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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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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.