stage 𝑚𝑟𝑒𝑓,𝑖,𝑗
evaporator when the defrosting initiating, which was evaluated by the refrigerant recovery
machine.
(7) - the total thermal energy stored within the heat exchanger shell and tube after a defrost Qℎ𝑥
period was complete, which was calculated according to:
[B.8]Qℎ𝑥 = ∑3𝑗 = 1𝑚ℎ𝑥,𝑗 ∙ 𝑐
𝑝,ℎ𝑥(𝑇𝑒 ‒ 𝑇𝑓,𝑖,𝑗)
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Where was the mass of evaporator at jth stage and was the specific heat of evaporator, 𝑚ℎ𝑥,𝑗 𝑐𝑝,ℎ𝑥
as all shells and tubes of evaporators were all made of stainless steel.
(8) - the sensible energy convected to the ambient, also the latent energy caused by Q𝑙𝑜𝑠𝑠
sublimation and re-evaporation from the tube surface, which was calculated by energy
conservation:
[B.9]Q𝑙𝑜𝑠𝑠 = Qℎ𝑔 + Q𝑒𝑙 + Q𝑔𝑣 ‒ Q𝑚𝑒𝑙𝑡 ‒ Q𝑒𝑥𝑐𝑒𝑠𝑠 ‒ Qℎ𝑥
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ACCEPTED MANUSCRIPT
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Highlights
1. A novel switchable dual channel cooling system was proposed.
2. Combined defrosting method for low temperature shell-tube heat exchanger was presented.
3. Uninterrupted and lower reduction performance was validated by industrial application.
4. Sensitivity characteristics of defrosting control strategy were analyzed.